ENHANCED COMPRESSOR DISTORTION
TOLERANCE USING ASYMMETRIC INLET
GUIDE VANE STAGGER
by
Andreas Schulmeyer
B.S., University of Illinois, (1987)
SUMMITED IN PARTIAL FULFILLMENT
OF THE REQUIREMENTS FOR THE DEGREE OF
MASTER OF SCIENCE
IN AERONAUTICS AND ASTRONAUTICS
at the
MASSACHUSETTS INSTITUTE OF TECHNOLOGY
August 1989
Massachusetts Institute of Technology
c
...
Signiture of Author
Department of Aeronautics and Astrondhtics, August, 1989
Certified by
Accepted by
Professor Edward M. Greitzer
Thesis Supervisor
Department of Aeronautics and Astronautics
.I
~-
E-~
Professor Harold Y. Wachman
Graduate Committee
ntC~,nt
t 2 91989
UORAM
Aero
VITHDRAWN
M.I.T.
LIBRARlfiS
Enhanced Compressor Distortion Tolerance'Using
Asymmetric Inlet Guide Vane Stagger
by
Andreas Schulmeyer
Submitted to the Department of Aeronautics and Astronauticýs
on August 28, 1989 in partial fulfillment c:f the
requirements for the Degree of Master of Science
in Aeronautics and Astronautics
ABSTRACT
This w6rk is experimental verification of a method
tto
increase axial compressor distortion tollerance.
The method
uses a restagger of the inlet guide vanes to achieve the
desired uniform inlet velocity profile.
In order to achieve
this the vanes are opened in the low total pressure region,
increasing mass flow and pressure rise in this section,
and
closing
the vanes
in the high total pressure region to
reduce the mass -flow and pressure rise in this section of
the compressor.
The
theoretical calculations showed that a ten degree
restagger
of the guide vanes would
achieve the desired
result for
a one dynamic head total
pressure
distortion.
This setup was then tested on the MIT low-speed single stage
compressor. The test data showed the improved performance of
the restaggered compressor.
The non--uniformity in the inlet
velocity is reduced by 50%, while presssure rise at stall is
5.3% higher
than without the restagger and the stall
flow
coefficient is 2.7% higher.
Thesis Supervisor: Edward M. Greitzer
Title: Professor of Aer'onaut-i cs and Astronautics
ACKNOWLEDGEMENTS
I
would
like
Professor E.M.
the
to
Greitzer,
unexpected
my
offer
mysteries
thanks
to
for his support in
that
the
advisor,
my
solving many of
experimental
results
provided.
Thanks to Gwo-Tung Chen for his theoretical work
set the basis for this experimental
the
which
investigation as well as
patient introduction he gave me to the methods of
this
scheme.
I
also
Dubrowski,
Roy Andrew,
experimental
I
and
appreciate
all the help
by
Viktor
Jim Nash and Jerry Guenette with the
setup.
would also like to thank all
at
provided
Ashdown
for
making
life
my friends at
at
MIT
so
the
much
GTL
more
enjoyable.
Most
of all
I would like to thank my parents,
and Helga Schulmeyer,
for all their love and
Gerhard
encouragement
that have made it possible for me to be where I am.
This
work
was
supported by the Air Force
Office
Scientific Research under contract AFOSR -87- 0398E.
of
TABLE OF CONTENTS
Abstract
2
Acknowledgements
3
List of Tables
5
List of Figures
6
List of Symbols
9
Chapter 1
Introduction
Chapter 2 Theoretical Model and Calculations
2.1)
2.2)
2.3)
2.4)
Introduction
Compressor Characteristic Calculations
Distortion Theory Summary
Restaggering Scheme Calculations
Chapter 3 Experimental Setup
3.1)
3.2)
3.3)
3.4)
3.5)
3.6)
Introduction
Blading Modifications
Blade Row Spacing
Distortion Screen Design and Placement
Instrumentation Layout
Experimental Procedure and Data Aquisition
Chapter 4 Experimental Results
4.1)
4.2)
4.3)
4.4)
4.5)
Compressor Characteristics
Restaggering Scheme Calculations
Baseline Compressor Behaviour
Restaggered Compressor Behaviour
Limitations
Chapter 5 Conclusions and Recomendations for Future Work
5.1) Conclusions
5.2) Recomendations for Future Work
References
Appendix
10
13
13
13
16
18
21
21
22
23
23
24
25
27
27
29
30
31
33
35
35
35
37
Actuator Disc Modeling
38
Tables
42
Figures
44
LIST OF TABLES
Tab I e
2.1
Compressor Geometry
42
3.1
Distortion Screen Total Pressure Loss Coefficient
43
LIST OF FIGURES
Figure
1.1
Typical Allocation of Required Surge Margin
44
1.2
Parallel Compressor Description of Response to
Circumferential Distortion
45
1.3
Effect of Non-Axisymmetric Vane Stagger on
Compressor Response to Total Pressure Distortion
46
2.1
Three Stage Compressor
Restaggering Schemes
47
2.2
Empirical Curve Relating Loss Coefficient and
Diffusion Factor
48
2.3
Multiplying Factor on Loss Coefficient as a Funtion
of Diffusion Factor
48
2.4
Deviation 'Adder' Applied to Carter's Rule as a
Function of Diffusion Factor
49
2.5
Calculated and Experimental Data for 4 Low Speed
Three-Stage Compressors
50
2.6
Compressor Speedline for Control Modelling
51
2.7
Experimental Speedline
52
2.8
Compressor Speedline with a 10 Degree Opening
and Closing of IGVs and Stators
53
2.9
Inlet Total Pressure Distortion Profile
54
2.10 Compressor Response to a 1.0 Dynamic Head
Distortion with a 5,10, 15 Degree Restagger of
the IGVs and Stators Near Stall
55
2.11 Compressor Response to a 1.0 Dynamic Head
Distortion with a 5,10,15 Degree Restagger of
the IGVs and Stators at Design Flow
55
2.12 Compressor Response to a 1.0( Dynamic Head
Distortion with a 8,12,16 Degree Graduated
Restagger of the IGVs and Stators Near Stall
56
2.13 Compressor Response to a 1.0 Dynamic Head
Distortion with a 8,12,16 Degree Graduated
Restagger of the IGVs and Stators at Design Flow
56
Response to Three
2.14 Compressor Response to a 1.0 Dynamic Head
Distortion with a Shifted Graduated Restagger
of the IGVs and Stators Near Stall
57
2.15 Compressor Response to a 1.0 Dynamic Head
Distortion with a Shifted Graduated Restagger
of the IGVs and Stators at Design
57
2.16 Compressor Response to a 1.0 Dynamic Head
Distortion with a 20,30,40 Degree IGV Restagger
Near Stall
58
2.17 Compressor Response to a 1.0 Dynamic Head
Distortion with a 11,22,33 Degree IGV Restagger
at Design
58
3.1
Axial Velocity Perturbation Introduced by a
10 Degree IGV Restagger vs. Blade Row Spacing
59
3.2
Circumferential Inlet Total Pressure Profile
with Inlet Distortion
60
3.3
Pressure Instrumentation Layout
61
3.4
Velocity Measured by Pitot Tube vs. Velocity
Measured by Total Pressure Kiel Probes and
Static Pressure Taps
62
3.5
Scannivalve 1 Pressure Transducer Calibration
Curve
63
3.6
Scannivalve 2 Pressure Transducer Calibration
Curve
63
3.7
Flow Coefficient from Averaged Pressure Readings
vs Coefficient from Averaged Local Flow
64
4.1
Compressor Speedline for 7.5, -7.5 and 22.5
Degree IGV Stagger
65
4.2
Compressor Speedline for 0.0 and 27.5 Degree
IGV Stagger
65
4.3
Compressor Speedline for 7.5,
Degree IGV Stagger
66
4.4
Inlet Static Pressure Profile at Design Flow
67
4.5
Inlet Velocity Profile Near Stall
67
4.6
Exit Static Pressure Profile at Design Flow
68
4.7
Compressor Speedline Curvefit for Restagger
Calculation
69
15.0 and 22.5
4.8
Change in Pressure Rise due to Change in
Stagger Angle vs Flow Coefficient
4.9
Compressor Characteristic Slope vs Flow
Coefficient
IGV
70
70
4.10 Calculated Compressor Response to a 1.0
Dynamic Head Distortion with 5,10,15 Degree
IGV Restagger Near Stall
71
4.11 Calculated Compressor Response tp a 1.0
Dynamic Head Distortion with 5,10,15 Degree
IGV Restagger at Design Flow
72
4.12 Calculated Compressor Response to a 1.0
Dynamic Head Distortion with 5,10,15 Degree
Square Wave Restagger at Design Flow
72
4.13 Compressor Speedline for 15 Degree IGV Stagger
with and without Inlet Total Pressure Distortion
73
4.14 Inlet Velocity Profile for Baseline Compressor
with Inlet Total Pressure Distortion Near Stall
74
4.15 Inlet Velocity Proflie for Baseline Compressor
with Inlet Total Pressure Distortion at Design Flow
74
4.16 Velocity Non-Uniformities for Baseline Compressor
'75
with 1.0 Dynamic Head Inlet Total Pressure Distortion
4.17 Magnitude of First, Second and Third Fourier
Coefficients of Velocity Non-Uniformity f'or
Baseline Compressor with Inlet Distortion
75
4.18 Magnitude of Axial Velocity Perturbation for an
Isolated Rotor with a 10% Inlet Velocity
Non-Uniformity
76
4.19 Restaggered and Baseline Compressor Speedline
with Inlet Total Pressure Distortion
77
4.20 Inlet Static Pressure Profile for Baseline and
Restaggered Compressor Near Stall
78
4.21 Inlet Velocity Profile for Compressor with Inlet
Total Pressure Distortion Near Stall
79
4.22 Inlet Velocity Profile for Compressor with Inlet
Total Pressure Distortion at Design Flow
79
4.23 Velocity Non-Uniformities for Restaggered
Compressor with 1.0 Dynamic Head Inlet Total
Pressure Distortion
80
4.24 Magnitude of First, Second and Third Fourier
Coefficients of Velocity Non-Uniformity for
Baseline and Restaggered Compressor
80
4.25 Phase of First, Second and Third Fourier
Coefficients of Velocity Non-Uniformity for
Baseline and REstaggered Compressor
81
4.26 Exit Static Pressure Profile for Comressor with
Inlet Total Pressure Distortionand IGV Restagger
at Design Flow
82
List of Symbols
C
Velocity
Ci
Blade chord
cp
Pressure rise coefficient across blade row
P
Non-dimensional pressure
R
Mean radius
r
Blade stagger angle
a
Absolute flow angle
7
Blade stagger angle
Flow coefficient
P
Density
0
Circumferential angle
P
Compressor pressure rise
Subscipts
e
Exit
in
Inlet
t
Total
x
Axial
y
Circumferential
Chapter 1 Introduction
distortion
Inlet
affecting the stall
margin in
important
an
factor
compressors.
