An instability associated with an intershaft squeeze film damper is described. The squeeze film is located between the intershaft bearing outer race and the low-speed shaft of a five-bearing, two-spool test rig. The instability is dominated by response of the third system mode to destabilizing excitation of the type described by Hibner, et al. Installing a spring cage in place of the intershaft damper removes the instability and produces satisfactory performance throughout the operating range

Alderson, R. G.

English

NASA Technical Reports Server

N87"22215
INSTABILITY OF AN INTERSHAFT SQUEEZE FILM DAMPER IN A
TWO-SPOOL ROTOR DYNAMICS SIMULATOR
R.G. Alderson
Garrett Corporation
Phoenix, Arizona 86010
ABSTRACT
An instability associated with an
intershaft squeeze film damper is de-
scribed. The squeeze film is located
between the intershaft bearing outer
race and the low-speed shaft of a five-
bearing, two-spool test rig. The
instability is dominated by response of
the third system mode to destabilizing
excitation of the type described by
Hibner, et al. Installing a spring
cage in place of the intershaft damper
removes the instability and produces
satisfactory performance throughout the
operating range.
INTRODUCTION
Trends in advanced technology en-
gines are toward significantly higher
turbine inlet temperatures, higher work
stages, increased thrust-to-weight ra-
tios, greater durability requirements,
and reduced life-cycle costs. These
trends are producing departures from
traditional rotor systems that have
significant effects on system rotor
dynamics.
One Garrett design in this area is a
two-spool engine that uses an inter-
shaft roller bearing to support the
turbine end of the high-pressure (HP)
spool. This eliminates the high-
temperature structure otherwise needed
to support the HP turbine-end bearing.
However, the intershaft bearing is a
path for nonsynchronous excitation not
found in traditional configurations.
Shaft diameters are kept to a mini-
mum in order that disk bores and bear-
ing diameters will be minimized. These
measures for improving disk low-cycle
fatigue lives and bearing lives are ac-
complished without requiring the low-
pressure (LP) spool to operate super-
critically. Although many of the modes
encountered in the operating range dis-
play some flexure, the synchronous
flexural critical speeds of the LP rotor
are outside of the operating range.
However, nonsynchronous excitation of
these modes by the HP rotor via the in-
tershaft bearing needs to be considered.
Shafting technology for advanced pro-
pulsion engines has not yet progressed
to the extent that satisfactory opera-
tion can be guaranteed in this regime on
the basis of analytical predictions
alone. Commitment of an advanced engine
design to a configuration incorporating
these features requires a program of
testing with a rotor dynamics simulator.
The dynamics rig designed to simulate
the engine dynamics is a two-spool, co-
rotating system driven by air impinge-
ment. Figure 1 depicts the dynamic sys-
tem in schematic form. Individual
spools are separately controlled and
driven. The LP spool is supported by
three bearings, two straddling the simu-
lated fan, with the third bearing aft of
the simulated LP turbine. The HP spool
is supported by two bearings, including
the intershaft bearing that has the
inner race carried by the HP rotor.
The rotor system is installed in a
rigid frame, with the spring cages used
as bearing supports simulating engine
structural stiffnesses. Each of the
four structure-supported bearings is
provided with an unsealed squeeze film
damper. Centering springs are not used.
315
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Originally, a squeeze film damper
was also used with the intershaft bear-
ing. This feature was employed in an
effort to exploit fully the benefits of
squeeze film dampers in controlling
unbalance. The oil film was formed
between the outer race and a housing
rotating with the LP spool. The subse-
quent need to eliminate a subsynchro-
nous instability led to removal of this
damper in favor of a spring cage.
Initial rig testing revealed sub-
synchronous whirl that appeared at a
part-speed condition and persisted
throughout the operating range. The
instability did not grow to destructive
proportions, probably due to relatively
well-balanced rotors. It was the domi-
nant feature of the rotor response,
however, and unacceptable for engine
operation.
