U.S. patent application number 10/649515 was filed with the patent office on 2004-02-26 for dual strategy control for a toroidal drive type continuously variable transmission.
Invention is credited to Loeffler, John M., Lohr, Charles B., McIndoe, Gordon M..
Application Number | 20040038772 10/649515 |
Document ID | / |
Family ID | 22553183 |
Filed Date | 2004-02-26 |
United States Patent
Application |
20040038772 |
Kind Code |
A1 |
McIndoe, Gordon M. ; et
al. |
February 26, 2004 |
Dual strategy control for a toroidal drive type continuously
variable transmission
Abstract
An apparatus and method for operating a continuously variable
transmission (CVT), such as a toroidal drive type transmission, is
disclosed. The CVT is selectively operated in either a torque
control strategy and a ratio control strategy, depending upon the
operating conditions of the vehicle. Thus, the CVT is operated in
such a manner as to benefit from the advantageous aspects of both
the torque and ratio control strategies, while avoiding the
disadvantageous aspects of both strategies. The transition from the
torque control strategy to the ratio control strategy (and vice
versa) can be accomplished by simultaneously calculating the
control pressures that would result from operation in both the
torque and ratio control strategies, and further assigning a
weighted value to each of such calculated control pressures based
upon the current operating conditions. The summation of such
weighted values provides a composite control signal that
facilitates a smooth transition between the two control strategies.
The transition from the torque control strategy to the ratio
control strategy preferably occurs before a mode shift is effected.
Negative feedback is provided in response to ratio changes effected
by the control signals to increase stability and to compensate for
sensitivity differences at different ratio angles, loading, speeds,
and temperatures.
Inventors: |
McIndoe, Gordon M.;
(Perrysburg, OH) ; Lohr, Charles B.; (Maumee,
OH) ; Loeffler, John M.; (Whitehouse, OH) |
Correspondence
Address: |
MACMILLAN SOBANSKI & TODD, LLC
ONE MARITIME PLAZA FOURTH FLOOR
720 WATER STREET
TOLEDO
OH
43604-1619
US
|
Family ID: |
22553183 |
Appl. No.: |
10/649515 |
Filed: |
August 26, 2003 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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10649515 |
Aug 26, 2003 |
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09666745 |
Sep 20, 2000 |
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6663532 |
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60154876 |
Sep 20, 1999 |
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Current U.S.
Class: |
476/1 |
Current CPC
Class: |
F16H 15/38 20130101;
F16H 61/664 20130101; F16H 37/086 20130101; F16H 2037/0886
20130101; F16H 61/6649 20130101 |
Class at
Publication: |
476/1 |
International
Class: |
F16H 059/00 |
Claims
What is claimed is:
1. A continuously variable transmission comprising: an input shaft;
an output shaft; a continuously variable drive section connected
between said input shaft and said output shaft, said continuously
variable drive section including a roller that is mounted on a
trunnion for movement therewith, wherein movement of said roller
causes a change in a ratio provided by said continuously variable
drive section between said input shaft and said output shaft; a
control system for selectively operating said continuously variable
drive section in a torque control strategy under first
predetermined operating conditions and in a ratio control strategy
under second predetermined operating conditions.
2. The continuously variable transmission defined in claim 1
wherein said control system operates said continuously variable
drive section in the torque control strategy when the rotational
speed of said output shaft is less than a predetermined value, and
operates said continuously variable drive section in the ratio
control strategy when the rotational speed of said output shaft is
greater than a predetermined value.
3. The continuously variable transmission defined in claim 1
further including first and second gear assemblies that are
alternatively connected between said continuously variable drive
section and said output shaft, wherein portions of said first and
second gear assemblies rotate at a synchronous speed at a
predetermined mode point.
4. The continuously variable transmission defined in claim 3
wherein said control system operates said continuously variable
drive section in the torque control strategy when the rotational
speed of said output shaft is less than the mode point, and
operates said continuously variable drive section in the ratio
control strategy when the rotational speed of said output shaft is
greater than the mode point.
5. The continuously variable transmission defined in claim 3
wherein said control system operates said continuously variable
drive section in the torque control strategy when the rotational
speed of said output shaft is less than the mode point, and
operates said continuously variable drive section in the ratio
control strategy when the rotational speed of said output shaft
approaches the mode point.
