Note: Descriptions are shown in the official language in which they were submitted.
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IMPROVED AUTOMATIC TRANSMISSION AND GEAR TRAIN
Inventors: Scott R. Daleriberg, Theodore V. Kachel
and John W. Creighton
BACKGROUND OF THE INVENTION
Field of the Invention
The present invention relates to automatic transmissions and gear trains. In
particular, the instant
invention relates to an automatic transmission and gear train with improved
resistance to
deterioration with an increased effective life.
Description of related art
Motorized vehicles include a powertrain that is comprised of an engine, mufti-
speed
transmission, and a differential or final drive. The mufti-speed transmission
increases the overall
operating range of the vehicle by permitting the engine to operate through its
torque range a
number of times. The number of forward speed ratios that are available in the
transmission
determines the number of times the engine torque range can be repeated. Early
automatic
transmissions had two speed ranges. This severely limited the overall speed
range of the vehicle
and therefore required a relatively large engine that could produce a wide
speed and torque
range. This resulted in increased engine wear and tear and reduced engine
efficiency. As a
result, manually-shifted transmissions that offered several speed ratios were
the most popular for
use in motor vehicles.
With the advent of three- and four-speed automatic transmissions, the
automatic shifting
(planetary gear) transmission increased in popularity. These transmissions
improved the
operating performance and fuel economy of the vehicle. The increased number of
speed ratios
reduces the step size between ratios and therefore improves the shift quality
of the transmission
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by making the ratio interchanges substantially imperceptible to the operator
under normal vehicle
acceleration.
One aspect of the instant invention is a redesign of an automatic transmission
that has a high
failure rate in its present application. The U.S. Postal Service CUSPS) fleet
has 160,000 Long
Life Vehicles (LLV's) on the road with a service life of 25 or more years.
These vehicles rely on
the 180C and/or the 70084 transmission. There has been a long felt need to
design a total
transmission system specifically for the vehicle driving conditions of this
fleet that would
provide durability, longevity and cost savings in the field. High vehicle down
time due to
inherent original transmission design problems have been plaguing this fleet
for many years.
During this time, the USPS has utilized local and regional rebuilders to
supply replacement
products to keep their fleet on the road at great cost.
Due to a shortage of quality replacement parts for the 180C transmission, a
70084 transmission
was introduced in the late '90's to replace the 180C. The 70084 transmission
is a GM unit that
has been in the marketplace for many years. This unit was not designed for the
Postal LLV
application and had to be modified to adapt to the vehicle. This kit included
a driveshaft, shift
linkage and torque converter. The lack of controlled testing and evaluation of
the 70084
transmission resulted high failure rates in the field. Thus, there is a clear
need for a transmission
that is an improvement over the 180C and 70084 transmissions.
The 1800 transmission has problems in five areas:
1. Planetary gear wear or tooth breakage.
2. Valve body malfunction. (2-3 shift valve sleeve sticks)
3. Torque converter fails.
4. Main pump fails.
5. Low band fails.
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The nature of the use of the 180C transmission by the Postal Service results
in constant shifting
from first gear to second gear and back down to first gear. This shifting
sequence occurs
hundreds of times per vehicle in one day of service. The 1800 transmission was
designed to
shift first to second and then to third. The converter clutch would then be
applied and the driver
and vehicle would proceed to its destination. Thus, most of the transmission's
useful service
would be spent in third gear. The 180C transmission as used by the Postal
Service, and in other
applications which require a great deal of stop and go driving, does not see
such ideal usage.
After repeated first to second and then second to first shifts the pinion
gears within the 180C gear
train begin to span. The spalling of the gears results in very small yet very
hard pieces of steel
flaking off the pressure angle surface of each pinion gear. These pieces of
steel are washed away
from the pinion gears by the oil that lubricates and cools these same gears;
thus the pieces of
hard steel are suspended in the oil. As the contaminated oil is pumped to all
critical areas of the
transmission, these pieces of metal act as an abrasive to all parts they
contact. The abrasive
action of the steel particles wears the surfaces of all mating and moving
parts. As gear
deterioration progresses the hard steel particles become more concentrated in
the oil and become
lodged in the valves of the valve body. This results in the transmission not
shifting back to first
gear, but instead using second as the lowest gear. This phenomenon was
confirmed in
transmission destructive tests. In order to continue the test, the valve body
had to be replaced for
the transmission to operate in a normal manner.
Other affected parts are the torque converter and the main pump assembly. It
is difficult to
determine which of these assemblies fail first since they are connected and
the contaminated oil
is pumped from one to the other. Observed deterioration includes the converter
hub wearing
where it is supported by the pump bushing, converter internal bearing
deterioration, pump outer
and inner gear wear, and pressure regulator valve malfunction. All of these
malfunctions are the
direct result of the suspended metal in the oil.
This deterioration can lead to main pump malfunction and oil loss between the
converter hub and
the pump. The loss of shifting control and pump pressure results in the rise
of transmission
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temperature, with temperatures of three hundred degrees or more conceivable.
Under these
conditions the seals that hold the oil pressure that applies the clutches and
bands begin to harden
and allow for oil leakage, thus also allowing the clutches and band to slip
under load, which
leads eventually to clutch and band failure. At this point complete
transmission failure occurs.
The vehicle will not move.
The conventional planetary gear system in the 1800 transmission includes a sun
gear and a ring
gear with pinion gears engaged between the sun and ring gears. The pinion
gears are connected
to and carned by a pinion Garner assembly. The pinion Garner assembly includes
a pinion Garner
body and pinion shafts, the pinion shafts being fixed to the pinion Garner
body. The 1800
utilizes a three-pinion arrangement of compound and plain gears (six gears
total, three compound
and three plain). As stated above, this transmission fails because of the
nature of the application.