The intent of this paper is to de-
scribe the characteristics of the
instability, and to discuss the influ-
ence of the intershaft squeeze film
damper.
EARLY TEST EXPERIENCE
Figure 2 shows the speed schedule
followed in early tests to approximate
engine conditions. Acceleration
through the speed range revealed some
system modes synchronous with the LP
rotor, while others were synchronous
with the HP rotor. Initially, the LP-
synchronous modes tended to be more
severe. But trim balancing effectively
reduced all synchronous response to
relatively low levels. Subsynchronous
response also was observed, beginning
at a part-speed condition and persist-
ing throughout the speed range. Trim
balancing was not effective in reducing
the amplitude of the subsynchronous
response.
Figure 3 is an RPM spectrum map that
typifies the system response. This
data was obtained from a proximity
probe sensing displacement of the simu-
lated HP turbine wheel. This spectrum
map is identical in character to maps
for the simulated HP compressor and LP
turbine wheels. These similarities em-
phasize the dynamic coupling provided by
the intershaft bearing. Proximity probe
data from the simulated fan showed less
prominent subsynchronous response,
probably due to the combined effects of
fan mass and LP shaft flexibility.
Figure 3 shows the main subsynch-
ronous response near 120 Hz, with a sec-
ondary subsynchronous response near
60Hz. The maximum amplitude of the 120
Hz component is in the range of 0.10 to
0.15 mm (4 to 6 mil) peak-to-peak for
the HP turbine, and about twiue that for
the LP turbine. The 120 Hz component
displays an increasing trend from its
initiation through maximum operational
speed for the rig. This response is
regarded as an instability. Although
these subsynchronous whirl amplitudes
are not a threat to rig integrity, the
presence of an instability would not be
tolerable for engine operation.
CHARACTERIZATION OF RESPONSE
Figure 4 combines Figures 2 and 3
with the results of undamped whirl speed
calculations. The solid lines define
system whirl frequencies for various
whirl ratios (LP spin speed/system whirl
speed). The analysis was done in the
whirl frame of reference, and extended
from zero whirl ratio (nonrotating sys-
tem) through unity whirl ratio (LP un-
balance excitation). Whirl ratios cor-
responding to HP unbalance excitation
vary according to the speed schedule.
At 100 percent speed, the whirl ratio
corresponding to HP unbalance is between
0.5 and 0.6.
Figure 4 also shows that the analyti-
cal model is in reasonable correlation
with the test data. Particularly
prominent is the agreement between re-
sponse peaks and calculated whirl speeds
for LP unbalance excitation (unity whirl
ratio). The first three mode shapes
corresponding to LP unbalance excitation
are shown in Figure 5. Amplitude and
phase data from the various displacement
transducers are consistent with these
mode shapes. All three modes display
316
shaft flexure but are not dominated by
it. Analysis shows that most of the
strain energy in these modes results
from support participation rather than
shaft flexure.
Figure 4 suggests that the sub-
synchronous activity around 60 Hz and
120 Hz is associated with the first and
third system modes, respectively. Fig-
ure 5 shows that these two modes (par-
ticularly mode 3) involve participation
by the intershaft bearing and squeeze
film damper.
INSTABILITY OF INTERSHAFT DAMPER
The intershaft bearing is installed
with the inner race mounted on the HP
rotor. The squeeze film is formed be-
tween the outer race and a housing that
rotates with the LP rotor. The damper
is of the open-end type. Following the
approach of Hibner, Kirk, and Buono
(Ref. I), the intershaft damper was
modeled by a solution of the Reynolds
Equation:
6 2 _8 + _-{ =
3h 3h (i)
(wl + w2 - 2_ _ + _ 3-[
Making the "short bearing" assump-
tion permits the first term on the left
to be neglected. For steady whirl in a
circular orbit,
h = C (I + e cose) (2)
as is shown in Figure 6.
