U.S. patent application number 09/844741 was filed with the patent office on 2002-10-24 for high thrust turbocharger rotor with ball bearings.
Invention is credited to Duve, Eric J., Heilenbach, James William, Panos, Jean B..
Application Number | 20020155009 09/844741 |
Document ID | / |
Family ID | 25293510 |
Filed Date | 2002-10-24 |
United States Patent
Application |
20020155009 |
Kind Code |
A1 |
Panos, Jean B. ; et
al. |
October 24, 2002 |
HIGH THRUST TURBOCHARGER ROTOR WITH BALL BEARINGS
Abstract
An improved turbocharger for railway locomotive sized engines
includes, in a preferred embodiment, first and second axially
spaced ball bearings supporting a rotor with the first bearing
being a hybrid ceramic ball bearing mounted to accept both radial
and axial loads acting on the shaft at the compressor end. The
first bearing is mounted on a reduced diameter portion of the
shaft, providing reduced bearing diameter to acceptably limit
centrifugal loading of ceramic balls in the bearing against a
surrounding bearing race. The first bearing has dual rows of
ceramic ball bearings mounted to share all axial thrust loads on
the shaft. The second bearing is also a ball bearing. Lubrication
of the bearings is preferably by direct impingement on the inner
race to minimize oil churning causing heating and power loss.
Additional features and advantages are disclosed.
Inventors: |
Panos, Jean B.; (York,
PA) ; Heilenbach, James William; (Riverside, IL)
; Duve, Eric J.; (Riverside, IL) |
Correspondence
Address: |
General Motors Corporation
Legal Staff
Mail Code 482-C23-B21
P.O. Box 300
Detroit
MI
48265-3000
US
|
Family ID: |
25293510 |
Appl. No.: |
09/844741 |
Filed: |
April 24, 2001 |
Current U.S.
Class: |
417/407 |
Current CPC
Class: |
F05D 2300/21 20130101;
F16C 25/083 20130101; F01D 25/166 20130101; F01D 25/164 20130101;
F16C 27/045 20130101; F01D 25/16 20130101; F02C 7/06 20130101; F16C
2360/24 20130101; F16C 19/56 20130101; F05D 2220/40 20130101; F01D
25/18 20130101 |
Class at
Publication: |
417/407 |
International
Class: |
F04B 017/00; F04B
035/00 |
Claims
1. A turbocharger having a housing carrying a rotor including an
axial flow turbine wheel and a radial flow compressor wheel
supported at opposite ends of a shaft carried in the housing on oil
lubricated first and second bearings spaced axially adjacent
compressor and turbine ends respectively of the shaft and providing
an overhung rotor mounting with axial thrust loading normally
applied to the shaft from both wheels in the same direction from
the turbine toward the compressor, and the improvement wherein:
said first bearing comprises at least one hybrid ceramic ball
bearing mounted to accept both radial and axial loads acting on the
shaft at the compressor end; said first bearing mounted on a
reduced diameter portion of the shaft and providing reduced bearing
diameter to acceptably limit centrifugal loading of ceramic balls
in the bearing against a surrounding bearing race.
2. A turbocharger as in claim 1 wherein said first bearing includes
dual rows of ceramic ball bearings sharing all the primary thrust
loads on the shaft to maintain thrust loading of the bearing within
acceptable limits.
3. A turbocharger as in claim 1 wherein said second bearing is a
single row ball bearing that primarily carries only radial loads
and is mounted on a second reduced diameter portion of the shaft to
acceptably limit centrifugal loading of balls in the second bearing
against a surrounding bearing race.
4. A turbocharger as in claim 3 wherein said second bearing is a
hybrid ceramic ball bearing.
5. A turbocharger as in claim 1 wherein said first bearing is
lubricated with a controlled amount of oil applied by direct
impingement to an interface of the balls with an inner race of the
bearing to limit oil churning.
6. A turbocharger as in claim 1 wherein said rotor includes an
axial fastener extending through the compressor and the shaft to
engage the turbine and hold components of the rotor together with
an initial axial load, the fastener being axially yieldable to
maintain the axial load on the rotor relatively constant under
varying temperature operating conditions of the rotor.