6. The continuously variable transmission defined in claim 1
wherein said control system gradually transitions from operating
said continuously variable drive section in the torque control
strategy to operating said continuously variable drive section in
the ratio control strategy.
7. The continuously variable transmission defined in claim 6
wherein said control system assigns a first weighted factor to the
torque control strategy and a second weighted factor to the ratio
control strategy.
8. The continuously variable transmission defined in claim 1
wherein said control system varies the first weighted factor and
the second weighted factor to the ratio control strategy in
response to the predetermined operating conditions.
9. A continuously variable transmission comprising: an input shaft;
an output shaft; a continuously variable drive section connected
between said input shaft and said output shaft, said continuously
variable drive section including a roller that is mounted on a
trunnion for movement therewith, wherein movement of said roller
causes a change in a ratio provided by said continuously variable
drive section between said input shaft and said output shaft; a
control system that is responsive to an input signal for effecting
movement of said trunnion and said roller; and a feedback mechanism
that is responsive to movement of said trunnion and said roller for
causing said control system to alter the movement of said
trunnion.
10. The continuously variable transmission defined in claim 9
wherein said feedback mechanism is responsive to axial movement of
said trunnion and said roller for causing said control system to
alter the movement of said trunnion.
11. The continuously variable transmission defined in claim 9
wherein said wherein said feedback mechanism is responsive to
rotational movement of said trunnion and said roller for causing
said control system to alter the movement of said trunnion.
12. The continuously variable transmission defined in claim 9
wherein said wherein said feedback mechanism is responsive to axial
movement and rotational movement of said trunnion and said roller
for causing said control system to alter the movement of said
trunnion.
13. The continuously variable transmission defined in claim 9
wherein said control system includes a trunnion control valve that
selectively provides pressurized fluid to a trunnion cylinder
containing a control piston, said control piston being connected to
said trunnion for movement therewith, said feedback mechanism being
responsive to movement of said trunnion for varying the operation
of said trunnion control valve.
14. The continuously variable transmission defined in claim 14
wherein said feedback mechanism includes a cam that is connected to
said trunnion for movement therewith and a link that extends
between said cam and said trunnion control valve such that movement
of said cam with said trunnion causes movement of said link for
varying the operation of said trunnion control valve.
15. The continuously variable transmission defined in claim 14
wherein said control system includes a trunnion actuator that is
connected to said trunnion control valve by a link such that
movement of said link by said trunnion actuator controls the
operation of said trunnion control valve.
16. The continuously variable transmission defined in claim 15
wherein said feedback mechanism includes a cam that is connected to
said trunnion for movement therewith and said link further extends
between said cam and said trunnion control valve such that movement
of said cam with said trunnion causes movement of said link for
varying the operation of said trunnion control valve.
17. The continuously variable transmission defined in claim 14
wherein said cam includes a ramped surface that is engaged by said
link.
Description
CROSS REFERENCE TO RELATED APPLICATION
[0001] This application claims the benefit of U.S. Provisional
Application No. 60/154,876, filed Sep. 20, 1999, the disclosure of
which is incorporated herein by reference.
BACKGROUND OF THE INVENTION
[0002] This invention relates in general to toroidal drive type
continuously variable transmissions such as for use in vehicles. In
particular, this invention relates to an improved apparatus and
method for operating such a toroidal drive type continuously is
variable transmission.
[0003] In virtually all land vehicles in use today, a transmission
is provided in a drive train between a source of rotational power,
such as an internal combustion or diesel engine, and the driven
axle and wheels of the vehicle. One common type of transmission is
a discretely geared transmission, which includes a case containing
an input shaft, an output shaft, and a plurality of meshing gears.
Means are provided for connecting selected ones of the meshing
gears between the input shaft and the output shaft to provide a
desired gear ratio therebetween. The meshing gears contained within
the transmission case are of varying size so as to provide a
plurality of such discrete gear ratios between the input shaft and
the output shaft. By appropriately shifting among these various
discrete gear ratios, acceleration and deceleration of the vehicle
can be accomplished in a relatively smooth and efficient
manner.
[0004] Another type of such transmission is a continuously variable
transmission (CVT), wherein the ratio between the input shaft and
the output shaft is not provided in discrete gear increments, as
described above in connection with the discretely geared
transmission, but rather is adjustable in a continuous or
infinitely variable manner over a predetermined range. One known
structure for a continuously variable transmission includes a
forwardly positioned continuously variable drive section that is
connected through an intermediate co-axial drive section to a
rearwardly positioned output gear section. A representative
structure for such a CVT is disclosed in U.S. Pat. No. 5,607,372.