modern axial
one half of the design stall
to
is
tolerance
may
margin (Figure 1.1)
Up
be
allocated to account for the effect of inlet distortion [13.
This
margin
flow
due
such phenomena as crosswind,
to
induced
attack
interaction
in
allows for the non-uniformities in
in
inlet seperation
supersonic
inlets.
or
the
high
inlet
angle
shock-boundary
of
layer
In addition to decreases
pressure rise and efficiency of the compressor,
effects of inlet distortion are rotating
stall,
possible
combustion
flame out or engine overheating.
these
For
axial
reasons the effect of inlet
compressors
years.
Methods
E2,3,43
has received much attention in
have been developed to predict
the
succeptability
performance
The standard methods to reduce
to distortion are to increase the
with
engine
tolerance
non-uniformities by modifying the engine inlets,
lower
on
past
and stability E5,63 of a compressor operating
an inlet distortion.
to
distortion
or
to
the operating point to increase the stall margin.
With
the capabilities of new electronic controls other
to
methods
increase
distortion
tolerance
have
become
available. The theoretical capability to control the stagger
setting of each individual inlet guide vane (IGV)
blade
[73
or
stator
allows the compressor to be divided into two
or
more 'parallel compressors', with an asymmetric restaggering
patern chosen to match the inlet distortion E83.
By opening
the
vanes
the
in
low
pressure
total
-flow
the
region
coefficient
is
increased and the operating point moves away
from stall.
In
the high total pressure region the vanes are
closed and the flow coefficient is
this
region
is
reduced,
operating far from stall
however,
the
because
reduction
in
flow coefficient does not pose a problem.
The
behaviour
can
manner
parallel
be
of
a compressor
restaggered
shown qualitatively in
compressor
the
in
response
to a inlet total pressure
this
of
a
distortion.
The low total pressure zone and the high total pressure zone
operating
inlet
points
are determined by the
distortion
and
the
slope
characteristic
(Figure 1.2).
coefficient in
the two regions,
closer to stall.
operating
reduces
this
magnitude
the
of
velocity non-uniformity by shifting
pressure
increase
in
of
the
the restaggering is
stall
the
compressor
a
more
low
the
region speedline to the left (Figure
around
zone
A non-axisymmetric vane stagger
total
coefficient
compressor
with the low pressure
pressure region speedline to the right and
result
the
This leads to a different flow
total
The
of
high
1.3).
uniform
circumference
margin gained by moving the
flow
and
low
an
total
pressure operating point further from stall.
A further benefit of using this method is
averaged
pressure
rise
is
closing
all vanes uniformly.
overall
engine
deliver
a
performance
not reduced
as
that the mass
much
as
A second benefit in terms
is
that
the
when
of
compressor
can
uniform flow to the combustor thus reducing
the
of
cd
I ikelyhoo
hot spots that
developing
damage
cat
the
turbine3.
ob jcect
Th e
demonstrate
this
type
toleraince
distortion
is
Turbine
...aboratory.
in
this
of"
thesis
is
scheme
for
of arn existi-ng
thesis.
The
results
t
of
f
experimenta I ly
to
impjroving
ccompressc-r.
single stage compressor
vehicle
described
the
this
of
the
at
inlet
The
test
MIT
Gas
exper iment
are
the
Chapter 2 Theoretical Model and Calculations
2.1)
Introduction
theoretical calculations performed as part of
The
distortion
control are the application to a specific single
stage compressor of the method described by Chen et al.
axial compressors.
the
The work in
characteristics
implementing
procedure
data.
control
a
consisted of calculating
E83
three
stage
compressor
and
this
compressor.
The
sceme on
lead to speedlines in
The
control
the
for
for
used to calculate the off-design performance
shown to
the
agreement with
was
experimental
computations also indicated that the
distortion
is capable of reducing the velocity defect due to a
distortion of greater than two dynamic heads by a factor
of
two (Figure 2.1).
This
compressor
chapter
deals
with
characteristic,
a
the
prediction
summary
of
the
of
the
distortion
computations and the application of the computations to
the
single stage compressor.
2.2) Compressor Characteristic Calculation
The
of
program used to calculate the off-design behaviour
the compressor was written by Chen E86
proposed by Raw and
Weir [93.
based on a
Losses are calculated -for the
blade rows by using a loss parameter based on the
factor
(Figure 2.2),
model
with a multiplication
diffusion
factor dependent
on blade Reynolds number (Figure 2.3) applied. Carter's rule
function
for deviation is modified by an "adder' which is a
of the diffusion factor of the blade rows (Figure 2.4).
The major assumptions of the analysis are:
- high
to
hub
tip ratio,
mean
so
line
velocity
triangles are used
- low Mach number so compressibility is neglected
- blades are represented by circular arcs
The
last assumption leads to deviations in
agreement
overestimated
neglecting
effects
experimental
with
by
neglecting
data,
real
as
the
reasonable
turning
-fluid effect
thickness turning is underestimated.
are of comparable magnitude [10)
order one,
is
and
by
These
two
for a solidity
of
the solidity of the rotor and stator rows in the
experimental compressor.
The
required
inputs to the program are the rotor
stator camber and stagger angles and the IGV leaving
The
and
angle.
compressor characteristic and the flow angles are
then
calculated for any desired range of flow coefficients.
For
the
calculations
experimentally
three
the
stage
compressors
agreement
between
obtained speedlines is
used
in
previous
calculated
arnd
close enough for
the
the
control calculations that followed (Figure 2.5).
The
showed
results for the single stage compressor,
a characteristic that continues to rise as the
coefficient deceases (Figure 2.6).
it
(or
was
the
however,
From these
calculations
not possible to predict the compressor stall
pressure rise at stall) with
any
flow
precision.
point
The
model does not perform satisfactorally
for the single
stage
compressor possibly because the correlations used to predict
optimized
are
losses
compressors.
multi-stage
for
The
factors are thus optimized for multi-stage compressors.
Even though the program calculated speedlines that
later obtained (Figure 2.7),
to
determine
the
amount
that
speedlines,
to the experimental
comparable
not
are
were
the calculated curves were used
location
and
of
restaggering
required to reduce the magnitude of the velocity distortion.
restagger ing
The
calculations
are
possible
method
uses
calculated
speedlines
changes
pressure rise due to a change in
in
as
the
with
the
the
relative
stagger and the
slope of "the characteristic as the relevant parameters.
the actual amount o-f restaggering required the
For
experimental
speedlines could then be used to repeat the calculations. To
the
find
optimum
restaggering
method
the
calculated
speedlines were then used with the point of inflection,
point
where the calculated speedline is the
'flattest',
the
as
the calculated 'stall' point.
The
program
speedlines.
table
The
The
then
used
to
calculate
various
final compressor geometry can be found
2.1 and the calculated characteristic in figure
stacking
scheme is
bracketed
closing
was
of the speedlines required for
shown in
by
this
in
2.6.
control
figure 2.8 where the nominal speedline is
the speedlines with a 10 degree
of the IGVs and stators.
and
From this family of curves
the parameters necessary for the restaggering
15
opening
calculations,
•a-o•• and
a''/a
, can be obtained.