7. A turbocharger as in claim 1 wherein a portion of the shaft
between the first and second bearings has a greater outside
diameter than said reduced diameter to maintain adequate bending
stiffness in the overhung rotor.
8. A turbocharger as in claim 4 wherein said second bearing is
slidable in the housing to direct primary thrust loading of the
rotor to the first bearing.
9. A turbocharger as in claim 8 wherein said second bearing is
mounted with a squeeze film damper to minimize whirl of the turbine
end of the rotor.
10. A turbocharger as in claim 1 wherein polygon projections in
mating recesses comprise drive elements between the shaft and the
connected turbine and compressor wheels.
Description
TECHNICAL FIELD
[0001] This invention relates to engine turbochargers and
particularly to a novel ball bearing mounting of a high thrust
turbocharger rotor.
BACKGROUND OF THE INVENTION
[0002] A turbocharger for a medium speed diesel engine, adaptable
for use in railway road locomotives and other applications, has a
rotor with a radial flow compressor wheel or impeller and an axial
flow turbine wheel or turbine, unlike typical automotive
turbochargers. The wheels are carried at opposite ends of a
connecting shaft supported at two spaced bearing locations with the
wheels overhung. This configuration is known as a flexible rotor,
since it will operate above its first, and possibly second,
critical speeds. It can therefore be subject to rotor dynamic
conditions such as whirl and synchronous vibration.
[0003] High thrust loads are created by the difference in air
pressures across the turbine and compressor wheels. These loads can
be quite large due to the relatively large radial area of the
wheels. The net thrust loads on the wheels are in the same
direction, creating a high overall thrust on the rotor. The radial
load due to the static weight of the rotor is comparatively
small.
[0004] Turbocharger design can include the use of sealing devices
at the rim of turbine wheel to help control pressure on the face of
the turbine wheel inboard of the blades. This is feasible because
the high temperature turbine end materials have more closely
matched thermal expansion coefficients than the aluminum wheel and
ferrous housing materials typical of the compressor end of the
turbocharger. Thus, at the turbine end, a reasonable range of
clearances can be obtained.
[0005] On the upstream end, the aim is to keep the flowpath
pressure off the face of the turbine wheel. This pressure pushes in
the same direction as the thrust on the compressor wheel. On the
downstream end, if the face could be pressurized it would help to
reduce the compressor wheel thrust effect by pushing the other way.
In practice, this is difficult, because the seal must be made very
tight or else an extremely high flow of pressurized air is
required, only to be directly exhausted out of the turbocharger
without being used to do any work.
[0006] Diesel locomotive engines, and turbochargers, may operate
over an extremely large range of conditions, from minus 40 degrees
at startup to the high temperatures and high turbine speeds
experienced in a high altitude tunnel. With aluminum compressor
wheels chosen for low inertia and quick response, their rotating
and static thermal coefficients are poorly matched to the housing
so that sealing the back face of the compressor wheel is not a
currently practical option. Since the compressor pressure ratio is
considerably higher than that of the turbine, a higher pressure
acts over an area about equal to that of the turbine.
[0007] Even with the use of seals where practical, and more so
without them, the high thrust loads acting on the rotor, as well as
the potential for whirl and vibration, have made hydrodynamic fluid
film bearings the universal choice for turbochargers of this type
as compared to the common use of ball bearings in automobile engine
turbochargers. Hydrodynamic fluid film bearings feature high load
capacity, variable stiffness, essentially infinite life if the
fluid film is maintained, and allow large shaft diameter for better
stiffness and lower vibration. However, they require high oil flow
and cause high power losses, which reduce overall efficiency.
[0008] Ball bearings require much lower oil flow and operate with
lower power loss for improved efficiency as well as more consistent
stiffness over the operating range. However, they have lower thrust
load capacity, have finite operating life due to metal fatigue of
the moving parts, and must be limited in diameter so that high
rotating speeds do not put excessive centrifugal loads on the
balls. As a result, ball bearings are not known to have been
successfully applied to turbochargers of the type described as used
in railroad engines and other applications.
