Gas turbine engine motor assembly

A gas turbine engine rotor assembly has an inertial damper loosely surrounding a portion of the main rotor shaft. The damper is axially compressed between facing abutment surfaces provided on a nut and bevel gear. The damper has two axially spaced rows of circumferentially spaced slots, with the interrupting web portions of each row being circumferentially staggered with respect to one another. The damper is configured and dimensioned so that the axial force exerted by the damper on the nut and bevel gear, the polar moment of inertia of the damper, and the coefficient of friction between the damper end faces and the abutment surfaces, cooperate to effectively damp the excitation of the fifth-stage blades in response to application of a source of vibrations at a particular frequency to the bevel gear.

BACKGROUND OF THE INVENTION 
The present invention relates generally to gas turbine engines, and, more 
specifically, to an improved gas turbine engine rotor assembly which 
incorporates an inertial damper for damping selected vibrations 
transmitted from one portion of the rotor assembly to another portion 
thereof. 
Gas turbine engines have a rotor assembly mounted for rotation within a 
stator. The rotor assembly typically has a plurality of axially spaced 
rows of circumferentially spaced blades mounted on a main rotor shaft. In 
the compressor stage of such rotor assembly, the blades of each row 
progressively decrease in size in the direction of flow, that is, the 
blades of each row are smaller than the blades of the immediately adjacent 
upstream row. 
A bevel gear is typically mounted on the main rotor shaft, and is 
continuously engaged with a cooperative auxiliary bevel gear located at 
the distal end of an auxiliary shaft. This auxiliary shaft is commonly 
known as a power-take-off shaft which often has its longitudinal axis 
arranged so as be perpendicular, or at least oblique, to the axis of the 
main shaft. A starter is operatively coupled to the auxiliary shaft. Thus, 
for example, during engine start-up, the starter causes the auxiliary 
shaft to rotate up to about 7,000 rpm. Such motion is transmitted through 
the auxiliary and main shaft bevel gears, to cause the rotor assembly to 
rotate. After the engine has been started, the powered rotation of the 
main shaft drives the rotation of the auxiliary shaft. In this mode, the 
starter is conventionally disconnected and an alternator which is 
connected to the shaft supplies electrical power to the system of which 
the engine is a part. Hence, the auxiliary shaft remains mechanically 
coupled with the main rotor shaft throughout the entire operating speed 
range of the main shaft which typically extends up to about 46,000 rpm. 
In certain engine configurations, it was observed on strain gauges that the 
fifth-stage compressor blades suffered from excessive vibratory stress 
levels at or near a resonant frequency during engine start-up. It was 
noted that the stress due to resonant frequencies of such blades had an 
apparent relationship to the number of teeth in the rotor assembly bevel 
gear. More particularly, it was observed that the fifth-stage blades were 
being excited at "38/rev" (i.e., 38 times the rotor speed), which number 
coincided with the number of teeth in the rotor assembly bevel gear. From 
this observation, it was deduced that the vibrational excitation of the 
fifth-stage blades was attributable to a source of vibrations transmitted 
from the auxiliary gear to the rotor assembly bevel gear, and through 
torsional vibration of the rotor shaft to the fifth-stage of the 
compressor. 
SUMMARY OF THE INVENTION 
Accordingly, one object of the invention is to provide an improved gas 
turbine engine rotor assembly. 
Another object of the invention is to provide a rotor assembly in which a 
particular row of blades is not vibrationally excited in response to 
application of a source of vibrations to a particular portion of the rotor 
assembly. 
Another object is to provide an improved inertial damper for use in a gas 
turbine engine, which damper is effective to damp the vibrations 
transmitted from a first portion of the rotor assembly to another portion 
thereof. 
Another object is to provide an improved inertial damper which is effective 
to damp selected vibrations transmitted from a rotor bevel gear to the 
fifth-stage row of blades. 
Another object is to provide an improved inertial damper for use in a gas 
turbine engine, which may be easily incorporated into the existing engine 
configuration and physical constraints, with minimal modification of 
existing parts and components therein. 
An improved gas turbine engine rotor assembly including an improved 
inertial damper for use therein, is disclosed. The improved damper is 
adapted to be mounted on the rotor assembly in such a manner as to exert a 
predetermined force on one abutment surface provided on the rotor 
assembly, and on another abutment surface provided on the bevel gear. The 
magnitude of this preload force, the polar moment of inertia of the 
damper, and the coefficient of friction between the damper and the two 
abutment surfaces are selected such that vibrational excitation 
transmitted through the bevel gear will be effectively damped, and will 
not cause another portion of the rotor assembly (e.g., the blades of the 
fifth-stage) to become vibrationally excited at or near a resonant 
frequency, all with an object of obtaining a preferred high-cycle fatigue 
life of such blades. At the same time, such inertial damper may be readily 
incorporated into an existing engine design and configuration with a 
minimum of modification to other structures. Thus, effective damping is 
provided at minimal cost and effort. 