An electromechanical control system is often provided for
controlling the operation of the CVT in a desired manner.
[0005] Traditionally, such a control system has been programmed to
operate the CVT in either the torque control strategy or the ratio
control strategy. When the CVT is operated in the torque control
strategy, the control system correlates the throttle pedal position
of the engine (as set by the driver of the vehicle using the
accelerator pedal) with a desired amount of thrust for the vehicle.
When the CVT is operated in the ratio control strategy, the control
system controls the input speed of the CVT to a predetermined value
that represents some ideal operating parameters for the engine at
the current conditions. Although both of these control strategies
have been effective, it has been found that both of such control
strategies have disadvantages under certain operating conditions.
Thus, it would be desirable to provide an improved control strategy
for a CVT that achieves the advantages of both of such control
strategies, while avoiding the disadvantages associated
therewith.
SUMMARY OF THE INVENTION
[0006] This invention relates to an improved apparatus and method
for operating a toroidal drive type continuously variable
transmission (CVT). The CVT is selectively operated in either a
torque control strategy and a ratio control strategy, depending
upon the operating conditions of the vehicle. Thus, the CVT is
operated in such a manner as to benefit from the advantageous
aspects of both the torque and ratio control strategies, while
avoiding the disadvantageous aspects of both strategies. The
transition from the torque control strategy to the ratio control
strategy (and vice versa) can be accomplished by simultaneously
calculating the control valve signals that would result from
operation in both the torque and ratio control strategies, and
farther assigning a weighted value to each of such calculated
control valve signals based upon the current operating conditions.
The summation of such weighted values provides a composite control
signal that facilitates a smooth transition between the two control
strategies. The transition from the torque control strategy to the
ratio control strategy preferably occurs before a mode shift is
effected. Negative feedback is provided in response to ratio
changes effected by the control signals to increase stability and
to compensate for sensitivity differences at different ratio
angles, loading, speeds, and temperatures.
[0007] Various objects and advantages of this invention will become
apparent to those skilled in the art from the following detailed
description of the preferred embodiment, when read in light of the
accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
[0008] FIG. 1 is a schematic top plan view of a toroidal drive type
continuously variable transmission in accordance with this
invention.
[0009] FIG. 2 is a graph that illustrates the rotational speed of
the output shaft illustrated in FIG. 1 as a function of the angle
at which the traction rollers are oriented relative to the disks
when both the Mode One clutch and the Mode Two clutch are engaged,
assuming a constant rotational speed of the input shaft.
[0010] FIG. 3 is a block diagram of a control system for operating
the toroidal drive type transmission illustrated in FIG. 1.
[0011] FIG. 4 is a schematic perspective view of the mechanical
control and feedback mechanism of the control system illustrated in
FIG. 3.
[0012] FIG. 5 is a schematic perspective view of one of the disks,
one of the trunnions, and one of the traction rollers illustrated
in FIG. 1, wherein the trunnion and the traction roller are shown
in a centered position relative to the disk.
[0013] FIG. 6 is a schematic perspective view similar to FIG. 5,
wherein the trunnion and the traction roller are shown in an
axially displaced position relative to the disk.
[0014] FIG. 7 is a schematic perspective view similar to FIG. 6,
wherein the trunnion and the traction roller are shown in an
axially and rotational displaced position relative to the disk.
[0015] FIG. 8 is a graph similar to FIG. 2 that illustrates the
dual strategy operation of the transmission control unit
illustrated in FIG. 3 in the torque control strategy during certain
situations and in the ratio control strategy during other
situations.
[0016] FIG. 9 is a chart illustrating the gradual transition from
the torque control strategy to the ratio control strategy by
assigning a weighted value to each of the calculated valve signals
based upon the current operating conditions.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT
[0017] 1. CVT Structure
[0018] Referring now to the drawings, there is schematically
illustrated in FIG. 1a portion of a toroidal drive type
continuously variable transmission (CVT), indicated generally at
10, in accordance with this invention. Although this invention will
be described in the context of the illustrated structure for the
CVT 10, it will be appreciated that this invention may be used in
conjunction with other CVT structures. Therefore, the scope of this
invention is intended to be limited only by the claims appended
hereto. The general structure and operation of the illustrated CVT
10 are known in the art, and only those portions of the CVT 10 that
are necessary for a complete understanding of this invention are
illustrated. A complete discussion of the structure and operation
of the illustrated CVT 10 can be found in U.S. Pat. No. 5,607,372,
the disclosure of which is incorporated herein by reference.