SUMMARY OF THE INVENTION
[0009] The present invention provides a turbocharger, adapted for
use in railroad locomotive engines and other applications, combined
with a ball bearing rotor mounting capable of accepting both radial
support loads and axial thrust loads applied to the rotor of a
railroad engine turbocharger.
[0010] In a preferred embodiment, the turbocharger includes a
housing carrying a rotor having an axial flow turbine wheel and a
radial flow compressor wheel. The wheels are supported at opposite
ends of a shaft carried in the housing on oil lubricated first and
second bearings spaced axially adjacent to compressor and turbine
ends respectively of the shaft. The arrangement provides an
overhung rotor mounting with axial thrust loading normally applied
to the shaft from both wheels in the same direction from the
turbine toward the compressor.
[0011] In the improved assembly, the first bearing includes at
least one hybrid ceramic ball bearing mounted to accept both radial
and axial loads acting on the shaft at the compressor end. The
first bearing is mounted on a reduced diameter portion of the
shaft, providing reduced bearing diameter to acceptably limit
centrifugal loading of ceramic balls in the bearing against a
surrounding bearing race.
[0012] Additional features may include a first bearing having dual
rows of ceramic ball bearings mounted to share all axial thrust
loads on the shaft. The second bearing may also be a ball bearing
and, optionally, a hybrid ceramic thrust bearing on a reduced
diameter shaft portion to limit centrifugal loading of the balls in
the bearing. Lubrication of the bearings is preferably by direct
impingement on the inner race to minimize oil churning causing
heating and power loss. The shaft between the bearings preferably
has a greater diameter than at the bearing locations to maintain
adequate bending stiffness in the overhung rotor. The second
bearing may be made slidable in the housing to direct all thrust
loads to the dual row first bearing. A squeeze film damper may
carry the second bearing to minimize whirl at the turbine end of
the rotor. The shaft may be separate from the compressor and
turbine wheels and include a yieldable fastener, such as a stud or
bolt extending through the compressor wheel and the shaft to engage
the turbine wheel and maintain a relatively constant clamping load
on the rotor.
[0013] These and other features and advantages of the invention
will be more fully understood from the following description of
certain specific embodiments of the invention taken together with
the accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
[0014] In the drawings:
[0015] FIG. 1 is a cross-sectional view of an engine turbocharger
having a ball bearing mounted rotor according to the invention;
[0016] FIG. 2 is an enlarged cross-sectional view of the rotor and
bearing mounting portions of the turbocharger of FIG. 1; and
[0017] FIG. 3 is a cross-sectional view through the rotor shaft
toward the end of compressor adapter showing a polygon drive
connection.
DESCRIPTION OF THE PREFERRED EMBODIMENT
[0018] Referring now to the drawings in detail, numeral 10
generally indicates an exhaust driven turbocharger for an engine,
such as a diesel engine intended for use in railway locomotives or
other applications of medium speed diesel engines. Turbocharger 10
includes a rotor 12 carried by a rotor support 14 for rotation on a
longitudinal axis 16 and including a turbine wheel 18 and a
compressor wheel 20. The compressor wheel is enclosed by a
compressor housing assembly 22 including components which are
supported on an axially facing first side 24 of the rotor support
14. An exhaust duct 26 has a compressor end 28 that is mounted on a
second side 30 of the rotor support 14 spaced axially from the
first side 24.
[0019] The exhaust duct 26 is physically positioned between the
rotor support 14 and the turbine wheel 18 to receive exhaust gases
passing through the turbine wheel and carry them to an exhaust
outlet 32. A turbine end 34 of the exhaust duct 26 and an
associated nozzle retainer assembly 35 are separately supported by
an exhaust duct support 36 that is connected with the exhaust duct
26 at the turbine end 34. The exhaust duct support 36 also supports
a turbine inlet scroll 38 which receives exhaust gas from an
associated engine and directs it through a nozzle ring 40 to the
turbine wheel 18 for transferring energy to drive the turbocharger
compressor wheel 20.
[0020] The rotor support 14 includes a pair of laterally spaced
mounting feet 42 which are rigidly connected to an upstanding
mounting portion 44 of the rotor support 14 and are adapted to be
mounted on a rigid base, not shown. The rotor support 14 further
includes a tapering rotor support portion 46 having ball bearings
48, 50 that rotatably support the rotor 12 and are subsequently
further described.