In a preferred embodiment, the improved damper is in the form of a 
cylindrical tube. The opposite end faces of this tube are adapted to bear 
against opposed and facing abutment surfaces on the rotor and bevel gear. 
The tube has at least two axially spaced rows of circumferentially spaced 
slots. The slots of each row are separated by web-like interruptions, and 
the interruptions of the various rows are circumferentially staggered with 
respect to one another. Thus, the damper is made to be axially flexible, 
and is axially compressed by a predetermined distance to cause the damper 
to exert the desired preload force on the abutment surfaces.

DETAILED DESCRIPTION 
Referring initially to FIG. 1, a portion of an improved gas turbine engine 
in accordance with a preferred embodiment of the invention, is generally 
indicated at 10. 
Engine 10 is shown as broadly including an elongated stator 11 and a rotor 
assembly 12 therewithin. The stator includes an outer casing 13 and an 
inner casing 14, which define therebetween an annular flow passageway 
which extends rearwardly from a forwardmost inlet 15 to an intermediate 
compressor portion, generally indicated at 16. A plurality of radially 
extending circumferentially spaced guide vanes, severally indicated at 18, 
join the stator inner and outer casings proximate the inlet. 
The rotor assembly 12 is shown as having a main shaft 19 suitably 
journalled on the stator at a forward end by means of an aft bearing 21, 
with an aft end of the shaft 19 being suitably connected to a high 
pressure turbine (not shown) which drives the shaft 19. Extending 
coaxially within the shaft 19 is a power shaft 17 supported at a forward 
end by a forward bearing 20, with an aft end of the shaft 17 being 
suitably connected to a low pressure turbine (not shown) which drives the 
shaft 17 for powering a propeller or helicopter rotor, for example. The 
outer race of forward bearing 20 is secured to the inner casing by a 
plurality of circumferentially spaced radially extending struts, one of 
which is indicated at 22. The outer race of bearing 21 is suitably secured 
to the rearward marginal end portion of the inner casing. 
The compressor portion 16 includes five axially spaced rows of 
circumferentially spaced rotor blades, which extend outwardly from the 
main shaft 19 into the annular passageway defined between the disk in 
which the roots of the blades are mounted, and the outer casing 13. The 
five rows of blades are severally indicated at 16A, 16B, 16C, 16D and 16E, 
respectively. Disposed upstream of each blade row is a row of stator 
vanes, with each vane/blade row combination being referred to as a 
"stage". The physical size of the blades in each row progressively 
decreases in the direction of flow through the engine, that is, the 
first-stage blades 16A are larger than the second-stage blades 16B, which 
are larger than the third-stage blades 16C, and so on. A centrifugal 
compressor 23 is also mounted on the main shaft 19 to the rear of the 
fifth-stage row of blades. 
A forwardly facing bevel gear, generally indicated at 24, is mounted on the 
main rotor shaft, just forwardly of bearing 21. This bevel gear 24 is in 
continuous meshed engagement with an auxiliary pinion or bevel gear 25 
secured to the inward marginal end portion of an auxiliary or 
power-take-off ("pto") shaft 26. A starter motor 28, shown as being in 
exploded aligned relation to the engine, is normally mounted on the 
stator, and is in releasable meshing engagement with auxiliary shaft 26. 
During engine start-up, starter 28 is operated to selectively rotate the 
main rotor 19 relative to the stator. However, after the engine has been 
started, starter 28 is disengaged. Other downstream components of the 
improved engine, such as the combustor and high- and low-pressure 
turbines, have been omitted in the interest of clarity. 