[0019] The illustrated CVT 10 is an infinitely variable and
regenerative transmission that includes a forwardly positioned
continuously variable drive section, indicated generally at 20,
that is connected through an intermediate co-axial drive section,
indicated generally at 30, to a rearwardly positioned output gear
section, indicated generally at 40. For purposes of clarification,
the terms front or forward refer to the left side of the CVT 10
shown in FIG. 1, while the terms rear or rearward refer to the
right side thereof. All three of the sections 20, 30, and 40 are
enclosed within a housing 11 and are driven off of an input drive
shaft 12 that is rotatably driven by an engine (not shown). The
housing 11 is divided by forward and rearward internal walls 13 and
14, respectively, into three chambers, one for each of the sections
20, 30, and 40 of the exemplary CVT 10.
[0020] The illustrated continuously variable drive section 20 is a
dual cavity toroidal type, including first and second outboard
traction disks 21 and 22 and a single integral inboard disk 23, all
of which are disposed concentrically about the input shaft 18. A
forward toric cavity is defined between the front outboard disk 21
and the inboard disk 23, while a rearward toric cavity is defined
between the inboard disk 23 and the rear outboard disk 22. First
and second traction rollers 24 are disposed in each of the toric
cavities. The rollers 24 are preferably disposed transversely on
opposites sides of each toric cavity. Each pair of the traction
rollers 24 is engaged between the associated one of the outboard
disks 21 and 22 and the inboard disk 23. The rollers 24 are
supported on respective trunnions 25 in such a manner as to be
movable relative to the outboard disks 21 and 22 and the inboard
disk 23 to initiate a change in the transmission ratio. The
mechanism for effecting such relative movement of the trunnions 25
and, therefore, the rollers 24 will be described below.
[0021] The front outboard disk 21 can be splined directly onto the
input shaft 12 for rotation therewith. The rear outboard disk 22 is
connected for rotation with the input shaft 12 in the manner
described below. The rear end of the rear outboard disk 22 has a
hollow cylindrical collar portion 22a that extends through an axial
opening formed through the forward internal wall 13 into the middle
chamber of the housing 11 containing the co-axial drive section 30.
The cylindrical collar portion 22a and the rear outboard disk 22
are supported for rotation within this axial opening by an annular
bearing (not shown) provided therein. The inboard disk 23 is
supported on the input shaft 12 by bearings (not shown) so as to be
rotatable relative thereto. The rearward end of the inboard disk 23
is splined for rotation with a first torque tube 26, which is
disposed concentrically about the input shaft 12. The torque tube
26 extends concentrically through an opening formed though the rear
outboard disk 22 and further through the axial opening formed
through the front internal wall 13 into the middle chamber of the
housing 11 containing the co-axial drive section 30.
[0022] The illustrated continuously variable drive section 20 of
the CVT 10 is a dual cavity, half toroidal (or "off-center") type
continuously variable drive, wherein the included angle between the
traction contacts (i.e., where the traction rollers 24 contact the
disks 21, 22, and 23) is less than one hundred eighty degrees.
Notwithstanding this, it will be appreciated that this invention
may be practiced with other structures for the continuously
variable drive section 20. For example, the continuously variable
drive section 20 may be formed having only a single toroidal cavity
if desired. Alternatively, the continuously variable drive section
20 may be a full toroidal (or "on-center") type continuously
variable drive, wherein the included angle between the traction
contacts is approximately equal to one hundred eighty degrees.
[0023] The illustrated co-axial drive section 30 is a planetary
type drive, although such is not required. It will be understood
that the structure of the co-axial drive section 30 can be varied
in a number of ways, including being an epicyclic (as shown) or
other conventional planetary gear assembly or planetary traction
drive. A first sun gear 31 is formed integrally with the rearward
end of the torque tube 26. The co-axial drive section 30 also
includes a planetary carrier, indicated generally at 32, for
carrying a plurality of pairs of axially joined first and second
planet gears 33 and 34 around the input shaft 12. Preferably, each
pair of compound planet gears 33 and 34 is of one-piece
construction. The illustrated epicyclic planetary co-axial drive
section 30 does not include a ring gear, as is typically present in
conventional planetary gear assemblies.