2.3) Distortion Theory Summary
Chen
is
section
This
a
fo:'r
a summar'y of the theory developed
compressor
operating
with
a
steady
distortion. The pressure rise across a blade row is
by
inlet
given by
an axisymmetric part and an asymmetric component due to
the
fluid acceleration within the blade row. The fluid mechanics
compressor response to total pressure inlet distortion
for
is described in [53.
With
a linearized upstream and downstream flowfield the
compressor performance is given by
where
SCi
cos r4
is
the
parameter associated with
within the blade rows.
mean radius,
For
C,
respectively,
r,
fluid
acceleration
R are the chord, stagger and
of the rotor.
a small disturbance the downstream pressure
field
satisfies
VP2 P =O
(2.2)
and the boundary condition at the compressor exit plane is
aPc
P a•
ac, + cac(
"=O
16
a• =oy
The
Eq. (2.2)
results
linearized
and
(2.3)
obtained
from
perturbing
are
i
6=-
aq
++ r ST-a
a6o
ao--
and
a6 ,P
= 0' sec' a
aZ1,=0
variables
Expanding all
a6aI
in
(2.5)
Fourier series
(2.6)
+oo
=C
6i
a, einae-lniz
(2.7)
z
+oo
65i=
69 =
-oc
n
d enin#
(2.8)
dn ini
(2.9)
b neini
+-00
Sy =
E
L=-0O
and subsituting Eq.
(2.6),(2.8) and (2.9) into Eq. (2.5) we
have
zn
an=
Comb i n i ng
-
IIn
bn +
sec'
(2.4),(2.7),(2.8) ,(2.9)
S
bn
-
-
in
dn)
89n
and (2.10)
c )d
sec a1'
- nA +- seC
•in
i
se
p-
(2.10)
en
c•
Tc-o
(2.11)
stagger
The
distortion,
distribution
eliminates
that
the
velocity
i.e.
(2.12)
b =O
6 =0,
is
given by
2.4)
d,
7--
en
a
-1
a
(2.13)
, 02,_eC2
a+
secoC1 '3.7
Restaggering Scheme Calculations
The program that calculates the characteristics for
compressor
YT/D~
also
the
distortion and the implemented restagger.
derivatives
the compressor with an inlet total pressure
profile to minimize
the
These
two
are needed to t•d calculate the velocity profile
and various stagger profiles.
of
variables a.i/ay and
two
that determine the behaviour of the compressor with an
inlet
of
provides
the
The goal
is
distortion
to find a stagger
the velocity non-uniformity at the face
compressor.
For the calculations the
total
inlet
pressure distortion is a sine shaped defect over 180 degrees
of the annulus with a magnitude of one dynamic head
(Figure
2.9).
Based on the work done by Chen two schemes were used in
the first set of calculations.
both
total
IGVs
The first is a restagger
and stators identical amounts in phase with
pressure distortion.
resulting
axial
operating
near
The results can be seen in
velocity distribution for
stall
(Figure 2.10)
and
the
at
of
the
the
compressor
design
flow
For both of these the velocity defect due to
(Figure 2.11).
the
both
cases
the
slope of the characteristic
identical.
defect
at both flow coefficients is
are
velocity
the
also of the same
order
The
in almost identical behavior for both cases.
resulting
magnitude of the velocity distortion is
defect remaining.
It
any
non-uniformity
leads
restaggering
reduced by almost
a
is
a
for both flows although there
two
of
factor
and al~Y
The capability to correct
almost
in
as
distortion are of the same shape and magnitude,
is
not possible to reduce the velocity
further with this scheme,
to an overcorrection in
as a
larger
parts
of
the
the 15 degree restagger
which can be seen in
circumference,
still
of the blade rows.
to
led
This
distortion
still
sampling
of
can
possible
two,
velocity
the stators and IGVs
differing
in phase with the
to
reduce
distortion.
pressure
the resultant velocity profiles,
the stator restagger is
be
the
attempt
restaggering
by
amounts,
where
the
seen in
3/4th of the
*figures 2.12 and 2.13.
case
of the
restagger,
IGV
Here again it
to reduce the velocity distortion by a
but the velocity profile still
A
was
factor
of
shows the same behaviour
as the previous calculations.
The
next
restaggered
was
extended
step allowed for a phase shift
between
regions and the pressure distortion.
in
this case to allow for
between the restaggering
in
the
the two blade rows.
the repsonse to a stator restagger is
calculated
The model
phase
In
the
shift
this case
separately
This extension of the
from the response to a IGV restagger.
model
assumes that the velocity perturbations introduced by
restagger
the
vice
and
:Linearization
in
the
with
the
calculated
DI/0
that
the
and
aayan
perturbations
velocity
this case
or in
introduced by the total pressure distortion,
by the restagger of the other blade row.
as
possible
is
model assumes
even
constant
remain
This
versa.
velocity
the
both are
introduced by the IGVs if
perturbation
separately
the stators does not alter
of
The final result is
blade
then obtained by summing the effect of the individual
This method
row restaggerings.
a
introduces the phase shift as
new degree of freedom and should thus lead to an improved
velocity profile.
-figures 2.14 and 2.15 show a reduction of the
As
non-uniformity can be achived in
of
four
The magnitude
this manner.
velocity distortion can be reduced by a
the
or
more.
detracts
from
The
complexity of
factor
scheme,
this
its attractiveness so that
flow
further
of
however,
schemes
were tested.
The
most
This
completly
limited
method
to
reduce
the
velocity
was found to require a restaggering of the
distortion
only.
efficient
method
removing
by
the
(Figures 2.16 and 2.17) is
the
velocity
distortion,
amount of restagger of the
IGVs
capable
of
but
is
IGVs
it
possible
without separating the airfoils.
It was decided to use only the restaggering of the IGVs
as
the
scheme
restaggering
is
to
be
implemented.
The
magnitude
then dependent on the size of the
20
of
pressure
distortion,
stagger,
a/ay
the change in pressure rise due to the change in
,
and
the
without stalling the blades.
maximum
restaggering
possible
Chapter 3 Experimental Setup
3.1>) Introduction
The
GTL
single stage compressor was modified for
the
were
the
experimental
investigation.
restaggering
of
spacing,
The
major
the blade rows,
changes
the reduction
of
stage
the installation of a distortion screen and
instrumentation
and
modification of the
data
added
aquisition\
The compressor has a hub to tip ratio of 0.74 and
programs.
the design flow rate for the baseline blade setting is 0.59.
A
detailed
description of the geometry is given
in
table
2.1.
3.2) Blading Modifications
The IGVs used previously were long chord sheet metal of
high solidity.
These were not suitable for the restaggering
required by the distortion ex>periment,
to
as the long chord led
interference with the outer casing.
incidence
tolerance.
The
replacement
They also have
set of IGVs
had
shorter chord thus allowing for the required movement,
although
incidence
they
are
tolerance
also
sheet
metal,
somewhat
as they are thicker,
with
a
low
a
and,
better
rounded
leading edge.
The
new
blades
were
1/16th
inch
shorter
than
the
previous set leading to a clearance around the centerbody of
the compressor.
As the centerbody was now supported only at
the
by six struts three of
front
end
22
the
46
IGVs
were
replaced
bolts
by
to
the
locate
centerbody
the
and
eliminate vibration.
3.3) Blade Row Spacing
The compressor had been set up with maximum spacing,
5"
gap
The
experiment.
using
of spacing was
effect
actuator
an
investigated
compressor radius (R),
setup
(Figure 3.1)
down
the
between the two was varied from
to zero.
As
the
calculations
show
the velocity correction introduced with minimal
(H/R
=
0.05)
is in phase with
movement from -90 to +90 degrees.
the
IGV
introduced
by the IGVs.
magnitude,
will
lead
This,
to
an
stagger
the
For a larger spacing
correction shifts in the direction of the
velocity
for
appendix
(H), non-dimensionalized by the
IGVs as another. The spacing
spacing
thus
(see
analysis
disc
previous
the
for
with rotor and stator lumped into one disc and
derivation),
existing
stator,
the rotor and
between
a
turning
together with the change in
incorrect
estimate
in
the
restaggering calculations based on zero spacing. The spacing
between
the
the IGV row and the rotor was therefore reduced
minimum
corresponding
possible
to
with
H/R = 0.25,
the
existing
to
compressor
although this still
left
a
significant gap.
The
the
program that calculates the compressor response to
inlet distortion was not modified
spacing
between
to account
*for
the blade rows due to time constraints
the program.