The portions of the engine just described, are conventional. However, in 
use, it was observed that the fifth-stage blades 16E exhibited reduced 
high-cycle fatigue life, due to an apparent resonance thereof at or near a 
particular rotor speed. More particularly, it was observed that such 
resonance occurred during start-up at rotor speeds of up to about 7,000 
rpm for the rotor shaft 19. Since the speed range of the rotor is from 
zero to about 46,000 rpm, the rotor passed through a resonance-inducing 
speed each time the engine was started. Strain gauges were then placed on 
the fifth-stage blades in an attempt to identify the source of the 
vibrations which were exciting such blades. In analyzing the data provided 
by these strain gauges, it was noticed that the blade stresses at or near 
the critical rotor speed (i.e., at about 7,000 rpm), had a frequency of 
38/rev. Since bevel gear 24 had thirty-eight teeth, this suggested that 
auxiliary gear 25 was transmitting vibrations to the bevel gear, which 
were causing the fifth-stage blades to vibrate at or near a resonant 
frequency. Accordingly, the improved damper, generally indicated at 29 in 
FIG. 1, was designed and configured so as to be operatively mountable on 
the rotor assembly 12 with a minimum of modifications to the other 
existing structure of the engine. 
Referring now to FIG. 2, the pertinent portion of the engine is shown as 
again including main rotor shaft 19, bearing 21, bevel gear 24, damper 29, 
a nut 30, the auxiliary shaft 26, and pinion gear 25. Other structure of 
the engine which is collateral to an understanding of the improved damper, 
has been omitted from FIG. 2 in the interest of clarity. 
The auxiliary or pto shaft 26 is shown as being rotatable about axis y--y 
which is perpendicular to a longitudinal centerline rotor axis x--x, and 
has a lower end portion fixedly mounted to pinion gear 25. Auxiliary shaft 
26 is shown as being journalled for rotation in a bearing 31, the outer 
race of which is adapted to be mounted on the engine stator portion. 
Pinion 25 has its downwardly and outwardly facing teeth arranged in 
continuous meshing engagement with the upwardly and leftwardly facing 
teeth of bevel gear 24. 
The illustrated portion of shaft 19 is depicted as being a tubular member 
elongated along horizontal axis x--x, and having an outer surface which 
sequentially includes in pertinent part: an outwardly facing horizontal 
cylindrical surface 32, an externally threaded portion 33, a leftwardly 
facing annular vertical surface 34, an outwardly facing horizontal 
cylindrical surface 35, an outwardly facing surface 36 defining spline 
teeth, and outwardly facing horizontal cylindrical surface 38, a 
leftwardly facing annular vertical surface 39, and an outwardly facing 
horizontal cylindrical surface 40 continuing rightwardly therefrom. The 
other details of shaft 19 are deemed to be collateral to a fundamental 
understanding of the invention, and have not been illustrated. The rotor 
shaft is adapted to rotate about axis x--x at angular speeds of from zero 
to about 46,000 rpm. 
Bearing 21 is of conventional design, and is arranged to surround the shaft 
19 such that the annular vertical right end face 42 of its inner race 43 
is arranged to abut shaft surface 39. The inwardly facing horizontal 
cylindrical surface 44 of the inner race is arranged in closely spaced 
facing relation to the right marginal end portion of surface 38. 
Bevel gear 24 is shown as being a specially configured member surrounding 
the shaft 19, and as having its annular vertical right end face 45 
abutting the left end face 46 of the bearing inner race. The bevel gear 24 
has its upwardly and leftwardly facing teeth, severally indicated at 48, 
in meshed engagement with the downwardly and rightwardly facing teeth 49 
of the auxiliary gear 25. The bevel gear 24 has an outer annular surface, 
which sequentially includes: an upwardly and rightwardly facing frusto 
conical surface 50, another upwardly and rightwardly facing frusto conical 
surface 51, and an outwardly facing horizontal cylindrical surface 52 
continuing rightwardly therefrom to join right end face 45. The various 
portions of the outer surface between surfaces 50,51 and 51,52 are shown 
as being in smooth continuous transition. The inner surface of the bevel 
gear 24 sequentially includes: an inwardly and leftwardly facing frusto 
conical surface 53 extending rightwardly and inwardly from teeth 48, an 
inwardly facing horizontal cylindrical surface 54, a rightwardly facing 
annular vertical surface 55, an inwardly facing horizontal cylindrical 
surface 56, a leftwardly facing annular vertical surface 58, an inwardly 
facing horizontal cylindrical surface 59, and inwardly facing surface 60 
defining spline teeth, and an inwardly facing horizontal cylindrical 
surface 61 continuing rightwardly therefrom to join right end face 42. The 
bevel gear is shown as having a splined connection, indicated at 62, with 
shaft 19, which is defined by the interdigitated spline teeth of surfaces 
36 and 60. Thus, the bevel gear is slipped over the shaft 19, and is moved 
rightwardly relative thereto such that bevel gear right end face 45 abuts 
bearing inner race left end face 46. 