The repeated shifts between first and second gear causes spalling of the
pinion gears. The
solution to the above described problems is to prevent this spalling. The
instant invention
comprises a system that would substantially extend the life of the average
1800 transmission.
This newly rebuilt transmission is the 280PSTM transmission.
SUMMARY OF THE INVENTION
One embodiment of the instant invention contemplates a unique transmission
gear train.
Another embodiment of the instant invention contemplates an eight gear
compound planetary
gear train with the accompanying front sun gear, rear sun gear and ring gear
of the appropriate
and unique size.
A further embodiment of the instant invention contemplates a complete
transmission system that
will perform at least about thirty three percent longer than current design or
more.
Another embodiment of the instant invention contemplates a 4 compound pinion
gear train.
An additional embodiment of the instant invention contemplates a complete
transmission system
having eight pinion gears, four of which are compound, four of which are
single.
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An additional embodiment of the instant invention contemplates a transmission
which is at least
about 3 decibels quieter than current transmissions.
BRIEF DESCRIPTION OF THE FIGURES
For a fuller understanding of the nature and desired objects of the present
invention, reference is
made to the following detailed description taken in conjunction with the
accompanying figures
wherein:
Figure 1 illustrates a compound pinion gear.
Figure 2 illustrates a long pinion pin.
Figure 3 illustrates a plain pinion gear.
Figure 4 illustrates a short pinion pin.
Figure 5 illustrates an output shaft with flange.
Figure 6 illustrates a top shell.
Figure 7 illustrates a lower shell.
Figure 8 illustrates a front sun gear.
Figure 9 illustrates a front sun gear hub.
Figure 10 illustrates front sun gear hub forging.
Figure 11 illustrates a sun gear and hub assembly.
Figure 12 illustrates a rear sun gear.
Figure 13 illustrates an assembled rear sun gear.
Figure 14- illustrates a rear band drum.
Figure 15 illustrates a ring gear.
Figure 16 illustrates a ring gear forging.
Figure 17 illustrates a lock plate.
Figure 18 illustrates a small washer.
Figure 19 illustrates a large washer.
Figure 20 illustrates a needle sleeve.
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Figure 21 illustrates a plug.
Figure 22 illustrates a needle roller.
Figure 23 illustrates an output flange casting.
Figure 24 illustrates an assembled body.
Figure 25 illustrates a complete assembly.
Figure 26 illustrates a complete planetary gear kit.
Figure 27 illustrates the 280 transmission pump.
Figure 28 illustrates the 280 bell housing geometry.
Figure 29 illustrates the 280 2/3 shift valve.
Figure 30 illustrates the 280PS flanged converter hub.
Figure 31 illustrates the specifications of the internal helical gear.
Figure 32 illustrates the specifications of the external helical gear.
Figure 33 illustrates a virtual gear mesh representation of the initial 15T to
26T design.
Figure 34 illustrates a virtual gear mesh representation of the revised 15T to
26T design.
Figure 35 illustrates a virtual gear mesh representation of the initial 15T to
20T design.
Figure 36 illustrates a virtual gear mesh representation of the revised 15T to
20T design.
Figure 37 illustrates a virtual gear mesh representation of the initial 34T to
20T design.
Figure 38 illustrates a virtual gear mesh representation of the revised 34T to
20T design.
Figure 39 illustrates a virtual gear mesh representation of the initial 20T to
ring gear design.
Figure 40 illustrates a virtual gear mesh representation of the revised 20T to
ring gear design.
Figure 41 illustrates the cycles of the automated test protocol.
Figure 42 illustrates a comparison of the durability of the 180c transmission
and the 280PS
transmission.
Figure 43 illustrates a comparison of the gear wear of the 180c transmission
and the 280PS
transmission.
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DETAILED DESCRIPTION OF THE INVENTION
For simplicity and illustrative purposes, the principles of the present
invention are described by
referring to various exemplary embodiments thereof. Although the preferred
embodiments of
the invention are particularly disclosed herein, one of ordinary skill in the
art will readily
recognize that the same principles are equally applicable to, and can be
implicated in other
compositions and methods, and that any such variation would be within such
modifications that
do not part from the scope of the present invention. Before explaining the
disclosed
embodiments of the present invention in detail, it is to be understood that
the invention is not
limited in its application to the details of any particular embodiment shown,
since of course the
invention is capable of other embodiments. The terminology used herein is for
the purpose of
description and not of limitation. Further, although certain methods are
described with reference
to certain steps that are presented herein in certain order, in many
instances, these steps may be
performed in any order as may be appreciated by one skilled in the art, and
the methods are not
limited to the particular arrangement of steps disclosed herein.
The current state of the art, the 180C transmission, utilizes a six pinion
compound planetary gear
train. The instant invention is directed to an eight gear compound planetary
gear train with the
accompanying front sun gear, rear sun gear and ring gear of the appropriate
and unique size.
Regardless of what engine is used in front of any transmission the total
torque produced by that
engine must be transferred to the ground by the transmission and associated
parts. Each pinion
gear transfers its share of the torque. Thus, if you had three pinion gears in
the carrier each
pinion would transfer thirty three percent of the torque generated at any one
time. If a fourth
pinion gear is added to the carrier the amount of torque transferred by each
gear is reduced to
twenty five percent. The importance of this change reveals itself in the
affective life of the four
pinion gears collectively. By reducing the load each gear must transfer you
increase the number
of times the gear can transfer the load. In the case of the transmission in
accordance with the
instant invention the affective life of the gear train is increased by thirty
three percent and the
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beginning of gear spalling is delayed by the same amount of time. One can then
conclude that
all other parts that make up the transmission and are cooled and lubricated by
the same oil would
experience a similar increase in effective performance. The longer the system
is kept free of
harmful contamination the longer the system will properly function.