[0021] Referring particularly to FIG. 2, the rotor 12 includes a
shaft 52 extending between and operatively engaging inner ends of
the turbine wheel 18 and the compressor wheel 20. A resilient
fastener in the form of a stud 54 extends through axial openings in
the compressor wheel 20 and shaft 52 and engages a threaded opening
in the turbine wheel 18. A nut 56 on the stud 54 engages a washer
on the outer end of the compressor wheel to clamp the rotor
components together with a desired preload. The stud is resiliently
stretched so that the preload remains relatively constant in spite
of variations in the axial length of the rotor assembly under
operating and stationary conditions.
[0022] In accordance with the invention, the rotor 12 is supported
by first and second axially spaced ball bearings 48, 50,
respectively. The bearings engage reduced diameter mounting
portions 58, 60 at opposite ends of the shaft 52. The mounting
portion diameters are sized to reduce the bearing race diameters to
maintain centrifugal forces on the bearing balls within acceptable
limits. The portions of shaft 52 between the mounting portions are
maintained large to provide a stiff connection between the
compressor and turbine wheels.
[0023] At the compressor end of the shaft, bearing 48 includes dual
rows of hybrid ceramic ball bearings having inner and outer races
62, 64 in axial engagement for transferring thrust loads. The inner
races 62 are clamped by a nut 66 against a shoulder 68 at the inner
end of the mounting portion 58. The outer races 64 are received in
a bore of a bearing housing 70 that is secured in and radially
located by the rotor support portion 46 of the rotor support 14.
The dual row bearing 48 transfers primary thrust loads to a radial
flange 72 of the bearing housing 70. A retainer plate 73 mounted on
the bearing housing 70 traps the outer races 64 in the bearing
housing and limits axial motion during axial thrust reversals.
[0024] An oil feed passage 74 in the bearing housing sprays oil
directly from the flange 72 into the bearing 48 between the inner
and outer races. Excess oil from the bearing drains in part through
a drain passage 76 into an open drain area 78. An oil seal member
80 is radially located by the bearing housing 70 but is axially
located by mounting to the rotor support portion 46. Member 80
cooperates with a seal adapter 82 fixed on a stub of the compressor
wheel 20 to limit oil leakage from the bearing 48 toward the
compressor wheel.
[0025] At the turbine end of the shaft 52, bearing 50 is a single
row bearing having inner and outer races 84, 86. The bearing 50 may
be a conventional or hybrid ceramic type and can be made smaller as
it carries primarily relatively light radial loads. The inner race
84 is secured by a nut 88 against a shoulder 90 at the inner end of
the reduced diameter mounting portion 60. The outer race 86 is
carried in a squeeze film damper (SFD) sleeve 92 that floats in a
SFD housing 94 fixed in the rotor support portion 46. Oil is
supplied to the SFD through a groove in the SFD housing which also
supplies an oil feed passage 96 that delivers oil directly to the
bearing balls between the races 84, 86. A preload spring stack 98,
between a flange of the SFD housing 94 and the SFD sleeve 92,
biases the sleeve and the bearing outer race 86 axially toward the
shoulder 90 to maintain continuous axial load on the balls during
limited axial bearing motion and avoid ball skidding and subsequent
fatigue.
[0026] An adapter 100 on a stub of the turbine wheel cooperates
with a seal member 102 mounted on the rotor support portion 46 to
limit oil leakage toward the turbine wheel. The bearing 50 is
drained directly into the central oil drain area 78 of the rotor
support portion 46.
[0027] Ball bearings used in high speed rotating machines tend to
be life limited by centrifugal forces acting on the bearing balls.
The size of the bearing balls and the diameter of the ball races
are thus important factors in the application of ball bearings to
turbomachinery. Accordingly, ball bearings are commonly used with
small automotive engine turbocharger rotors because the small
diameters of balls and races permit long life with conventional
bearing materials. For the same reasons, ball bearing applications
are not found in large engine turbochargers with large diameter
shafts and heavy thrust loads requiring larger bearings.