[0024] The illustrated planetary carrier 32 is formed from three
components, namely, a front support member 32a, a rear support
member 32b, and a central member 32c extending therebetween. The
front support member 32a is disposed concentrically about and is
splined for rotation with the rearwardly extending collar portion
22a of the second outboard disk 22. The rear support member 32b has
a rearwardly extending, hollow cylindrical support portion 32d
formed thereon that extends through an axial opening formed through
the rearward internal wall 14 into the rear chamber of the housing
11 containing the output gear section 40. A plurality of bearing
pins 35 (only one is illustrated) extends between and is secured to
the front support member 32a and the rear support member 32b.
Preferably, the bearing pins 35 are equally spaced about the
planetary carrier 32 and the input shaft 12. The central member 32c
of the planetary carrier 32 includes a hub portion 32e that is
splined onto the input shaft 18 for rotation therewith.
[0025] Each pair of planet gears 33 and 34 is mounted on one of the
bearing pins 35 for rotation relative thereto. The forward planet
gears 33 mesh with the input sun gear 31 provided on the rear end
of the torque tube 26, while the rearward planet gears 34 mesh with
a second sun gear 37 provided in the co-axial drive section 30. The
second sun gear 37 is smaller than the first sun gear 31 and is
splined for rotation with the forward end of a second torque tube
38. The rearward end of second tube 38 extends co-axially through
the rearwardly extending, hollow cylindrical support portion 32d of
the planetary carrier 32 and through the axial opening formed
through the rearward internal wall 14 into the rear chamber of the
housing 11 containing the output gear section 40.
[0026] Within the rear chamber of the housing 11, a Mode One
planetary gear assembly, indicated generally at 41, is provided.
The Mode One planetary gear assembly 41 includes a sun gear 42, a
planetary carrier 43 carrying relatively rotatable planet gears 44,
and a ring gear 45. The Mode One sun gear 42 is provided on the
rearward end of the torque tube 38 for rotation therewith. The Mode
One planetary carrier 43 extends rearwardly into cooperation with a
Mode One clutch 46, the purpose for which will be described below.
The Mode One planet gears 44 mesh with the Mode One sun gear 42 and
the Mode One ring gear 45. The Mode One ring gear 45 is splined to
the input shaft 12 for rotation therewith.
[0027] Also within the rear chamber of the housing 11, a Mode Two
planetary gear assembly, indicated generally at 51, is provided.
The Mode Two planetary gear assembly 51 includes a sun gear 52, a
planetary carrier 53 carrying relatively rotatable planet gears 54,
and a ring gear 55. The Mode Two sun gear 52 is also provided on
the rearward end of the torque tube 38 for rotation therewith. The
Mode Two planetary carrier 53 is formed integrally with or
connected to the rearward internal wall 14 of the housing 11. The
Mode Two planet gears 54 mesh with the Mode Two sun gear 52 and the
Mode Two ring gear 55. The Mode Two ring gear 55 extends rearwardly
into cooperation with a Mode Two clutch 56, the purpose for which
will also be described below.
[0028] The Mode One clutch 46 is provided to selectively connect
the Mode One planetary carrier 43 to an output shaft 57 for
rotation therewith. Similarly, the Mode Two clutch 55 is provided
to selectively connect the Mode Two ring gear 55 to the output
shaft 57 for rotation therewith. The output shaft 57 extends
outwardly from the housing 11 of the CVT 10 and can be connected in
a conventional manner to an axle assembly (not shown) or other
mechanism to rotatably drive the wheels of the vehicle. Therefore,
depending upon the operation of the clutches 46 and 56, either the
Mode One planetary gear assembly 41 or the Mode Two planetary gear
assembly 51 will be connected to transmit power to the output shaft
57.
[0029] In operation, the input shaft 12 of the CVT 10 is rotated
continuously by the engine of the vehicle. The rotation of input
shaft 12, in turn, directly rotates the first outboard traction
disk 21, the planetary carrier 32 of the co-axial drive section 30,
and the ring gear 45 of the Mode One planetary gear assembly 41,
all in the same rotational direction. For the purpose of this
description, the rotation of the input shaft 12 will be referred to
as being in a positive direction, and any oppositely rotating
element will be referred to as being in a negative direction.