Through an extensive process of root cause analysis, five areas of deficiency
were identified
within the 180C transmission; deficiencies in the inner diameter of the
hydraulic pump, one
deficiency in the alignment of the bell housing, one deficiency was found in
the low servo apply
circuit, one deficiency was detected in the 2/3 shift valve in the valve body,
and one deficiency
was found in the low rear band. These areas of deficiency are listed below and
are addressed in
the subsequent like number sections.
1. The outer pump gear moves out of alignment with the inner pump gear. (metal
generated by the pump)
2. The inner gear contacts the inner diameter of the pump crescent. (noise is
created)
3. The oil pressure developed by the pump is not consistent.
4. _The amount of oil delivered to the transmission lube circuit and torque
converter
feed circuit is not consistent. (not enough oil to reliably cool and lubricate
the gear train
and not enough oil to cool and lubricate the torque converter under all
conditions)
5. The bushing bore center of the bell housing that aligns the torque
converter and
the transmission pump assembly is out of alignment.
6. The aluminum two to three (2/3) shift valve sleeve wears resulting in
erratic and
unpredictable operation of the 2/3 shift valve.
7. The low band does not h~ld the rear band drum and sun-gear under all
conditions.
1. Metal generated by the transmission pump. (Outer gear moves out of
alignment with the inner
gear.)
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Metal particles are generated by the rotary motion of the inner and outer gear
that creates the
volume of oil that results in the necessary oil pressure to operate and
sustain the transmission. In
order to prevent the creation of metal particles from the oil pump, the inner
and outer gear must
be kept in their proper position so the motion of the outer gear does not
improperly mesh with
the inner gear. The inner gear is held in place by the converter hub and the
converter hub
bushing that is located in the bell housing. If either of these parts are not
straight (greater than
.002" TIR) the inner gear will cause a problem. The problem of inner gear
support is discussed
as part of the proper remanufacture of the torque converter and the bell
housing that follows.
The support of the outer gear is another problem. The natural action of the
two gears is to move
away from one another in the area of the pressure port. Thus, the outer gear
has a tendency to
move to the left and the inner gear moves to the right. The outer and inner
gears are constructed
of powder metal and the pump housing is constructed of iron. The outer
diameter of the outer
gear is in constant contact with the pump housing. ~ver time, the outer
diameter of the outer
pump gear begins to wear into the pump housing. This action creates metal
particles in much the
same manner as the pinion gears. In order to prevent this, the pump used in
the 280PSTM
transmission system has a SAE 660 bronze sleeve inserted in the housing in the
area of the outer
gear. The bronze sleeve is machined to the proper clearance for the outer gear
and forms a
perfect wear surface that will assure long and sustained performance.
2. Noise created by inner gear contacting inner diameter of pump crescent.
The tips of the inner pump gear teeth contact the leading edge of the inner
diameter of the pump
crescent as the bell housing bushing wears. When this undesirable interference
occurs between
inner pump gear teeth and the pump crescent, noise and small metal particles
are generated. The
noise is unacceptable from an operational perspective and the metal that is
generated as a result
of the interference is detrimental to the longevity of the transmission. Both
of the above
mentioned problems are eliminated by modifying the geometry of the inner pump
crescent. The
pump of the transmission of the instant invention has the pump crescent inner
diameter changed
to remove both the noise and eliminates the creation of metal particles
generated by the inner
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pump gear contacting the crescent inner diameter. The new geometry of the
inner crescent
creates .060"clearance between the inner gear and the leading edge of the
inner pump crescent
(initial point of contact) and decreases to .000" in a consistent arc 60 to 65
degrees from the
initial point of generation. In so doing the tips of the teeth of the inner
gear can not contact the
leading edge of the inner pump crescent yet by design will still contain the
oil within the inner
pump crescent which maintains all the previous oil pumping capacity that the
pump was
designed to create. This new design thus eliminates both the noise and the
generation of metal
particles, contributing to the overall longevity of the transmission. See
Figure 27.
3. Inconsistent oil pressure.
The oil pressure generated by the pump is regulated by the pressure regulator
valve which is
preloaded by the pressure regulator spring. The combination of these two parts
working in
conjunction with the vacuum modulator (which senses engine load) and the
transmission
governor (which senses vehicle speed) creates the appropriate oil pressure
required to hold all
clutches from slipping and produce oil to lubricate and cool the total
transmission during
operation. The USPS's unique application of this transmission requires the
lubrication oil to be
regulated in the first and second gear ranges, where it spends most of its
operating life, in such a
way as to supply oil to lubricate the gear train and prevent the deterioration
of meshing
components. It was observed during root cause analysis that lubrication oil
pressures varied
greatly in the transmission. The cause was determined to be the pressure
regulator spring
creating significantly different preload pressures on the pressure regulating
valve. This resulted
in either a surplus of lubrication oil or a deficiency of lubrication oil. The
new design pressure
regulator spring incorporated in the 280PSTM pump uniformly preloads the
pressure regulator
valve thus assuring correct and uniform oil pressure which maintains correct
clutch pressure and
lubrication oil flow.