23
this
on
3.4) Distortion Screen Design and Placement
The
which
lead
used
in
total
pressure
consists of 7
sections
a
to give
designed
of one dynamic head.
distortion
It
to an approximation of the sine
the
coefficients
screen
the
was
screen
rescaling
for each section were determined by
used
by
Bruce E113
The
to
obtain
circumferrential
loss
pressure
total
The
calculations.
defect
shaped
a
sinusoidal
velocity
profile.
pressure
loss coefficient of each section can be
the
extension and
*found
in
table 3.1.
screen
The
was
attached
to
the
struts
forward
supporting the centerbody 23 inches ahead of the IGVs,
it
where
could easily be removed to obtain uniform flow compressor
data.
With this large spacing some mixing was ex:pected,
the
total pressure profile measured at the IGV plane is shown in
figure
3.2.
The
overall profile approximates the
sinusoidal total pressure defect,
-90
desired
except for the reading at
degreesthe interface of the low and high total pressure
regions, which is lower than expected.
3.5)
Instrumentation Layout
The
instrumentation
circumferential
profile
was
in
set up to
a
front of the IGV plane
measure overall compressor performance.
for total pressure measurement,
and
to
pressure
Ten more sets were
downstream of the stator (Figure 3.3).
24
detailed
Twenty kiel probes,
and twenty static
taps were installed upstream of the IGVs.
installed
give
The close
an increase in
to
led
factor was calculated by representing
probe
its
and
-factor of
and
the static pressure measured
correction
of
taps
of the kiel probes behind the static pressure
spacing
wake as a source.
The resulting
the
a
kiel
correction
1.08 for the axial velocity agrees with the
ratio
the axial velocity measured by a pitot tube compared
to
obtained from the kiel probes and static pressure taps
that
(Figure 3.4).
Four
of
the static pressure taps,
and two in
plane
calculations
the downstream set,
two at
the
inlet
were not used in
as they gave incorrect readings whose
the
origion
could not be found.
The calibration curves for the two scannvalves used
obtain
the
pressure readings are shown in
figures 3.5
to
and
3.6.
3.6) Experimental Procedure and Data Aquisition
"The
basic plan of the experiment was as
follows.
compressor characteristics for IGV settings of +15,
relative to the design staggering were first
this
lT/ay
data
amount
of
restaggering
distortion
would
can be calculated.
required to
introduced by the screen.
then
This
remove
non-ax i symmetricly
staggered
obtained.
determines
the
The final
show the difference between the
compressor
0,
-15
With
the
velocity
speedlines
uniformly
and
The
show
and
the
improvement in the circumferrential velocity profile.
The
individual speedlines are taken starting with
25
the
completely open.
throttle
new
operating
point is
stall
is
slowly closed to a
point and the procedure repeated
reached.
Each of the
compressor
static pressures to give the total to static
rise.
until
the
operating
calculated using an average of the inlet total and
points is
exit
The throttle
The
axial
velocity is calculated by
pressure
averaging
the
local velocities obtained from the total and static pressure
probes.
also
definition of the axial flow
This
used
for
the distorted flow where it
coefficient
is
represents
an
average -flow coefficient for the whole compressor.
In
distorted flow the overall flow coefficient for
the
compressor obtained from averaging the individual velocities
gave a value up to 5% higher than a method using the average
pressure readings (Figure 3.'7).
all
axial
flow
coefficients
velocities at the inlet plane.
26
As defined here,
are
those
based
therefore,
on
the
Chapter 4 Experimental Results
4.1)
Compressor characteristics
The
an
IGV stagger settings were used.
IGV
from
previous
speedline
calculated
results
the
family
were performed
stagger
The
the
'/o
agreement
the
with
design
the
so new calculations of
using
the
.
As
4.1).
(Figure
degrees
of
setup was
inital
was only acceptable near
• = 0.59,
coefficient of
optimum
of 7.5
angle
stagger
expected
flow
were
needed to calculate the variablesaP/o and
speedlines
Several
data taken
of
sets
first
the
experimentally
obtained speedlines.
Two
of
+/-
speedlines were then taken for a change of stagger
degrees from nominal
15
increased
stagger
of
(Figures
22.5 degrees led to
rise and stall flow coefficient,
pressure
The
4.1).
a
decrease
in
as predicted
by
theory.
The
speedline with the IGV stagger setting of -7.5
degrees
showed an unacceptable drop in pressure
rise
near
dropping below the pressure rise for the +7.5 degree
stall,
stagger.
reason
T'he
separation
of
angle of attack.
of aY/7
for
the lower pressure rise
the airflow on the IGVs at a
sign of
T/qY
negative
was inappropriate f'or
range of stagger angles to be used in
4.2
the
As the l inear model uses only a mean value
a change in
Figure
high
is
the control sceme.
shows a speedline for a IGV stagger
degrees . Using linear
the
of
0
interpolation of the three speedlines
with stagger angles of +7.5,
0 and -7.5 degrees stagger the
27
which
at 3 degrees,
is
= 0
0a~y
stagger for which
critical
was taken as the point where the IGVs separated.
avoid
To
(Figure 4.2).
degrees
27.5
for
which
the
of
of
stagger
This speedline along with
they were chosen -for the
and
sign
the
15 and 22.5 degress showed the desired
speedlines "for +7.5,
stacking
angles
a speedline was taken with an IGV
changes
Dii/l~
stagger
(Figure
experiment
4.3).
nominal setting around which the
The
was
staggering
degrees
of
15
allowed for a maximum restagger of up to
10
then
which
non-axisymmetric
designed was a stagger
angle
degrees before separation occured.
static
The
(Figure
4.4)
accounted
pressure
showed
for.
A
readings
factor for
therefore obtained from the raw data.
(Figure
upstream struts supporting the centerbody,
and
in
probe
be
was
4.5).
The
the eccentricity
the variation of the blabe tip clearence
other imperfections in the compressor are accounted for
this correction.
The correction in
readings accounts for the permanent,
uniformities
only
each
not
The corrected velocity
profile was then satisfactory near stall
of the centerbody,
could
non-uniformiteis that
correction
flow
undistorted
for
in
changes
the static
i.e.
pressure
systematic,
the compressor and the subsequent will
due
to
the inlet
distortion
and
the
nonshow
IGV
restagger.
The
variations
same
correction,
and other
to account for blade
imperfections of the
to
compressor,
performed on the exit static pressure non-uniformities,
28
blade
was
but
then it was not possible to reduce the static
even
.05
at the exit plane to less than
non-uniformity
pressure
dynamic
heads (Figure 4.6).
4.2) Restaggering Scheme Calculations
order polynomial
third
fits
All
speedlines.
experimental
speedlines were fit
(Figure 4.7)
point of the nominal speedline.
compared with those
speed 1 ines.
generated
were
aYIJ/
the
The
the
with
a
and the resulting curve
used for a calculation cr'fTl/ayandaTy/aat
4.9) are
from
and DI/ao were determined
parametersaT//ay
The
any
operating
These values (Figure 4.8 and
calculated using the computer
experimental values
of ~DP/ayand
used to find the restaggering required to reduce
velocity
non-uniformity
created
by
the
distortion
screen. For the 1.0 dynamic head total pressure distortion a
restagger
of 10 degrees was needed (Figures 4.10 and
obtain
to
a sizable reduction in
the velocity
4.11)
distortion.
The actual restagger of the IGVs was a square wave with
a
+10
degree restagger
degree
restagger in
in
the low pressure zone and a
the high pressure
region.
wave
restagger
that
would show a benefit similar to the
method
of
was chosen as a simple sector
a sinusoidal restagger,
seen
a
in
square
restaggering
more
complicated
where each blade
have to be restaggered by a different amount.
using
"The
-10
would
The results of
square wave restagger in the calculations
can
be
figure
the
10
4.12.
The figure shows that with
degree square wave restaggering it
29
is
possible to reduce the
velocity non-uniformity by a factor of two.
4.3) Baseline Compressor Behaviour
The
compressor
the
run with
was
screen
distortion
installed and uniform IGV stagger angle to find the baseline
response.
characteristics with and without
The
figure 4.13.
are shown in
lower and the stall
The
0.37.
drop
distortion
7.9%
is
The pressure rise at stall
flow coefficient decreased from 0.38
in
pressure rise is
to
as
to be expected
a
result of the total pressure inlet distortion.
velocity profiles for the distorted flow operating
The
conditions
For all
4.15).
are
show a more pronounced effect (Figures 4.14
flow coefficients the shape of the
profiles
similar and follow that of the total pressure
The
large change in
position
three
can
be
supporting
flow coefficient at the
struts
which replaces the
profile.
degree
0
explained by the presence of one
IGV
and
the
of
at
this
location.
This large spike appears in
the swirl
introduced by the inlet distortion leads to a non-
uniform
incidence
angle
on
the distorted flow,
the
IGVs
around
circumference.
At this larger,
the
therefore have a stronger influence
strut will
as
the
non-uniform incidence angle
on
the
local flow coefficient.