Then (i) reduces to
32P 12pc sin8 (WL__)
= - C 2 (I+_ cos0) 3
(3)
Assuming that the pressure is symmetric
across the width of the squeeze film,
(3) can be integrated to obtain the
pressure distribution:
6_/(z2-L_ _ sin@
P(Z,8) = - C2 (l+e cosS) 3
(WL-%) (4)
For the open-end damper, the circumfer-
ential pressure distribution is approx-
imated by the "w-film" in which pressure
is positive over half of the annulus and
zero elsewhere. Two cases need to be
considered in the integration of the
pressure to obtain the radial and tan-
gential damper forces. In the first
case, the damper spin speed is less than
the journal whirl rate. The opposite
situation constitutes the second case.
Case 1 - _ > WL, P=0 for 0_< 8_<
L
P(Z,8) cos8 RdSdZ =
2_RL3 _ (%-wL) (5)
- C2 (i-_2) 2
which tends to center the journal.
L
Ft = 2 _2n/_[ P(Z,8) sin@ RdedZ =
o
- 2c _ (i__213/2 (b-w E) (6)
which tends to oppose whirl (stabiliz-
ing).
Case 2 - $ < WL, P=0 for _ _< 8 <_ 21T
L
F r = 2 P(Z,e) COS@ RdedZ =
2pRL 3 e2
_ C2 (1__2) 2 (WL-_) (7)
which tends to center the journal, as
in Case i.
L
F t =/o_/o _ P(Z,8} sin8 RdedZ =
+ (i_e213/2 (WL-_) (8)
31/
which tends to promote whirl (destabil-
izing). This is the opposite of the
effect noted in Case i.
This analysis, together with the
mode shapes shown in Figure 5, offers
an interpretation of the system re-
sponse shown in Figure 4. As the LP
rotor is accelerated, LP-synchronous
response in each of the first three
system modes can be seen at LP speeds
around 3300, 5000, and 7200 rpm. Ac-
cording to the analysis, intershaft
damper forces due to LP-synchronous
whirl are small since _ = W L. Simulta-
neously, HP-synchronous whirl is pre-
sent due to excitation of the second
and third system modes by the HP rotor.
Excitation of the sixth mode probably
occurs also. Stabilizing forces are
developed in the intershaft damper due
to HP-synchronous whirl since _ > W L.
Evidently these stabilizing forces, to-
gether with damping furnished by the
grounded squeeze film dampers, are the
dominant influences on system response.
For LP speeds above 8000 rpm, the
character of the response changes.
Following the expected sharp drop in
LP-synchronous response as the LP rotor
passes through the third system mode, a
subsynchronous response appears around
120 Hz. The amplitude of the sub-
synchronous response continues to in-
crease throughout the operating range.
The unique feature of the third system
mode (in contrast with the first and
second modes) is the degree of partici-
pation of the intershaft damper. This
is illustrated in Figure 5.
A plausible explanation of the re-
sponse derives from the abundance of
transients in the system. Noncircular
damper orbits and the nonlinearity of
squeeze film damper forces are primary
sources of transients which ordinarily
are of little consequence. Even though
system response to transients will tend
to include response in the system
modes, system damping usually suppres-
ses the transients immediately. But
because the intershaft damper partici-
pates strongly in the third system
mode, destabilizing forces become signi-
ficant since _ < W L for this mode. It
appears that the secondary subsynchro-
nous response around 60 Hz that occurs
in the same speed range as the 120 Hz
instability is a similar but weaker phe-
nomenon. Damper participation in the
first system mode is significantly re-
duced compared with the third mode.
Virtually no intershaft damper activity
occurs in the second system mode, nor is
there subsynchronous response associ-
ated with it.