4. Inconsistent lubrication oil flow to the gear train and torque converter.
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It was observed through root cause analysis that even with consistent pressure
regulation of the
transmission oil pump that under certain extreme conditions the oil supplied
to the transmission
gear train and the torque converter was severely reduced or cut off
completely. Both of these
conditions created high wear conditions which reduced the life of the
transmission. In order to
prevent the reduction or loss of lubrication oil, a new method of supplying
oil to the critical
lubrication circuit had to be created that would supply lubrication oil to the
transmission under
any condition. This problem was solved by re-engineering the oil lubrication
circuit to allow a
constant flow of oil form the pressure port to the lubrication port without
the influence of the
pressure regulation system. A passage (equal to .040") was created between the
pressure port of
the pump, before regulation, and the lubrication circuit. In so doing,
lubrication oil was assured
under all conditions guaranteeing adequate lubrication and cooling to critical
components and
thus extending the life of the transmission. See Figure 27.
5. New bell housing design.
It was observed through root cause analysis that the inner gear of the
transmission pump would
move, over time, due to the wear of the bell housing bushing. The wear of the
bushing allowed
the converter hub and the inner pump gear to move in the direction of the
leading edge of the
pump crescent and thus create noise and metal to metal contact. In order to
reduce the possibility
of such an occurrence, the center of the bell housing bushing was moved in the
opposite
direction and in the same distance of the measured wear. The new position of
the bell housing
bushing center is a function of the distance off center in relation to the
alignment dowels located
on the bell housing which are used to align the entire transmission to the
engine. During
transmission operation the pressure created within the transmission oil pump
pushes the inner
gear and the converter hub off center by the amount of clearance that exists
between the
converter hub and the bell housing bushing, minus the oil film created between
the two. When
this misalignment occurs, all the corresponding parts that depended on the
alignment of the bell
housing bushing experienced misalignment by the same amount (.002" to .005").
The bell
housing bushing bore should be on center with the crankshaft center which is
correspondingly on
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center with the transmission pump inner gear. When the pressure created inside
the pump is
exerted on the inner gear and the torque converter hub, the center line of
these parts is forced off
center during operation. When normal wear was added to this equation the
misalignment was
magnified adding to premature transmission failure. The new bell housing
design moves the
center line of the bell housing bushing .0025" off center in the opposite
direction of pressure,
thus when the pump pressure is exerted on the pump gear and converter hub the
entire assembly
of associated parts achieves alignment during operation.
This new alignment design removes the accumulative misalignment that
previously existed
between the tolerance fits required between the converter hub and the inner
pump gear, the
tolerance fit required between the pump stator shaft spline and the torque
converter stator sprag
race, the tolerance fit required between the transmission input shaft spline
and the torque
converter turbine spline and the tolerance fit required between torque
converter nose pilot and
the engine crankshaft torque converter pilot hole . This new design achieves
operational
alignment which reduces the wear on all affected components, thus reducing
metal particle
generation to the greatest extent possible and extending transmission life of
the 280PSTM. See
Figure 28.
6. Valve body (2/3 shift valve and sleeve combination)
If kept free of contamination the valve body will conceivably function
indefinitely. However,
when contamination occurs, some of the components will wear. The point of wear
is the 2/3-
boost sleeve, which is made of aluminum. The transmission of the instant
invention has a new
design boost sleeve installed in this position that is especially designed for
this application. The
geometry of the boost sleeve has been changed to allow more oil to react to
the shifting lands of
the shift valve. In so doing the valve is more sensitive, and produces a more
repeatable point of
shift. The material from which the boost sleeve was made is still aluminum but
the alloy is
changed to 5056 in order to increase wear resistance and reduce the binding
affects that small
particles of metal create on the inner shift valve. The leading edges of the
valve sleeve are very
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sharp by design and function as wipers to the valve spool in much the same way
that any oil seal
protects the inner parts of a mechanism form external dirt contamination. By
increasing the
valve sleeve's resistance to wear the sharp edges are maintained for a longer
period of time
during contamination. This new 2/3 shift valve spool will extend the life of
the transmission
even after metal contamination begins by keeping the small abrasive metal
particles from getting
between the valve and spool thus maintaining proper valve function. See Figure
29.
7. Low servo spring:
The rear clutch band that holds the rear reaction sun-gear and band drum from
turning would,
under extreme conditions, allow the sun-gear and drum assembly to rotate when
it should not.
This condition contributes to premature band friction failure which adversely
affects the
operating life of the transmission. In order to eliminate this malfunction a
new low servo spring
was designed. The low servo spring is applied within the system in a release
configuration thus
assuring that when the rear band holding function is not utilized (third gear)
the band is off and
the band drum rotates freely. The new design low servo spring tension is
reduced. In order for
the low band to be applied the force created by the oil pressure exerted
against the low servo
piston must overcome the spring pressure created to hold it off. By reducing
the amount of
negative pressure created by the low servo spring, the amount of force
remaining to apply the
band was increased resulting in more pressure applied to the low band friction
material. In so
doing, the holding capacity of the low band is increased eliminating the
slippage of the rear band
thus contributing to the extended life of the transmission.
The instant invention also contemplates the remanufacture of the torque
converter of the 180C
transmission. The remanufacture of a torque converter involves the proper
handling and
restoration of five parts:
1. The primary pump.
2. The stator assembly.
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3. The turbine.
4. The clutch and damper assembly.
5. The cover.
The torque converter begins the remanufacturing process by being cut in half.