Nut 30 is shown as being a specially configured annular member having an 
annular vertical left end face 63, and annular vertical right end face 64, 
and a stepped outer surface which sequentially includes a polygonal 
surface 65, a rightwardly facing annular vertical abutment surface 66, and 
an outwardly facing horizontal cylindrical surface 68 continuing 
rightwardly therefrom to join right end face 64. The inner surface of the 
nut is shown as having an internally threaded portion 69 extending 
rightwardly from left end face 63, and having a rightwardly and inwardly 
facing frusto conical surface 70 continuing therefrom to join right end 
face 64. The nut is threaded on to shaft threads 33 on the shaft 19, and 
is selectively tightened until nut right end face 64 abuts the inner 
marginal end portion of bevel gear abutment surface 58. 
Referring now to FIGS. 2-5, damper 29 is shown as being a horizontally 
elongated cylindrical tube compressively sandwiched between the bevel gear 
24 and the nut 30. As best shown in FIG. 4, the damper has an annular 
vertical left end face 71, an annular vertical right end face 72, an 
horizontal cylindrical outer surface 73, and a horizontal cylindrical 
inner surface 74. The inner margin of the damper left end face 71 is 
arranged to engage nut abutment surface 66. The damper right end face 72 
is arranged to engage bevel gear abutment surface 58. The damper inner 
surface 74 is spaced from the nut surface 68 to create up to about 2 mils 
(0.002 inches) of maximum diametral clearance to prevent undesired 
unbalance due to offset of the damper 29. A minimum diametral clearance of 
about 1/2 mil (0.0005 inches) is preferred to prevent binding of the 
damper 29 and allow unrestricted circumferential movement. 
As best shown in FIGS. 3-5, the preferred form of the damper 29 is provided 
with two axially spaced rows of circumferentially spaced through slots, 
there being three of such slots in each row. The two rows are preferably 
provided in the central portion of the damper, with the slots of the left 
row being arranged to the left of the axial midpoint of the damper, and 
the slots of the right row being arranged to the right of such midpoint. 
Thus, in the preferred embodiment, the left and right rows of slots are 
centered about such midpoint. The slots of the left row are severally 
indicated at 75, and the webs or interruption therebetween are severally 
indicated at 76. Similarly, the slots of the right row are severally 
indicated at 78, and the webs or interruptions therebetween are severally 
indicated at 79. The slots of each row severally occupy arc distances of 
about 110.degree., with the interrupting web portions occupying arc 
distances of about 10.degree.. The slots and webs of the two rows are 
equally dimensioned and proportioned, but are circumferentially staggered 
with respect to one another, as shown in FIG. 5. If desired, such slots 
may be made by conventional electro discharge machining. The purpose of 
these rows of slots is to cause the damper to be flexible to compression 
in an axial direction. Thus, the damper has a relatively low spring rate 
in the axial direction. 
The damper is mounted on the rotor assembly as shown in FIG. 2, with its 
right end face 72 engaging bevel gear abutment surface 58, and with its 
left end face 71 abutting nut abutment surface 66. The damper is formed to 
have a particular uncompressed axial length, and the nut is configured so 
as to have a predetermined shorter axial dimension between its right end 
face 64 and its abutment surface 66. Hence, when the nut is threaded on to 
the rotor such that the nut end face tightly engages bevel gear surface 
58, the damper will be axially compressed by a known distance (i.e., the 
difference between the undeflected free axial length of the damper and the 
axial distance between nut abutment surface 66 and nut end face 64). The 
spring rate of the damper, coupled with its axial compression by a known 
amount, causes the left and right end faces of the damper to exert a known 
preload force on each of the nut and bevel gear abutment surfaces 66,58, 
respectively. In a specific embodiment, the damper is configured so that 
such preload force is about ten pounds. The preload force is selected to 
provide maximum damping for each application. 
Furthermore, the damper 29 is configured and dimensioned so that the 
magnitude of such preload force, the coefficient of friction between the 
damper end faces and the facing abutment surfaces, and the polar moment of 
inertia of the damper, are such that both end faces 71 and 72 of the 
damper 29 will "skid" or slip relative to the abutment surfaces 58 and 66 
when the rotor is torsionally excited by the bevel gear 24 during 
start-up, thereby to damp the vibrations transmitted from the bevel gear 
through the shaft or other vibrationally-conductive structure to the 
excitable rotor portion, such as the blades of the fifth-row. 