The
non-uniformity
using three methods.
second
is
in
the inlet velocity
The first
method is
is
measured
the rms value,
the sum of the absolute values of the
the
difference
between the average and local flow coefficient and the third
30
is
Fourier analysis of the velocity profile showing
a
magnitude
measures
trend
the first
of
of
non-uniformity
that of
to
opposite
the
shows
data
the experimental
theoretical
increased
The velocity distortion
4.16 and 4.17).
towards
lower flow coefficients for the theoretical
distortion becomes more severe -for
the
in
coefficients
This
model,
higher
flow
the experiment.
change in
behaviour
setup
of the compressor.
rotor
and
compressor
a
calculations
(Figure
while
these
Using
coefficients.
three
the
is
thought to be due
to
Due to the separation between the
the stationary blade rows the behaviour
similarities to that of an
has
the
the
of
isolated
blade
rotor. Using an actuator disc model of an isolated rotor the
observed in
trend
the experiment can be predicted.
calculations show (Figure 4.18)
theory
will
therefore
experiment.
magnitude
This
of
the
the velocity non-uniformity,
in this case the maximum amplitude of a sinusoidal
distortion,
As
velocity
increase for higher flow coefficients. The
confirms
change
the
in
trend
behaviour
the individual distortions,
observed
will
in
the
modify
the
but it does
not
effect the control scheme otherwise.
4.4) Restaggered Compressor
With
Behaviour
the restagger implemented the same sets
were then taken to determine the improvement in
The
4.19.
The stall pressure rise is now
31
data
performance.
average performance given by the speedline is
Figure
of
shown
in
0.38 compared
to
from 0.37 to 0.38.
in
The loss in
the uncorrected case.
ý
undistorted
The
this
at
back
stall
has
the
to
the pressure and velocity profiles show
the flow
of
improvement
in
decreased
is
The restaggering of the IGVs
= 0.38.
changes in
pressure
smaller than
pressure rise is
coefficient
flow
the
brought
moved
coefficient
the baseline 0.36 and the stall flow
the
the region where
static
The
clearly.
more
are
vanes
opened relative to the uniform stagger setup (Figures 4.20).
This
to
leads
of' the distortion for all
reduction
non-uniformity
setup
(Figures
baseline
case,
and
velocity
increase of the
an
and 4.22)
shows
a
restaggered
change
but a direct comparison is
the non-uniformities in
a
The
flow coefficients.
of the velocity profiles in the
4.21
thus
the
from
due to
difficult
the profiles.
A more detailed examination of the restaggered profiles
shows a large increase in
and
290
degrees,
This "spike'
restaggering provides too much
total
is
region,
correction here.
thus leading to
two
The
restagger.
reduced even though this section is
pressure
wave
the result of
of the restagger square wave
effects
pressure is
the region between 250
as the calculations for the square
restagger had predicted.
side
velocity in
an
The static
not in
a low
increases
in
velocity.
There is also an.effect due to the spacing between
blade rows.
the
As the actuator disc calculations had shown the
spacing would lead to a correction that would be larger than
the calculated correction for zero spacing,
32
in
the direction
of the swirl
in
the
this
induced by the IGVs.
experiment due to the abrupt change in
location
therefore
be
from
5
25
to
degrees.
removed
if
the
using
a
measurement
interactively,
restagger
velocity to restagger the IGVs in
The increases in
360 degrees in
the
The spike is more extreme
4.24).
and 4.22
the
local
flow
200
are again a result of
for distortion magnitude
velocity
for all flow coefficients (Figure 4.23 and
For the higher flow coefficients the restagger is
effective as for the lower flow coefficients,
the
phase
and
struts.
desig~ned *for a flow coefficient of
of
of
could
implemented
a clearer decrease of a factor of two in the
non-uniformities
as
were
at
that region.
The comparison of the criteria
show
spike
The
the flow coefficient at 135,
figures 4.21
supporting
stagger
of the first
ý = 0.45.
three harmonics
not
as it was
A comparison
(Figure
4.25)
shows only small changes in the phase of the harmonics.
4.5) Limitations
These results of the response of the compressor to
distortion
and the resulting restaggering show some of
the
the
limits of the basic scheme.
The
exit
static pressure profile does not agree
the uniform pressure assumption of the model
(Figure
with
4.26).
This non-uniformity existed with and without the distortion,
and
the conclusion
is
that it
is
33
due to the compressor
and
not due to the restaggering of
The
of
shift
the IGVs.
of the static pressure profile in
rotor rotation relative to the original
distortion is
spacing
sudden
regions
ignored in
change in
leads
pressure
which assumes no
with
the
stagger angle at the interface of the
two
the
between
the linear model,
total
direction
blade rows.
This together
to an interaction of the
included in the parallel compressor model.
34
two
sections
not
Chapter 5 Conclusions
and Recomendations for Future Work
5.1)
Conclusions
of
results
The
the
of
implementation
is
They show that it
non-ideal test conditions.
very
possible to
the velocity distortion at the compressor face by
reduce
factor of two with an IGV restagger of 10 degrees.
with
non-
the
are encouraging even given
stagger
axisymmetric
the
A
a
setup
closer more typical spacing between the blade rows and
aerodynamically
compressors,
improved
IGVs,
typical of
current
axial
should increase the ability to reduce velocity
distortions and extend the method to even larger inlet total
pressure distortions.
Non-axisymmetric
restaggering also reduced the loss
in
pressure rise due to the inlet total pressure distortion.
5.2)Recomendations for Future Work
The
restaggering required of the IGVs is
possible
in
current
engines.
The
in
the
possibility
range
of
this method opens further questions on overall
implementing
engine performance that need to be investigated.
The
ability
distortions
has
in
of the scheme to adapt to different
plays a role in its usefulness.
If each
inlet
blade
to be controlled independently to achieve the reduction
velocity
non-uniformity
the improvements will
35
be
much
smaller,
due to the weight and complexity of the actuators,
than if
sectors of 60 to 90 degrees can be restaggered by a
A future investigation would therefore test
single actuator.
a
compressor
inlet total pressure distortion
with
in a certain section with varying
restagger
and
a
between
phase
the two, to determine how accurately the restaggering has to
be
positioned
to
f low
obtain a desired decrease in
non-
uniformity.
Similarly
the location and magnitude of the distortion will
the
level
of
determine
the number of sensors required to
accuracy needed to show
compressor behaviour.
an
depend
on
improvement
in
Here the work would concentrate on the
number of probes required to locate the inlet distortion.
The
implementation of this method in
require a quick method,
then
would
specify
sector
to
the
engine
extend
required
and
magnitude
distortion as well as the location of the current
point on the speedline.
with
would
such as a lookup table,
the amount of restagger
according
an
for
of
that
each
the
operating
Future work in this area would deal
the amount of time required between the sensing of
an
inlet distortion and the repositioning of the IGVs.
Before this method can be applied to actual compressors
further
gains
gained
sensing
work will have to be done to compare
of
implementing this method instead of
the
relative
improvements
by an additional compressor stage and the amount
required
to
make the
restagger ing
method of improving distortion tollerance.
36
an
of
efficient
REFERENCE
1)
Seymore, J. C., 'Aircraft Perormance Enhancement with
Active Compressor Stabilization',MIT Thesis Sept 1988
2)
Mazzaway, R. S., 'Multiple Segment Parallel Compressor
Model for Circumferential Flow Distortion' , Engineering
for Power, Vol. 99 No. 2, April 1977
3)
Stenning,
A.
Compressors',
March 1980
H. ,
'Inlet
Distortion Effects in
Journal of Fluid Engineering,
Axial
Vol.
102,
4)
Reid, C., 'The Response of Axial Flow Compressors to
Intake Flow Distortion:', ASME Paper 69-GT.-29, 1969 ASME
Gas Turbine Conference
5)
'A Method of Assessing
Hynes, T. P. and Greitzer, E. M.,
Effects of Circumferential Flow Distortion on
Compressor Stability', ASME Journal of Gas Turbines and
Power,
6)
Hynes, T. P., Chue, R., Greitzer, E. M. and Tan, C. S.,
'Calculation of Inlet Distortion Induced Compressor
Flow Field Instability', AGARD Conference Proceedings
No. 400, 'Engine Response to Distorted Inflow', March
1987
7)
Epstein, A. H., '"'Smart' Engine Components: A Micro in
Every Blade?', Aerospace America, Vol. 24, January 1986
8)
9)
Chen, G. T., 'Active Control of Turbomachinery
Instability - Initial Calculations and Results', MIT
Thesis August 1987
Raw, J. A., and Weir, G. C., "'The Prediction of OffDesign Characteristics of Axial and Axial/Centrifugal
Compressors', SAE Technical Paper 800628, April 1980
10) Rannie, W.D. ,'The Axial Compressor Stage', Section F
of 'Aerodynamics of Turbines and Compressors',
Princeton University Press, 1964
11)
Bruce, E. P., 'Design and Evaluation of Screens to
Produce Multi-Cycle
20% Amplitude Sinusoidal
Velocity Profiles', AIAA Paper No.