Because of the splined connection between the forwardly extending,
hollow cylindrical support 32a and the rearwardly extending collar
portion 22a of the second outboard disk 22 the rotation of the
planetary carrier 32 of the co-axial drive section 30 causes the
second outboard traction disk 22 to also rotate in the same
positive direction as the input shaft 12.
[0030] The rotating outboard disks 21 and 22 impinge on and rotate
the traction rollers 24 in a manner that is well known and standard
for toroidal type drives. The traction rollers 24 then impinge on
and cause the inboard traction disk element 23 to rotate it in a
negative direction. The inboard traction disk element 23 thus
rotates the first sun gear 31 of the co-axial drive section 30 in a
negative direction through the torque tube 26. As a result, the
planetary gears 33 and 34 are rotated in a positive direction by
the first sun gear 31, and the second sun gear 37 is rotated in a
negative direction by rearward planetary gears 34. As a result, the
Mode One sun gear 43 is also rotated in a negative direction.
[0031] As discussed above, the Mode One sun gear 42 is rotated in a
negative direction, while the Mode One ring gear 45 is rotated in a
positive direction. Depending upon the ratio of the continuously
variable drive section 20, the Mode One planetary carrier 43 will
either (1) rotate in a positive direction when the influence of the
Mode One ring gear 45 is greater than the influence of the Mode One
sun gear 42, (2) rotate in a negative direction when the influence
of the Mode One sun gear 42 is greater than the influence of the
Mode One ring gear 45, or (2) remain stationary when their
influences balance each other. The sum of these two influences,
whatever it is, is generated through the Mode One carrier 43 to the
Mode One clutch 46. If the Mode One clutch 46 is engaged, then this
rotational output is then transmitted to the output shaft 57,
causing the output shaft 57 to rotate in the same direction of
rotation as the Mode One carrier 43. In this way, the Mode One
planetary gear assembly 41 functions as a summing or mixing
planetary gear assembly, enabling the output shaft 57 to be rotated
in either a positive direction, a negative direction, or kept
rotationally stationary.
[0032] When the Mode One planetary gear assembly 41 is engaged
through the Mode One clutch 46, the CVT 10 operates in a
regenerative mode, wherein some of the power/speed that is
transmitted from the engine through the input shaft 12 and into the
Mode One ring gear 45 is siphoned off through the Mode One planet
gears 44 and the Mode One sun gear 42 and routed back to the second
sun gear 37 of the co-axial drive section 30. From the second sun
gear 37 of the co-axial drive section 30, this siphoned off
power/speed is transmitted back through the co-axial drive section
30 and the continuously variable drive section 20 to the input
shaft 12 of the engine. How much power is siphoned off in this
manner is determined by the ratio or angle of the rollers 24 in the
toric cavities. The use of this regenerative design is generally
considered to be necessary when no external coupling device (such
as a mechanical friction clutch or a fluid torque converter) is
provided between the engine and the CVT 10.
[0033] As the rotation of the second sun gear 37 of the co-axial
drive section 30 is affected, so too is the rotation of the first
sun gear 31 and the planetary carrier 32 through the pairs of
planet gears 33 and 34. Changes in the rotation of carrier 32
affect the rotation of the input shaft 12 directly through the hub
portion 32e and indirectly through the toroidal drive 12 by
directly affecting the rotation of the second outboard traction
disk 22 and, in turn, the first outboard traction disk 21. In this
way, the CVT 10 exhibits a recirculatory power loop between the
output gear section 40 and the continuously variable drive section
20 through the co-axial drive section 30.
[0034] If enough power/speed is diverted away via the Mode One sun
gear 42, the Mode One carrier 43 will not rotate at all. This
condition is often referred to as a geared neutral condition. If
the rotational speed of the Mode One sun gear 42 further increases
beyond the geared neutral condition, the rotation of Mode One
carrier 43 will reverse in direction and go opposite to the
direction of rotation of the Mode One ring gear 45. In this way,
the CVT 10 is operated in a reverse ratio. The Mode One planetary
gear assembly 41 can, therefore, operate the CVT 10 from a small
reverse regime through a geared neutral or zero output speed and
then into a forward speed.
[0035] However, because of the regenerative operation described
above, the Mode One carrier 43 is somewhat limited in its ability
to provide higher forward speeds. Thus, the Mode Two planetary gear
assembly 51 is provided to supplement the forward ratio of the Mode
One planetary gear assembly 41 to enable greater forward speeds to
be attained. When a predetermined rotational speed of the Mode One
planetary gear assembly 41 is neared or reached, the CVT 10 is
shifted from operation through the Mode One planetary gear assembly
41 to operation through the Mode Two planetary gear assembly 51.