The closure weld
that was applied when first manufactured must be removed by the use on a lathe
and the pad
plane and z-axis are held to within .002" TIR. The replaceable parts, two
thrust bearings and an
o-ring seal, are removed and discarded. The parts then are washed and sent to
their respective
remanufacturing cells.
The primary pump is comprised of a series of half moon shaped internal fins
that are attached to
steel stamping of the same shape in a radial pattern. An impeller hub is then
welded in the center
of the stamping that holds the fins. The restoration of 280PSTM converter
begins with the
welding of the half moon fins to the impeller stamping and the removal of the
old impeller hub.
The impeller hub removal process consists of removing the complete center of
the primary pump
to which the old impeller hub was welded. This extensive removal of material
is required in
order to replace the bearing and impeller hub support area with a newly
designed one piece
impeller hub and bearing support. This new impeller hub and bearing support
also moves the
heat affected zone caused by the welding attachment process to a larger
diameter eliminating the
distortion of the bearing surface which is the result of contraction of the
three alloys during the
welding process. The installation of a new flanged impeller hub increases the
strength of the
main bearing thrust area between the primary pump and the stator and reduces
the bearing
surface's tendency to distort during operation due to the lateral load exerted
on the hub by the
inner pump gear. At the same time the added flange strength prevents the
impeller hub from
being forced off center by the force of the same inner pump gear. (See
illustration 18) The
primary pump is then machined to restore all datum plains. The primary pump is
then washed a
second time and is ready for assembly.
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The stator assembly is made up of the aluminum housing which contains the fins
that redirect the
oil form the turbine to the primary pump and a one way roller clutch. The
bearing cap is
removed from the primary pump side of the stator and the one way roller clutch
is inspected. If
any contamination is detected the entire assembly is rebuilt. This involves
the removal of the
inner race, the rollers, and the discard of internal springs. The inner race
is tumbled to remove
all traces of the old roller signature and the clutch is reassembled with new
springs and new or
good rollers. The bearing cap is then inspected for wear and discarded is wear
is detected. The
new or good used cap is replaced and the stator is moved to the assembly area.
The turbine is built in much the same way as the primary pump. The points of
wear are the input
shaft spline and the thrust surface. The turbine spline is inspected and if
more than 10% wear is
detected the turbine is discarded or if needed, the old spline hub is removed
and a new spline hub
is welded in the center of the turbine. The C shaped thrust washer is
inspected and if worn
discarded. After final inspection the turbine is moved to the assembly area.
The clutch and damper assembly is made up of a formed steel plate to which the
friction ring is
bonded and a series of springs attached and arranged in a circle about the
turbine spline hub.
The points of failure are broken springs, worn or contaminated friction ring,
and worn seal
surface. If springs are broken the damper assembly is discarded, if seal
surface is damaged the
surface is restored, if the friction ring is worn it is cutoff and new
friction ring is bonded to the
damper assembly. After a final inspection the clutch and damper assembly is
moved to the
assembly area.
The cover is a single piece. This piece contains five critical areas. The
first is the pilot diameter.
This pilot must measure .825"+l-. 002". If the pilot is out of specification
the pilot is welded and
a new diameter is cut to specification. All of the pad surfaces are
reconditioned as well as the
threads contained in each pad. The friction surface is then restored by
machining .010" off the
surface to produce a good friction surface. After final inspection the cover
is moved to the
assembly area.
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The assembly process proceeds as follows, new 'bearings are installed on
either side of the stator
and the stator installed on the primary pump. A new o-ring is installed on the
turbine and the
turbine is installed on the stator. The reconditioned clutch and damper
assembly is assembled on
the turbine hub and the cover is installed on the turbine. The converter then
proceeds to the
closure welder where a new closure weld is installed.
The converter run-out and end play is checked and the finished converter is
balanced to 10 grams
or less. The run-out is checked by placing the impeller hub in a six jaw chuck
and extending the
pilot vertically. An indicator reading is taken form the extended pilot. The
total indicator
reading can not exceed .010" or the converter is rejected.
When all of the above part improvements are incorporated into the 280PSTM
transmission of the
instant invention, the result is a transmission that will function in service
at least thirty percent
longer than the current state of the art transmission, the 1800. All of the
observed areas of
failure have been addressed in this system. The strength of planetary gear
train has been
increased and thus the effective life extended. The pump has had a SAE 660
bronze sleeve
installed to eliminate the outer pump gear wearing against the pump housing.
The pump inner
crescent has been changed to eliminate inner gear interference thus removing
noise and any
possibility of metal generation. The pressure regulation system has been
corrected and the
lubrication circuit changed to allow for full time lubrication of the
converter and gear train. The
bell housing has been changed to allow for in service alignment thus
eliminating wear on all
associated parts. The rear servo spring has been changed to afford greater
rear band holding
capacity. The converter has been remanufactured to the highest standards that
exist within the
art today. The valve body has been remanufactured to remove all wear from
previous life and
has been restored to exceed OEM standards. The friction material on the rear
band and the
intermediate clutch has been changed to eliminate the affects of increased
heat, and finally the oil
that is recommended is by far the accepted standard within the industry for
extending the life of
any transmission.
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Rigorous testing of one embodiment of the instant invention that simulated
long term use of the
transmission identified areas where additional improvements can be made:
a) Bearing pin wear of the 15 tooth short pinion directly in line with the
separation forces
coming from the 20 tooth pinion
b) Loss of alignment with mating 20 tooth long pinion and 26 tooth sun gear
causing tip
loading, wear and pitting
c) Contamination of transmission oil with the worn particles causing
accelerated failure in
other mechanical and hydraulic components.