74-623, AIAA 8th
Aerodynamic Testing Conference, July 1974
37
Appendix
Actuator Disc Modeling
the origional calculations assumed
As the model used in
a compressor with closely spaced blade rows the changes
to
the large seperations in
be
investigated seperately.
the experimental
due
machine had to
Actuator disc theory allows
quick and simple estimation of the trends introduced by
a
the
spacing.
The
change
first set of calculations are an estimate
the
of
in ability to correct the flow due to the seperation
between the blade rows. The flow field is divided into three
regions,
the flow upstream of the IGV plane,
gap between IGVs and rotor,
stator.
all
The
the flow in
and the flow downstream of
the
the
rotor and stator are lumped into one disc,
the restaggering occours in
The flow in
the IGV blade row.
the three regions is
=
V1
-CO
respectively.
3
then governed by
=L ein
V22 = -)2
V2
as
= M e -in tan •1 (x/R) + in
(A-)
= -03 =N e-in tanP3 (xR) + in
The
magnitude of L is determined by the
magnitude of the inlet distortion and M and N are determined
by the loading on the blade rows.
The stream functions that satisfy these equations are
'1
=
A e ine +ndR + B e ein - nx/R + W e in e
38
'2 = C emne + nx/R + D eine - nx/R + E ein( - x/R tan a2)
in ( - n/R tan
in 0
Y 3 = F e ino + nx/R + G e - nx/R + H e
-
The
magni tude
of
constants
the
at
x =-- Co
at
x = co
At
each
leaving
is
bounded
=>
B = 0
is
bounded
=>
F = 0
of the blade rows three
angle
and at the
h.
'I
boundary
conditions
Contenuity has to be satisfied,
to be met.
have
cc
is
are
H
through
A
determined by the boundary conditions at x =
two blade rows x = 0,
)
the
flow
and
specified by the blade setting
the
pressure rise is given by the compressor characteristic.
at
x = 0
A + W = C + D + E
- itan.a
P2
R ae
A = C - D - i tan a 2 E
ae
)=
2
([2U1
( -3)
+2VlV]cp +
ae
2+V12] a)c
ae
ae
at x: =h
n R
C enh/R + D e-n R + E e-inh R tan a 2 = G e-h/
+H e-in R tan a3
in tan a2 (C en
R
+ D e- nW + E e- inR tan a2) = -
( A-4)
G e- nR + H e- nW tan a3
R
2 ]c+ [U+ +Vi )
R ae - ao )=2 ([2U2 ae +2V
a P
ae
the pressure depedency can be removed from the
equations by substituting the momentum equation
39
dynamic
Ua+ +
5x Rao
then
are
velocities
The
(A--5)
pRaO
the
in
expressed
streamfunctions and the equations linearized by dropping all
The resulting 6 x 6 system can then
second order terms.
evaluated
for
desired inlet distortion or
any
be
blade
IGV
distribution.
The other application is the calculation of the
the
flowfield
is divided into an upstream
this case
In
of an isolated rotor on a velocity distortion.
effect
downstream
and
region.
V 2i
=-
(A-5)
V2
2
= -02
In this case the pararmeter was the flow coefficient of
the compressor. As a result of this the pressure rise of the
blade
row
had
to be included as a function
coefficient.
The
experimental
speedline
only
intended
model
uses a curve-fit of
for this,
The change in
the
magnitude
for by using the slope of the characteristic at the
coefficient.
The
flow
of
are
the
pressure rise around the
circumference due to the velocity non-uniformity is
flow
flow
nominal
as the calculations
to show the trend in the
velocity distortion.
the
of
direction in
the
acounted
average
downstream
region is now also dependent on the flow coefficient, as the
relative leaving angle is
asumed to be constant.
Applying the boundary conditions at x =
and at x = 0
and performing the same substitutions as before leads to a 3
£40
x
3
rotor.
system
that can be solved to find the effect
of
the
IGV
Rotor
Stator
Hub Diameter (rrmn)
444
444
444
Casing Diameter (mn)
597
597
597
Number of Blades
46
44
45
Chord (mrr)
33
38
38
Solidity at Midspan
0.9
1.0
1.0
Aspect Ratio
2.2
1.9
1.9
Camber (deg)
12
25.5
30.0
Midspan Stagger (deg)
15
28.7
45.0
1.6
0.8
Blade Clearance (rrmn)
Table 2.1
Compressor Geometry
1.5
Sector
(deg)
Loss Coefficient
Pt/.5 Rho Cx^2
-90 to -70
1.02
-70 to -30
1.34
-30 to
30
1.72
30 to
70
1.34
70 to
90
1.02
90 to -90
0.32
Table 3.1
Distortion Screen Total Pressure
Loss Coefficient
43
OPERATING LINE SHIFT
OR
SURGE LINE SHIFT
5.5%
5.5%
REMAINING
AUGMENTOR SEQUENCING
7.5%
INLET DISTORTION
2.5%
ENGINE VARIATION
1.5%
CONTROL TOLERANCE
1.5%
REYNOLDS NUMBER
Fig. 1.1
USABLE SURGE MARGIN
MINIMUM SURGE MARGIN
7
Typical Allocation of Required Surge Margin [1]
C
cli
I* 1
I 1 1W
I
10Tal
essure Zone
0.C
i--
PT
aQ)
High To
Press. Z
0
I2
SInlet Distortion
'1)
a..
O
I
0
u
0.2
I
I
0.4
0.6
I
0.8
Mass Flow, Cx/U
Fig. 1.2 Parallel Compressor Description of Response to
Circumferrential Distortion
45
CuJ
4-
I
D
.
N IUT
Il
Rss. Zone
C'3
0
High Total
Press. Zon
a-
Opening
Vanes
L0
U)
U)
E
0
0)
Fig. 1.3
ominal
Mass Fl ow, Cx/U
Effect of Non-Axisymmetric Vane Stagger on Compressor
Response to Total Pressure Distortion
.08
-9-
0
"
o
.04
0
u
.o -. 04
n9
0
Fig. 2.1
360
120
240
Circumferential Position, 9 (Deg)
Three Stage Compressor Response to Three Restaggering
Schemes [8]
'06
1
Total pressure
.05 - loss parameter
Fully stalled
CjaLS
2C1 Co
Cos 82
02
cascade
.04
.03
.02
Slgnlficant cascade
--- stall starts here
.01
0
.3 .4
.0a
*j *q
.
.6 .1.7
DFAC =1 -V+(S) ACu
v, c 2TV -
.1
.I
0U
*
.8 .9
.
Fig. 2.2 Empirical Curve Relating Loss Coefficient and
Diffusion Factor [91
n
i.0n
5
%.U
Multiplying
factor on
@ML
3.0
2.0
,
l
ix
- _
fig 151
SP-38,
,51
-
I
-
1
'---royno
rence
efe
2
3
e
L
L Cc
a number 500,000
1 1 1 11 1
1 1 1
1
Ii
II
I
I
IN)'" power law
-
--
I
4
6
7
8
9
10
11
-
~-
~-
1
R*-blade chord reynolds number x 10-
Fig. 2.3 Multiplying Factor on Loss Coefficient as a Function
of Diffusion Factor [9]
43
'- I
9
DFAC = 1 - V + (S)ACu
Fig. 2.4
Deviation 'Adder' Applied to Carter's Rule as a
Function of Diffusion Factor [9]
1.5
1.0
/s
.5
03
0.2
0.4
0.6
0.8
1.0
1.2
Fig. 2.5 Calculated and Experimental Data for 4 Low Speed
Three-Stage Compressors [8]
ROTOR STAGGER = 28.70
ROTOR CAMBER = 25.50
STATOR STAGGER = 45.00
STATOR CAMBER = 30.00
o
o
-- ,
00
II
C'
-
a-fI, C)
1IGV EXIT ANGLE = 15.0
0.35
0.40
0.15
0.50
0.55
FLOW COEFFICIENT
Fig. 2.6
0.60
0.65
(0)
Compressor Speedline for Control Modeling
0.70
0
IGV STAGGER
,7
0
.- ,
0.40
0.45
0.50
0.55
0.60
FLOW COEFFICIENT - ()
Fig. 2.7
Experimental Speedline
0.65
0.70
0.75
15.0
ROTOR STAGGER - 32.80
ROTOR CAMBER - 25.50
STATOR STAGGER - 45.00
STATOR CAMBER * 26.40
OPENED VANES
CC
<0
o
I 0in
- 0
"O~
CLOSED VANES
a00
U)c
0=.