This can be accomplished by disengaging the Mode One clutch 46 and
engaging the Mode Two clutch 56. This operation is referred to as a
mode shift. Preferably, the mode shift occurs at or near the point
at which the rotational speed of the Mode One carrier 43 is
approximately equal to the rotational speed of the Mode Two ring
gear 45. This point is referred to as a mode point. The operation
of the CVT 10 to perform the mode shift at the mode point is
described in further detail below.
[0036] When the Mode Two planetary gear assembly 51 is engaged
through the Mode Two clutch 56, power from the input shaft 12 is
transmitted through the continuously variable drive section 20 and
the co-axial drive section 30 to rotate the Mode Two sun gear 52 in
a negative direction. This causes the Mode Two planet gears 54 to
rotate in a positive direction, which results in the Mode Two ring
gear 55 rotating in a positive direction. The Mode Two ring gear 55
directly rotates the output shaft 57 through the Mode Two clutch
56. Because the Mode Two carrier 53 is fixed to the rearward wall
14 of the housing 11, there is only one input and one output for
the Mode Two planetary gear assembly 51. Thus, the Mode Two
planetary gear assembly 51 also allows the CVT 10, while in its
forward regime, to reach its upper operating speeds by transmitting
power directly from the input shaft 12 to the output shaft 57. When
the Mode Two planetary gear assembly 51 is engaged, the CVT 10 is
operated in a split torque mode. Most of the torque is passed
through the continuously variable drive section 20, the co-axial
drive section 30, and the Mode Two planetary gear assembly 51 to
the output shaft 57. A small amount of torque, however, is passed
from the input shaft 12 directly to the planetary carrier 12 of the
co-axial drive section 20, where it is summed with the other
torque.
[0037] Although the output gear section 40 has been described and
illustrated as including two planetary gear assemblies 41 and 51,
it will be appreciated that other types of gear assemblies can be
used. Also, the output gear assembly 40 may further include one or
more additional planetary gear assemblies (not shown) if desired
for further extending the available overall ratios of the CVT 10.
Lastly, the operation of the two illustrated planetary gear
assemblies 41 and 51 is controlled by the clutches 46 and 56.
However, other control devices, such as brakes and the like, may be
used to control the operation of the two planetary gear assemblies
41 and 51.
[0038] FIG. 2 is a graph that illustrates the rotational speed of
the output shaft 57 as a function of the angle at which the
traction rollers 24 are oriented relative to the disks 21, 22, and
23 when both the Mode One clutch 46 and the Mode Two clutch 56 are
engaged, assuming a constant rotational speed of the input shaft
12. As shown by the dotted line 60, when the Mode One planetary
gear assembly 41 is connected to the output shaft 57 by the Mode
One clutch 46, the rotational speed of the output shaft 57
decreases as the angle of the traction rollers 24 increases.
However, as shown by the solid line 61, when the Mode Two planetary
gear assembly 51 is connected to the output shaft 57 by the Mode
Two clutch 56, the rotational speed of the output shaft 57
increases as the angle of the traction rollers 24 increases. The
two lines 60 and 61 intersect at a point 62, which is referred to
as the mode point. The mode point 62 represents the angle of the
traction rollers 24 where the rotational speed of the output shaft
57 is the same regardless of whether the Mode One clutch 46 or the
Mode Two clutch 56 is engaged. Ideally, to obtain a smooth
transition, the above-described mode shift occurs only when the
traction rollers 24 are positioned at (or at least near) the mode
point 62.
[0039] 2. Control System and Feedback Mechanism
[0040] Referring now to FIG. 3, there is illustrated a block
diagram of a control system, indicated generally at 100, for
operating the CVT 10 described above. The control system 100
includes a transmission control unit 101 that may, for example, be
embodied as any conventional microprocessor, programmable
controller, or similar electronic computing device. Typically, the
transmission control unit 101 communicates with a similar
electronic engine control unit 102 that is provided on the vehicle
for controlling the operation of the engine of the vehicle. The
interactions between the transmission control unit 101 and the
engine control unit 102 are generally conventional in the art.