[0028] The present invention overcomes these problems by combining
several features that make the application of ball bearings
practical in engines of a size useful in diesel road freight
locomotives and other comparable applications. For example, at
least the larger, thrust carrying bearing 48 is mounted on a
reduced diameter portion 58 at the end of the shaft. This allows
the bearing race diameter to be reduced while the portion of the
shaft between the bearings remains large as is needed for adequate
stiffness. A double row bearing is used if needed to carry the high
thrust loads involved. Also, hybrid ceramic ball bearings are used
in, at least, the high load position. The ceramic balls are lighter
than alloy steel but have high capacity so that the centrifugal
force of the balls is reduced and the fatigue life of the bearings
is extended.
[0029] Because the diameters of the shaft ends are reduced, a more
compact and efficient drive coupling is provided between the shaft
52 and the turbine and compressor wheels 18, 20. P3 polygon shaped
openings 104 are presently preferred to couple the shaft 52 to
mating polygon projections 106 extending from adapters 82, 100,
which are pressed onto the wheels 18, 20 and provide running lands
or surfaces for the labyrinth seals. FIG. 3 is a cross-sectional
view through the shaft 52 toward the end of the adapter 82 showing
the shape of the preferred P3 polygon projection 106 which mates
with a similarly shaped polygon opening 104 in the adapter 82. If
desired the projections could extend from the shaft and mate with
openings formed in the adapters.
[0030] In operation of the turbocharger 10, pressurized exhaust gas
is delivered through the turbine inlet scroll 38 to the turbine
wheel 18 where it imparts energy to the turbine blades to drive the
rotor 12 and is then exhausted at a lower pressure. Higher gas
pressure on the inlet face of the turbine wheel yields an axial
thrust force in the direction of the compressor wheel. The rotating
compressor wheel 20 draws in ambient air moving axially and
exhausts it radially at a higher pressure to the compressor housing
22. The outlet pressure acts against the inner side of the
compressor wheel 20 and yields an additional axial thrust force on
the rotor, adding to the thrust of the turbine wheel 18. These
thrust forces are absorbed fully by the dual row ceramic ball
bearing 48 which carries the thrust loads from the turbine shaft 52
to the bearing housing 70 and thus to the rotor support 14.
[0031] The thrust loads generate forces much higher than the radial
support loads, which are shared between the rotor bearings 48 and
50. Bearing 50 is allowed to move axially with its squeeze film
damper (SFD) sleeve 92 in the SFD housing 94. However, it is
expected to handle small transient thrust loads opposite to the
direction of primary thrust forces. The spring stack 98 biases the
bearing outer race toward the shaft shoulder 90 to maintain a small
axial load on the bearing balls. Thus, bearing 50 carries primarily
radial loads and may be made smaller than bearing 48. The axial
loading of bearing 50 helps to avoid ball skidding which could
adversely impact bearing fatigue life. The squeeze film damper is
applied to counteract so called shaft whirl where the shaft or
turbine wheel tends to orbit if the bearing is too lightly loaded.
However a squeeze film damper may not be required in all
turbocharger applications.
[0032] Lubrication of the bearings by direct impingement of oil at
the ball/race interface together with limiting the amount of oil
delivered and draining excess oil quickly, avoids oil churning,
bearing overheating and failure. The power losses from pumping the
oil and viscous resistance of prior hydrodynamic bearings are
greatly reduced with the ball bearings and oil delivery system of
the disclosed embodiment.
[0033] Advantages of the present invention over turbochargers using
the current bearing technology include, without limitation, reduced
oil consumption and horsepower loss without the need for expensive
dynamic air seals, improved rotor dynamics with the use of bearings
more appropriate for the relatively light radial loads, simplified
shaft seals due to the low oil consumption, and a potentially less
complex oil supply system for the turbocharger.
[0034] While the invention has been described by reference to
certain preferred embodiments, it should be understood that
numerous changes could be made within the spirit and scope of the
inventive concepts described. Accordingly, it is intended that the
invention not be limited to the disclosed embodiments, but that it
have the full scope permitted by the language of the following
claims.
* * * * *