-0
EXIT
0.35
0.40
0.45
0.50
0.55
FLOW COEFFICIENT
Fig. 2.8
0.60
0.65
0.70
(40)
Calculated Compressor Speedlines with a 10 Degree
Opening and Closing of IGVs and Stators
53
LuJ
U)
(n
cc
CE
I0I--
45.00
90.00
135.00
180.00
225.00
270.00
315.00
THETR
Fig. 2.9 Inlet Total Pressure Distortion Profile
360.00
0.0 IGV MAX SWING
5.0 IGV MAX SWING
in
0.0 STATOR HRAX SWING
5.0 STATOR HAX SWING
0
eo0
ZCC0;
co
>- 0;
I-o
0
r,•
Oo
1
-J=
_o,
i-o
o0.0r=,
-JI o
cr
0
PHI
0.50
X•
-(I
IGV PHASE
STATOR PHASE
45.00
90.00
135.00
180.00
225.00
270.00
315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 2.10 Compressor Response to a 1.0 Dynamic Head Distortion
with a 0,5,10,15 Degree Restagger of the IGVs and
Stators Near Stall
T
in0.0 IGV MAX SWING
m 5.0 ICV HAX SWING
0.0 STATOR MAX SWING
5.0 STATOR MAX SWING
0
0
In
0
z 00
C;
in
o
co
05
IU,
x
-4U
0. 0T
o
PHI
0.70
IGV PHASE
0.0
STATOR PHASE 0.0
-·I
45.00
90.00
135.00
180.00 225.00
270.00
315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 2.11 Compressor Response to a 1.0 Dynamic Head Distortion
with a 0,5,10,15 Degree Restagger of the IGVs and
Stators at Design Flow
55
(90.0 IGV HAX SHING
M 8.0 IGV HAX SWING
in
0.0 STATOR HRAX SWING
6.0 STATOR HRAX SWING
a
o
C;
C;
atn
z9
Qo
05
--
oC;
CC
0. v-u
o
.o
-J
LLJW
a:
•o
IO,
PHI
0.50
X•
-4(I
IGV PHASE
0.0
STATOR PHASE 0.0
45.00
90.00
135.00
180.00
225.00 270.00
315.00
360.00
CIRCUMFERENTIAL POSITION
Fig. 2.12 Compressor Response to a 1.0 Dynamic Head Distotrion
with a 0,8,12,16 Degree Graduated Restagger of the
IGVs and Stators Near Stall
SI0.0 IGV HAX SWING
m 8.0 IGV HAX SWING
PHI
0.0 STATOR HAX SWING
6.0 STATOR HAX SWING
0.70
IGV PHASE
0.0
STATOR PHASE 0.0
45.00
90.00
135.00
180.00
225.00 270.00
315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 2.13 Compressor Response to a 1.0 Dynamic Head Distotrion
with a 0,8,12,16 Degree Graduated Restagger of the.
IGVs and Stators at Design Flow
0.0 IGV MAX SWING
5.0 IGV HAX SWING
!)
0.0 STATOR MAX SWING
3.0 STATOR HAX SWING
0
o
<3U,
0
edW
zQo
•I-o
Yo
C3
CC0.
-J
L
)
0
_ I
PHI
0.50
0:0
X
Q:=
IGVYPHASE
-15.0
STATOR PHASE -30.0
45.00
7
90.00 135.00 180.00 225.00 270.00 315.00 360. 00
CIRCUMFERENTIAL POSITION
Fig. 2.14 Compressor Response to a 1.0 Dynamic Head Distotrion
with a Shifted Graduated Restagger of the IGVs and
Stators Near Stall
0.0 IGV MAX SWING
5.0 IGV MAX SWING
0.0 STATOR HAX SWING
3.0 STATOR HAX SWING
i-e
i-0
I-
>-
-J
U
-.J
a:
6-
PHI
0.70
IGV PHASE
-15.0
STATOR PHASE -30.0
45.00
90.00
135.00
180.00
225.00
270.00
315.00
360.00
CIRCUMFERENTIAL POSITION
Fig. 2.15 Compressor Response to a 1.0 Dynamic Head Distotrion
with a Shifted Graduated Restagger Of the IGVs and
Stators at Design Flow
0.0 IGV HAX SWING 0.0 STATOR HRAX SWING
20.0 IGV HAX SWING 0.0 STATOR HRAX SWING
PHI
0.50
IGV PHASE
0.0
STATOR PHASE 0.0
'5.00
90.00
135.00
180.00
225.00 270.00
·-·_,
__
315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 2.16 Compressor Response to a 1.0 Dynamic Head Distortion
with a 0,20,30,40 Degree IGV Restagger Near Stall
SWING 0.0 STATOR HAX SWING
0.0 IGV MAHRX
SWING
m 11.0 IGV MAX SWING 0.0 STATOR MAHRX
z
c
to
0
.I
u
cc
I._i
PHI
x
a:
X
cr
0.70
IGV PHASE
0.0
STATOR PHASE 0.0
50.00
CIRCUMFERENTIRL POSITION
Fig. 2.17 Compressor Response to a 1.0 Dynamic Head Distortion
with a 0,11,22,33 Degree IGV Restagger at Design Flow
= 0.050 H/R
= 0.250 H/R
= 0.500 H/R
O0
THETR
Fig. 3.1
Axial Velocity Perturbation Introduced by a 10 Degree IGV
Restagger v&. Blade Row Spacing
It¢•
O
(c
U
ZI
z
D~
O)
uJ
CrC
aJ
cc
-135.00 -90.00
-45.00
0.00
115.00
90.00
135.00
CIRCUMFERENTIRL POSITION
Fig. 3.2 Circumferential Inlet Total Pressure Profile with
Inlet Distortion
60
180.00
o
a
8
8
8
8A
8
8
A
A
8
o
-r,
--0 0
z
o
-O
d(
0
0 STATIC PRESSURE TAP
A TOTAL PRESSURE KIEL PROBE
02
DISTORTION SCREEN
8.00
16.00
24.00
32.00
40.00
48.00 56.00 64.00
CIRCUMFERENTIAL POSITION
Fig. 3.3
Pressure Instrumentation Layout
-
0.40
0.5
0.50
0.55
0.60
FLOW COEFFICIENT
0.65
0.70
0.75
(0]
Fig. 3.4 Velocity Measured by Pitot Tube vs. Velocity Measured
by Total Pressure Kiel Probes and Static Pressure Taps
62
0.0003
RESIDUAL
CORRELATION
1.2231
Lo
I
-30.00
0.00
-20.00 -10.00
PRESSURE
10.00
20.00
40.00
30.00
(IN H20)
Fig. 3.5 Scannivalve 1 Pressure Transducer Calibration Curve
RESIDUAL
CORRELATIO
5808
-30.00 -20.00 -10.00
PRESSURE
Fig. 3.6
0.00
10.00
20.00
30.00
40.00
(IN H20)
Scannivalve 2 Pressure Transducer Calibration Curve
63
tn
Ca
ULL
Oo
0
J c
C)
-D
A
U)
-V
0
CC,
LU
CIC30
i-n
z=r
IL o
,-,O
U
LL. c3
LL=
U-
Dt(t
-o
IL
WI
r
0.35
0.40
0.45
FLOW COEFFICIENT
Fig. 3.7
0.50
IV
•I
0.55
0.60
0.65
0.70
(AVERRGE PRESSURE)
Flow Coefficient from Average Pressure Reading vs
Coefficient from Averaged Local Flow
64
C
ITV STAGGFR
7.5
22.5
ýC)
u
-a
c;
-7.5
I-o
Lu
LL C
Li I
LLJ
a
LLI
rr-"
LuLuin
C
U) ,
(ILu
a-
0.40
Fig. 4.1
0.45
0.65
0.50
0.55
0.60
FLOW COEFFICIENT (0)
0.70
0.75
Compressor Speedlines for 7.5, -7.5 and 22.5
Degree IGV Stagger
IMV STAGGER
0.0
27.5
zLU
u
L)
L..
U-
M
Lu
c)
Lu
Cr)
cn
Lu
cc
Cr)
a:
a-
0.40
0.q5
0.50
0.55
0.60
FLOW COEFFICIENT
Fig. 4.2
0.65
0.70
0.75
(0)
Compressor Speedlines for 0.0 and 27.5 Degree
IGV Stagger
65
C
IGV STRGGER
U~)
U)CU
°-
k--O
D
7.5
A
22.5
+
15.0
I.Ln
_O
z c;
.o,
LLiO
Pi uu
C-)
LU
0
~0
0.40
0.45
0.50
0.55
0.60
0.65
0.70
0.75
FLOW COEFFICIENT (4)
Fig. 4.3
Compressor Speedline for 7.5, 15.0 and 22.5
Degree IGV Stagger
0
0
...........
U,
CO
acc'
(30
CC
U,
PHI = 0.57
45.00
90.00
135.00 180.00 225.00 270.00 315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 4.4 Inlet Static Pressure Profile at Design Flow
0
U,
<10
zC3
Zo
0o
-.
a
cc
L-o
0-0
U
0
'
>-
PHI = 0.40
45s.oo00
90.00
135.00 180.00 225.00 270.00 315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 4.5
Inlet Velocity Profile Near Stall
67
o
oo
(Z.
LU 0
Cr
(o
aeo
cc0
z
~0
LU
IJo
cc'
q:
i-
a.