One surprising result from the evaluation of the components was that there was
relatively little
wear on the 20 tooth pinion bearing pin which is made of the same material and
heat treatment,
although a slightly larger size. It would be expected that the torques in the
20 tooth gear would
be higher, and the pin wear would be larger, which was not the case. This
raised the question as
to the actual torques achieved in each gear in the mesh during the validation
testing.
Calculation of Transmission Torques
An evaluation was conducted to determine the speed and torques that are
achieved in each gear
for each speed condition of the transmission (i.e. 15', 2°d and 3'd
gear shift condition). During this
evaluation, it was discovered that the 1s' gear condition had the highest
torques in the system,
which for this transmission happens to be the most often used shift condition
in the field and
during validation testing.
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Table 1 shows data representing the 1S' Gear Configuration:
Speed and Torque of each Component
With Fixed 34 T Output Sun - 1s' Gear Configuration
+ve Is CW Let N = the speed of S1 when the carrier is held rigidly. Let T~=the
torque of the
output Garner
26 T S' 15 T P' 20 T PZ 74 T Rl CARRIER 34 T SZ
Speed With Si N Sy SIlV .S1N
Fixed
N _ + ,+ O _
Carrier
Pi Pa R' SZ
Adjustment
S
1V
For Fixed I
,+
34 T Output
Durin ls~
Gear
Total Speed NCl+s'~ NS,~S NS,~S +R SIN 0
-P~ + p ~
~ NS,~S
Equation s Z - - '
' - - 2
Value for
N Based
on Sl Speed 1983.333
of
3500 RPM
Achieved
Durin Testin
Total Speed 3,500 -1,921 4,095 2,214 1,517 0
Value
(RPM)
Total Torque
Equation _
@ 100% 's S., S '~'
Efficiency ' 1- 1-~- 1 ~-. T~.
Based 1 ~- Z -
on a Known ,S~ P P,, Rl
output Torque
T
Total Torque162.5 -296.0 138.9 256.9 375.0 -
(ft
lbs) Based
on T
Value of
375 ft lbs
From this data, it is clear that the 15 tooth pinion has a significantly
higher operating torque in 1s'
gear than the 20-tooth pinion. The torques are 214% higher even though the
bearing pin is
slightly smaller on the 15 tooth gear bearing pin than on the 20 tooth gear
bearing pin.
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Component Design Review
The components of the transmission were evaluated to determine how to further
improve
performance in the following areas:
1) Delay the 15 tooth gear bearing pin from wearing
2) Improve the pitting resistance and bending load carrying capacity of the
gear sets to delay
the wear of components.
3) Improve the alignment of gears to delay the wear of components.
4) Improve the load sharing of each of the planet gears to prevent premature
failure of only
one or two components.
Bearing Pin and 15 Tooth Gear Interface
The needle bearing sizing and spacing was evaluated and compared to the needle
bearing sizing
and spacing used on the 20 tooth pinion gear. In terms of load carrying
capacity, the bearing for
the 20 tooth pinion is capable of carrying approximately 2.42 times the load
of the 15 tooth
pinion, because there are 23 needles as compared to 19 on the 15 tooth pinion,
and there is a
double bearing arrangement. For a given size needle roller, the load carrying
capacity can be
conservatively estimated as being directly proportional to the number of
needle rollers.
In addition, it was found that in the maximum material condition, there was a
.529 mm gap,
which could potentially allow for misalignment of the needles during rotation.
This gap is
recommended to be approximately .130 mm.
A redesign of the bearing pin arrangement was conducted with emphasis on
increasing the load
carrying capacity on the 15 tooth pinion bearing as much as possible. The new
design consists of
a bearing pin of diameter 11.170 +/- 0.005 mm, 21 needle rollers (instead of
19), and an inner
bore on the 15 tooth gear of 15.10 +/- 0.005 mm. The maximum gap is now .130
mm. Using
this approach, the load carrying capacity by conservative estimates will
increase by 10.5°70.
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In addition, the benchmarked design washers do not allow for transmission
fluid to flush through
the needle bearing pack. With improved flushing, overheating problems in the
bearing will be
reduced and a low friction coefficient maintained. To improve the flushing of
the fluid, the steel
and phosphor bronze washer inner diameters have been increased to 12.55 +.07 l
- 0 mm.
The material and heat-treating choices were reviewed for the bearing pin to
determine if a more
wear resistant pin can be produced. The current material is SAE 52100 Quench
and Tempered to
60-64 HRC. A change in the allowable surface roughness from 0.4a to 0.2a to
delay the wear of
these components is advantageous.
Gear Pitting Resistance & Bending Strength
All existing meshes are geometrically sound (i.e. no contact in undesirable
areas, minimum
contact ratios above 1.0 (other than ring gear mesh), and sufficient backlash)
for the mesh
condition as described by the drawings. However, for the most part, each of
the meshes has not
been optimized for Pitting Resistance and Bending Strength. A redesign and
optimization
process for the meshes has now been conducted. The methodology that is used to
quantify these
design improvements is to use AGMA's (American Gear Manufactures Association)
calculating
methodology for Pitting Resistance (I) and Bending Resistance (J) Factors.
The new gear geometry is detailed in Figures 19-20. Figures 21-28 illustrate
the initial and
revised mesh conditions in each stage. Tables 2-4 show the initial and final I
and J factors and
the % improvement achieved. In each case, improvements to pitting resistance,
and bending
strength have been achieved by changing the designs to full fillet radii from
sharp fillet radii, and
by optimizing tooth proportions to provide more load sharing.