REMOVAL OF INSTABILITY
A limited test program was undertaken
in which various system parameters were
modified in an attempt to remove the in-
stability. These modifications included
clearance changes in the intershaft dam-
per, the HP compressor-end damper, and
the LP turbine-end damper. Support
stiffness at the LP turbine-end bearing
also was varied. Significant reduction
of subsynchronous response was not ob-
served.
Removal of the instability was accom-
plished by refitting the test rig with a
spring cage in place of the intershaft
damper. The result of this modification
is shown in Figure 7, an rpm spectrum
map of the simulated HP turbine response
after the refit. The subsynchronous
responses that appeared previously above
8000 rpm are absent. Although a
slightly different speed schedule was
used in this test, many of the synchro-
nous features identified previously can
be seen. The speed schedule modifica-
tion does not affect the validity of the
test. The modified speed schedule was
followed merely to give a better simula-
tion of engine conditions than the
schedule shown in Figure 2.
CONCLUSIONS
Use of a squeeze film damper in an
intershaft application is responsible
for the instability observed during
testing of a two-spool rotor dynamics
simulator. Replacing the intershaft
318
damper with a spring cage removed the
instability. A limlted program of
testing alternate damper clearances and
support stiffnesses revealed no signif-
icant effects on the instability. The
instability was driven by destabilizing
hydrodynamic forces in the intershaft
damper that develop when the system
tends to whirl at a speed less than the
damper spin speed. The instability
appears to be associated with the third
system mode, in which the intershaft
damper is influential.
C Damper Radial Clearance
e Eccentricity
F r Damper Oil Film Radial Force
F t Damper Oil Film Tangential Force
h Damper Oil Film Thickness
P Damper Oil Film Pressure
R Damper Radius
t Time
NOMENCLATURE
Z Damper Axial Coordinate
Eccentricity Ratio
8 Angle From Line of Centers, In Direc-
tion of Rotation
Oil Kinematic Viscosity
$ Journal Whirl Rate
W 1 Inner Journal Spin Speed
W 2 Outer Journal Spin Speed
W L LP Rotor Spin Speed
REFERENCES
Hibner, D.H., Kirk, R.H., and Buono, D.F., "Analytical and Experimental Investi-
gation of Intershaft Squeeze Film Dampers, Part I-Demonstration of Instability,"
Journal of Engineering for Power, Trans. ASME, Vol. 99, Series A, No. i, Jan.
1977, pp. 47-52.
LP ROTOR , HP ROTOR _
Figure i.
ERSHAFT BEARING
Rotor Dynamics Simulator Schematic.
319
O_
UJ_
ILl
O_
O9
U_
-I- o
n o
"r
--J
o
0.00
/
/
fl
//z
/
5.00
REFERENCE SPEED
0,00
KRPM )
/
/
15.00
Figure 2. Rotor Dynamics Simulator Speed Schedule.
System Includes Intershaft Squeeze Film Damper.
320
15
oRIG|NAL )v_,_r_ _
pooA
A
v
V3
(,/,)
5 s
1 i I
100 200 300 400 500
FREQUENCY (Hz)
Figure 3. RPM Spectrum Map for Simulated HP Turbine.
15
10
v
UJ
u9
5 5
Figure 4.
100 200 300 400 500
FREQUENCY (Hz)
RPM Spectrum Map with Predicted Whirl Map Superimposed.
321
°t
q
T"
_T ''| c
, i u
Figure 5. Deformed Shapes for System Modes 1 - 3.
322
ORIGINAL PAGE IS
POOR QuALrrY
LINE OF 01L
CENTERS FILM
OUTER
JOURNAL
Ft
Figure 6. Intershaft Squeeze Film Damper Geometry.
I0
Figure 7.
[ [ t I
i00 200 300 400 500
FREQUENCY (Hz)
RPM Spectrum Map for Simulated HP Turbine.
Spring Cage Replaces Intershaft Squeeze Film Damper.
323


Instability of an intershaft squeeze film damper in a two-spool rotor dynamics simulator

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