.o,
U,
ivl
PHI - 0.57
45.00
90.00 135.00 180.00 225.00 270.00 315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 4.6 Exit Static Pressure Profile at Design Flow
68
7.5
l--
*
A
22.5
27.5
75
FLOW COEFFICIENT
Fig. 4.7
(0)
Compressor Speedline Curvefit for Control Variable
Calculation
CALCULATED
D
EXPERIMENTAL &
oC3
C0
70
FLOW COEFFICIENT ()Change
n 1G
Fig. 4.8 Change in Pressure Rise due to Change In IGV Stagger
Angle vs Flow Coefficient
CALCULATED
EXPERIMENTAL
o
&
70
FLOW COEFFICIENT
(0)
Fig. 4.9 Compressor Characteristic Slope vs Flow Coefficient
(! 0.0 IGY MAX SWING
m 5.0 IGV MAX SWING
PHI
0.0 STATOR HRAX SWING
0.0 STATOR HAX sqING
0.116
IGV PHASE
STATOR PHASE
41.00
Fig. 4.10
90.00 135.00 180.00 225.00 270.00 315.00 360.00
CIRCUMFERENTIAL POSITION
Calculated Compressor Response to a 1.0 Dynamic Head
Distortion with 0,5,10,15 Degree IGV Restagger Near
Stall
0
0
a
+
o
0.0 IGV HAX SWING 0.0 STATOR MHAX SWING
5.0 IGV MAX SWING 0.0 STATOR MAX SWING
10.0 I[GVHAX SWING 0.0 STATOR MAX SWING
15.0 IGV MAX SWING 0.0 STATOR HRAX SWING
eC<1
o
e'o
(1
(..
X
C, M
r- o ,
c-J
0v
0.58
IGV PHASE
0.0
0.0
STATOR PHRSE
PHASE 0.0
PHRSE
[GV s5.00
STRTOR
90.00
135.00 180.00o 225.00 270.00 315.00 360.00
CIRCUMFERENTIAL POSITION
Fig. 4.11
Calculated Compressor Response to a 1.0 Dynamic Head
Distortion with 0,5,10,15 Degree IGV Restagger at High
Flow Coefficient
I30.0 IGV HAX
I SWING
-
5.0 IGVHRX
SWING
0.0 STATOR nRX SWING
0.0 STRTOR MAX SWING
180.00
225.00
B
z
o
!0
I.C3,
U
I-0
-C.
LU
-J
u
cz:
~r
'15.00
90.00
135.00
270.00
315.00
360.00
CIRCUMFERENTIAL POSITION
Fig. 4.12 Calculated Compressor Response to a 1.0 Dynamic Head
Distortion with 0,5,10,15 Degree Square Wave IGV
Restagger at Design Flow
a
0
C)ll
C)(
- n
U') (\2
c)
* O0
I-ý
U
0
LL.
on
LLJ
L_)
CE
•.
O
cn.
o,
o
DISTORTED
m
UNDISTORTED
r(I-
a-
_
0.40
·
0.45
0.50
0.55
0.60
0.65
0.70
0.75
FLOW COEFFICIENT ($)
Fig. 4.13 Compressor Speedlines for 15 Degree IGV Stagger with
and without Inlet Total Pressure Distortion
LOW
__ PT REGION
3.00
CIRCUMFERENTIAL POSITION
Fig. 4.14 Inlet Velocity Profile for Baseline Compressor with
Inlet Total Pressure Distortion Near Stall
RFfI¶M
PTRFGIC
10W PT
Iflu
PHI - 0.42
-135.00 -90.00
-45.00
0.00
45.00
90.00
135.00
180.00
CIRCUMFERENTIAL POSITION
Fig. 4.15 Inlet Velocity Profile for Baseline Compressor with
Inlet Total Pressure Distortion at Design Flow
o
0
>-0
tn
BASELINE SUM
ca
0C3
zo
Lc
LL_
0
z
I C
CC
a
z
BASELINE RMS
•ZO
cFl.
crc
0.35
0.60
0.50
0.55
0.0
0.45
FLOW COEFFICIENT (0)
0.65
0.70
Fig. 4.16 Velocity Non-Uniformities for Baseline Compressor
with 1.0 Dynamic Head Inlet Total Pressure Distortion
BASELINE COMPRESSOR
oFIRST HARMONIC
SSECOND
L_
HARMONIC
+THIRD HARMONIC
z
LU Ln
LLlt
LU
(-3Oc
C
a
L_
U_
:3o
oSa
LL
C)
0
LUO
:3,
..........
cE
,.-••
._
Q_
0.40
Fig. 4.17
0.145
0.50
0.55
0.60
0.65
FLOW COEFFICIENT (0)
0.70
0.75
Magnitude of First, Second and Third Fourier
Coefficients of Velocity Non-Uniformity for Baseline
Compressor with Inlet Total Pressure Distortion
o
Cý
Cr)
o"
Ln
Cu
xo
_.o
x C
"-' cu"
U~U
Z
N-- ,
0
z
n- •
C
cc
C.
Uj C3
0.35
0.40
0.45
0.50
0.55
RVERAGE FLOW COEFFICIENT
Fig. 4.18
0.60
0.65
(W)
Calculated Magnitude of Axial Velocity Perturbation
for a Isolated Rotor with a 10% Inlet Velocity NonUniformity
0.70
0
0
U-(,c
*0
z
LLJ
-0
LL.u.
LLtrOi0
L.U
U
LU
o
RESTAGGERED
(n
A
BASELINE
cc:
LU
ar
a-
0.40
Fig. 4.19
0.65
0.60
0.55
0.50
0.115
FLOW COEFFICIENT (0)
0.70
0.75
Restaggered and Baseline Compressor Speedlines with
Inlet Total Pressure Distortion
(
LLJ
ccZI
Co
I:
IW--,
" o
U,
a-0o
YOD
Q
U
ICU
cr.
UU)
1.00
CIRCUMFERENTIAL POSITION
Fig. 4.20 Inlet Static Pressure Profile for Baseline and
Restaggered Compressor Near Stall
N
PHI
PT
OL
REG
!0N
a 0.41
45.00
90.00
135.00
180.00
225.00
270.00
315.00
360.00
CIRCUMFERENTIAL POSITION
Fig. 4.21
Inlet Velocity Profile for Compressor with Total
Pressure Inlet Distortion and IGV Restagger Near
Stall
e
Q
z
0cr
I-
LI-
aU
I-
-J
Lii
).00
CIRCUMFERENTIAL POSITION
Fig. 4.22 Inlet Velocity Profile for Compressor with Inlet
Total Pressure Distortion and IGV Restagger at
Design Flow
79
>-.
BASELINE SUM
O
Z"C
Zo
RESTAGGERED SUM
Zo o
U_)
z..
mZmO
Ocr
.
3
o Co
BASELINE RMS
RESTRGGERED RMS
cr
~p"-~a-t~-~
0.35
Fig. 4.23
(to
0
0
8-
z
~
0.:55
0.60
0.50
0.40
0.45
FLOW COEFFICIENT (0)
0.65
0.70
Velocity Non-Uniformities for Baseline and Restaggered
Compressor with 1.0 Dynamic Head Inlet Total Pressure
Distortion
BASELINE COMPRESSOR
oFIRST HARMONIC
-SECOND HARMONIC
+THIRD HARMONIC
RESTAGGERED COMPRESSOR
xFIRST HARMONIC
oSECOND HARMONIC
*THIRD HARMONIC
Lu ,
CO
U
U0
Jc
~0
LL
Lu
M
0
Lu
z
Cr
C;,
LU
C.-,
_j
cLuO
0 Do
0
•
z-ra- o
._J
Q..
0.40
0.45
0.50
0.55
0.60
0.65
FLOW COEFFICIENT (0)
0.70
0.75
Fig. 4.24 Magnitude of First, Second and Third Fourier
Coefficients of Velocity Non-Uniformity for Baseline
and Restaggered Compressor with Inlet Total Pressure
Distortion
80
BASELINE COMPRESSOR
eFIRST HARMONIC
&SECONO HARMONIC
+THIRD HARMONIC
RESTAGGERED COMPRESSOR
xFIRST HARMONIC
*SECOND HARMONIC
*THIRD HRARMONIC
o
Io
z.
LUJ
0
WO
CU
C-1
P-0
U(JWo
,,co
LU.
0
o
Uj
LLI0
00
L. Q
00
U 0
(n.
a-I
0.40
0.oL5
0.50
0.55
0.60
FLOW COEFFICIENT
0.65
0.70
0.75
(0)
Fig. 4.25 Phase of First, Second and Third Fourier Coefficient
of Velocity. Non-Uniformity for Baseline and Restaggered
Compressor with Inlet Total Pressure Distortion
0
a
Wo
'r
U
Cr•o
cr*.
z0
LU
cco
0
cn
CC
ccC13
(n:o a
a-
U,
0
UNDISTORTED
RESTAGGERED
PHI = 0.60
45.00
90.00
135.00
180.00
225.00
270.00
315.00
360.00
CIRCUMFERENTIAL POSITION
Fig. 4.26 Exit Static Pressure Profile for Undistorted and
Restaggered Compressor with Inlet Total Pressure
Distortion at Design Flow
82