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Table 2- I & J Factor Comparisons for 15-26 Tooth Mesh
15 Tooth 26 Tooth
Short Sun
Pinion
InitialRevised InitialRevised
Im rovement Im rovement
Bending
Strength
,4566 .5136 12.5 % .3940 .4984 26.5 %
Geometry J
Factor
Pitting
Resistance ,1455 .1721 18.3 % .0805 .0918 14.0 %
Geometry I
Factor
Table 3- I & J Factor Comparisons for 15-20 Tooth Mesh
15 Tooth 20 Tooth
Short Sun
Pinion
InitialRevised InitialRevised
Im rovement Im rovement
Bending
Strength ,4642 .5026 8.3 % .3942 .4928 25.0 %
Geometry J
Factor
Pitting
Resistance ,1322 .1522 15.1 % .0965 .1087 12.6 %
Geometry I
Factor
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Table 4- I & J Factor Comparisons for 34 - 20 Tooth Mesh
34 Tooth 20 Tooth
Sun Sun
InitialRevised Initial Revised
Im rovement Im rovement
Bending
Strength
,4235 .4892 15.5 % .4354 .5501 26.3 %
Geometry J
Factor
Pitting
Resistance
,0839 .0978 16.6 % .1500 .1755 17.0 %
Geometry I
Factor
A review of the existing mesh characteristics for the internal l external
combination reveal that,
geometrically, that the contact ratio is slightly less than the 1.0 that is
the usual practice in gear
design. In order to correct this condition (which will also improve the
strength performance) the
minor diameter of the ring gear are reduced. I and J factors are not
calculated for internal gear
meshes as AGMA does not have an accepted practice to do this.
Review of Alignment Issues
As stated previously, pinion-bearing pin wear causes loss of alignment of the
gear meshes. This
will eventually lead to premature pitting, wear and breakage of gear teeth,
which further
contaminates the transmission oil in the system. In addition, positioning and
size control of the
planets will improve load sharing and further reduce wear.
Improvements in positioning and alignment of the gear meshes improve the life
of the gear train.
Positioning and alignment have been improved by the following methods:
1) Addition of .25 mm of tip relief on the gears by modification of the
shaving cutter tools
2) Increase the crowning on the gear teeth form .006 mm to .03 mm.
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3) Reduce total composite error on the planets to a maximum specification of
.038 mm to
improve consistency and load sharing.
Additional Improvements
Gear hub breakage has been of concern. The hub material is currently 4140,
quenched and
tempered. The material has been changed to 8620, then normalized and
carburised to improve
this condition.
The additional modifications to the transmission of the instant invention can
be summarized as
follows:
1) 15 Tooth Pinion Bearing Assembly is to be resized for an improved fit
2) 15 Tooth Pinion Bearing Pin to have surface roughness controlled to 0.2a
finish
3) All gears to be redesigned to increase the bending strength and pitting
resistance.
Profiles will have full fillet radii, and long addendum tooth profiles.
4) Gears to have tip relief added to reduce alignment concerns.
5) Gears to have .030 mm of crowning added to reduce alignment concerns.
6) Total composite errors on planet gears to be maintained to below .038 mm to
improve
load-sharing capability.
7) Gear hub material to be changed to 8620 normalized and carburised to reduce
hub
cracking concerns.
The transmission of the instant invention must be used as a total replacement
of the previous
design. The individual dimensions of the critical parts, that being the front
sun gear and hub
assembly, the planetary Garner and pinion gears, the rear sun gear and band
drum, and the ring
gear are not compatible with any other critical part of the previous design.
These parts will
however fit in any current market application (180C transmission) when
installed as a complete
assembly. Examples of which would be an Opal Olympia, the GEO Tracker,
Chevette, and
Postal delivery vehicles. The proper installation of these parts in a 180C
transmission requires
only an ordinary level of skill in the art of transmission rebuilding.
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Example 1
The disclosed planetary gear train is compound in nature. The transmission is
composed of eight
pinion gears, four of which are compound, and four that are plain, all eight
gears are assembled
into one carrier utilizing a forward and a rear sun gear, a ring gear, a band
drum and a one way
roller clutch. The four plain pinion gears mesh with the four compound gears
and the forward
sun gear and one way roller clutch. The four compound pinion gears also mesh
with the four
plain pinion gears, the front sun gear, the ring gear and the rear sun gear. A
2.8 reduction is
achieved by applying the rear band and driving the planetary assembly by the
front sun gear
through the one way roller clutch. (First gear) A 1.8 reduction is achieved by
applying the
intermediate clutch, thus driving the planetary by the ring gear and over
running the one way
roller clutch. (Second gear) Third gear is achieved (one to one) by applying
the direct clutch,
and releasing the rear band maintaining the application of the intermediate
clutch thus locking
the ring gear and front sun gear. The unique nature of this transmission lies
in the size of the
pinion gears, the pressure angle to which they are cut and the method of
assembly. This unique
combination produces a gear train that is thirty three percent stronger than
current design and is
four decibels quieter in operation.