More specifically, the improved rotor assembly 12 in accordance with the 
exemplary, preferred embodiment of the invention includes the damper 29, 
which dissipates by friction vibrational excitation energy and 
predeterminedly includes sufficient axial flexibility for making the 
damper 29 practical without compromising structural integrity when being 
operated at high rotational speed up to about 46,000 rpm. 
Damping is accomplished by selecting the polar moment of inertia of the 
damper 29 and the frictional force between the damper 29 and the abutment 
surfaces 58 and 66 so as to provide for deliberate relative torsional 
slippage therebetween when the bevel gear 24 is vibrationally excited at a 
predetermined frequency, such as that occurring due to the 38/rev 
excitation of the bevel gear 24. 
At such a condition of excitation, the bevel gear 24, shaft 19 and nut 30 
will torsionally vibrate or oscillate, and the damper 29 will tend to 
resist that vibration due to its polar moment of inertia. If the 
frictional forces between the damper 29 and the abutment surfaces 58 and 
66 are relatively low, the damper 29 will tend to remain at the nominal 
speed of rotation of the shaft 19 and not follow the vibrational motion of 
the bevel gear 24. If the frictional forces are relatively high, the 
damper 29 will follow or vibrate with the bevel gear 24, and effective 
damping will not occur. If a moderate amount of friction is provided, that 
friction will dissipate torsional vibration energy due to the torsional 
slippage between the damper 29 and the abutment surfaces 58 and 66. 
Accordingly, the polar moment of inertia and the axial preload of the 
damper 29 are selected in a trade-off to insure that the resistance to 
torsion forces due to the polar moment of inertia become greater than the 
frictional forces due to the axial preload of the damper 29 at the 
predetermined condition causing the torsional excitation. At such a 
condition, the contact between the damper 29 and the abutment surfaces 58 
and 66 is such as to allow torsional slippage which frictionally 
dissipates excitational energy. 
In a preferred embodiment, the polar moment of inertia had a value of 
3.9.times.(10).sup.-4 inch-pounds-second.sup.2 and the axial preload force 
was about 10 pounds. 
An accurate amount of axial preload is required on the damper 29 to ensure 
slippage of the damper 29 at a predetermined level of torsional vibration. 
Although the damper 29 could be imperforate, it would also be relatively 
axially stiff and the dimensional tolerances of the damper 29 relative to 
the abutment surfaces 58 and 66 required to insure a predetermined axial 
preload would be substantially small and not practical. The axial preload 
could also be generated by use of an additional spring biasing means, but 
would therefore be relatively more complex and would have to be 
additionally effective at the relatively high rotational speeds. 
In accordance with the preferred embodiment, the slots 75 and 78 are 
provided to create a member (i.e., the damper 29) which not only provides 
for frictional damping, but also inherently provides relative axial 
flexibility due to the slots 75 and 78. The axial flexibility allows for 
sufficient axial compression of the damper 29 between the abutment 
surfaces to insure an accurate and relatively low amount of axial preload 
force. For example, the axial preload force of 10 pounds in the preferred 
embodiment was obtained by axially compressing the damper 29 between the 
abutment surfaces 58 and 66 a distance of 10 mils (i.e., 0.010 inches). 
Since the slots 75 and 78 are oriented in the circumferential direction, 
undesirable stress in the damper is not generated under the high 
centrifugal loads occurring at the high rotational speeds of up to about 
46,000 rpm. And, the damper 29 retains adequate structural integrity and 
maintains a constant preload at all speeds. 
The improved engine rotor assembly and damper are capable of many changes 
and modifications. For example, the damper may be provided with more than 
two axially spaced rows of circumferentially spaced slots, if desired, and 
these may not necessarily be arranged symmetrically with respect to the 
axial midpoint of the damper. The slots of each row may be "saw cut" or 
otherwise formed, as desired. Various spacers (not shown) might be 
provided between the bevel gear 24 and the bearing inner race 43, and/or 
between inner race 43 and the shaft 19. The engaged surfaces of the 
damper, nut and bevel gear (i.e., surfaces 66,71 and surfaces 72,58) may 
be suitably finished (i.e., polished, roughened, etc.) so as to determine 
a selected coefficient of friction therebetween. 
While there have been described herein what are considered to be preferred 
embodiments of the present invention, other modifications of the invention 
shall be apparent to those skilled in the art from the teachings herein, 
and it is, therefore, desired to secure in the appended claims all such 
modifications as fall within the true spirit and scope of the invention. 
Accordingly, what is desired to be secured by Letters Patent of the United 
States is the invention as defined and differentiated by the following 
claims.