Assembly of the gear train:
1. Compound pinion gear (Figure 1)
A. number if 20
teeth
B. module 1.5
C. pressure angle18
D. helix angle 18.16
E. pitch diameter30.0
F. major diameter33.58
G. minor diameter28.64
H. pitch diameter22.50
I. pin diameter 3.0
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J. material SAE 8620
2. Plain pinion(Figure 3)
gear
A. number of 15
teeth
B. module 1.5
C. pressure 18
angle
D. helix angle18.16
E. hand left
F. pitch diameter22.50
G. major diameter28.07
H. minor diameter21.67
I. pin diameter3.0
J. material SAE 8620
3. Out put shaft and
yoke. (Figure 5)
A. number of teeth 27
(shaft)
B. diametrical pitch 24
C. pressure angle 30
D. major diameter 29.30
E. minor diameter 26.82
F. pin diameter 1.5
G. material Shaft ' SAE 4140
H. material flange SAE 1045
I. method of attachmentfriction welding
4. Upper shell (Figure 6)
A. Material SAE 1010
B. Method of attachment Laser welding (As per figure 6)
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5. Lower shell (Figure 7)
A. Material SAE 1010
B. Method of attachment Laser welding (As per figure 6)
6. Assembled body
7. Front sun gear (Figure
8)
A. Number of teeth 26
B. Module 1.5
C. Pressure angle 18
D. Helix angle 18.16
E. Major diameter 44.60
F. Minor diameter 38.00
G. Material SAE 8620
8. Front sun gear hub
(Figure 9)
A. Number of teeth (OD) 12
B. Module 5.25
C. Pressure angle 20
D. Helix angle 0 Spur
E. Hand 0 Spur
F. Major diameter 70.45
G. Minor diameter 66.25
Internal spline data.
A. Number of teeth26
B. Module 1.5
C. Pressure angle 18
D. Major diameter 42.80
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E. Minor diameter 38.65
F. Helix angle 18.20
G. Hand Right
H. Material SAE 4140
9. Sun gear and hub assembly (Figure 11)
10. Rear sun gear (Figure
12)
A. Number of teeth 34
B. Module 1.5
C. Pressure angle 18
D. Helix angle 18.16
E. Hand Left
F. Major diameter 57.42
E. Minor diameter 50.65
G. Material SAE 8620
11. Rear band drum (Figure
14)
A. Number of teeth 34
B. Module 1.5
C. Pressure angle 18
D. Helix angle 18.15
E. Major diameter 51.25
F. Minor diameter 55.34
G. Material FC250
12. Rear sun gear and band drum assembly
13. Ring gear (Figurel5)
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Internal gear specifications
A.Number of teeth 74
B.Diametrical pitch17
C.Pressure angle 18
D.Helix angle 18.16
E.Hand left
F.Material SAE 1045
External specifications
A. Number of teeth 36
B. Module 3.83245
C. Pressure angle 24.33' 36"
D. Major diameter 138.070
E. Minor diameter 132.92
F. Material SAE 1045
The unique angle of the gear pitch, the number and diameter of the pinion
gears, the size of the
front and rear sun gear and the ring gear allow for the transmission of no
less than thirty three
percent more torque. This unique design and combination of parts is the first
of its kind. Current
design of the 180C transmission gear train incorporates six pinion gears.
Three of which are
compound and three that are single. The 280PSTM transmission incorporates
eight pinion gears,
four of which are compound, and four are single. The addition of one more
pinion gears in the
gear train reduces the amount of torque that each gear must transfer by one
third. Thus,
increasing the torque bearing capacity of the gear train by 33%. Current
testing has established
that the eight pinion planetary with the unique gear cut angle, as disclosed
herein, performs
quieter and for a longer period of time than the current six pinion OEM
design. The transmission
will function a minimum of thirty three percent longer than the current 1800
transmission.
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Example 2
The 1800 automatic transmission was tested and compared to an automatic
transmission in
accordance with instant invention, the 280PSTM transmission. The testing
protocol was designed
to reproduce 180C field failures, create a worn or failed gear train, provide
indicators of wear /
failure, provide fluid sample, analyze pressure, temperature, noise, and
vibration. The test
protocol accelerated years of field wear into the test period
A computer algorithm was designed to automate the test. The program ran 11
cycles / hour, each
consisting of four shifts in first and second gear. This produced high torque
and RPM and
created high fluid temperatures
Each cycle was as follows:
Start in 1st Gear
Ramp input speed from 850 RPM to 3500 RPM
Add output torque of 375 lb-ft, hold for 40 seconds
Remove load, but maintain input speed of 3500 RPM
Shift to 2nd Gear
Use vacuum to induce 1 - 2 shift
Apply 375 lb-ft torque, hold for 50 seconds
Release load and drop input speed to 700 RPM
Use vacuum to induce 2 - 1 downshift
Repeat first and second gear cycle described above
Go to 850 RPM idle, wait 5 seconds, end test cycle
Start next cycle
As part of the test protocol, each transmission ran for up to 1500 cycles.
Fluid samples were
taken every 30 cycles and evaluated. Speeds, loads, temperatures and pressures
were recorded
during every cycle
Figures 41-43 illustrate the results of these tests.
The 280PSTM operated normally through 1472 cycles, or three times the life of
the 1800 unit,
which failed catastrophically at 497 cycles. At the time of its failure, the
1800 unit had 75%
more iron in the fluid (indicative of gear wear) and 375% more lead in the
fluid (indicative of
bearing wear) than the 280PSTM.
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While the invention has been described with an emphasis on particular
embodiments thereof,
those skilled in the art may make various modifications to the described
embodiments of the
invention without departing from the scope of the invention. Although the
invention has been
described and disclosed in various terms and certain embodiments, the scope of
the invention is
not intended to be, nor should it be deemed to be, limited thereby and such
other modifications
or embodiments as may be suggested by the teachings herein are particularly
reserved, especially
as they fall within the breadth and scope of the claims here appended. Those
skilled in the art
will recognise that these and other variations are possible within the scope
of the invention as
defined in the following claims and their equivalents.