[0041] The transmission control unit 101 receives a number of input
signals that are representative of various operating conditions of
the CVT 10 and other portions of the vehicle. As shown in FIG. 3,
such input signals can include a variety of signals that are
representative of the operating conditions of the vehicle engine,
such as throttle pedal position, engine power and/or governed
speed, engine temperature, and the rotational speed of the input
shaft 12 to the CVT 10 (which is rotatably driven directly by the
engine). Such signals may, if desired, be provided to the
transmission control unit 101 from the engine control unit 102.
Alternatively, such signals may be provided to the transmission
control unit 101 from discrete sensors. The input signals to the
transmission control unit 101 can also include a variety of signals
that are representative of the operating conditions of the CVT 10,
such as transmission temperature, the rotational speed of the
inboard disk 23, the rotational speed of the output shaft 57
(which, depending upon the direction of rotation may be positive or
negative), the axial displacement of the trunnion 25, the angular
displacement of the trunnion 25, and the magnitude of the control
pressure supplied to actuate the trunnion 25. Such signals are
preferably provided to the transmission control unit 101 from
discrete sensors. Lastly, the input signals to the transmission
control unit 101 can further include a variety of signals that are
representative of the operating conditions of other portions of the
vehicle, such as brake actuation and gear selector position. The
above list of input signals to the transmission control unit 101 is
not intended to be exhaustive, but rather merely exemplary of some
of the input signals that can be used by the transmission control
unit 101 to control the operation of the CVT 10, such as throttle
valve position and fuel governor position, for example.
[0042] In response to these input signals, the transmission control
unit 101 generates output signals to control the gear train
selection in the output gear assembly 40. As discussed above, the
illustrated CVT 10 includes the pair of planetary gear assemblies
41 and 51 that are selectively engaged and disengaged to rotatably
drive the output shaft 57 as desired. Thus, in the illustrated
embodiment, the output signals from the transmission control unit
101 control the operation of the Mode One clutch 46 and the Mode
Two clutch 56 in the manner described above. The specific
mechanisms for operating such clutches 46 and 56 are well known in
the art and need not be explained here for a complete understanding
of this invention. The specific strategy for shifting between the
Mode One clutch 46 and the Mode Two clutch 56 will be described in
further detail below.
[0043] The transmission control unit 101 also generates output
signals to control the movements of the trunnions 25 so as to
effect changes in the ratio of the continuously variable drive
section 20. As mentioned above, the rollers 24 are supported on
respective trunnions 25 in such a manner as to be movable relative
to the outboard disks 21 and 22 and the inboard disk 23 to initiate
a change in such ratio. Accordingly, by controlling the movements
of the trunnions 25, the angles of the rollers 24 relative to the
disks 21, 22, and 23 (and, thus, the ratio of the continuously
variable drive section 20) can be controlled in a desired manner.
The specific structure for accomplishing this is illustrated
schematically in FIG. 4. As shown therein, a trunnion actuator 103
receives output signals from the transmission control unit 101 to
control the operation thereof. The trunnion actuator 103 is
generally conventional in the art and is adapted to generate
mechanical movements in response to the output signals from the
transmission control unit 101. The trunnion actuator 103 may, for
example, be embodied as any known device, electromechanical or
otherwise, that is responsive to the output signals from the
transmission control unit 101 for causing corresponding linear
reciprocating movement of an output member 103a.
[0044] The output member 103a of the trunnion actuator 103 is
connected through a mechanical link 104 to an input member 105a of
a trunnion control valve 105. In the illustrated embodiment, the
output member 103a of the trunnion actuator 103 is connected to a
first end of the mechanical link 104, while the input member 105a
of the trunnion control valve 105 is connected to a central portion
of the mechanical link 104. However, the trunnion actuator 103 and
the trunnion control valve 105 may be connected to the mechanical
link 104 in any desired manner. The trunnion control valve 105 is
also conventional in the art and may, for example, be embodied as
any known hydraulic valve that is capable of controlling the flow
of fluid therethrough in response to movement of the mechanical
link 104. Typically, the trunnion control valve 105 is connected
between a source of pressurized hydraulic fluid 106 and each of the
trunnion cylinders 107 that are mechanically connected to the
trunnions 25 shown in FIG. 1. By supplying pressurized fluid to
these trunnion cylinders 107, mechanical movement of the trunnions
25 can be effected.
[0045] The manner in which the trunnion 25 can be moved by
operation of the trunnion control valve 105 is shown in FIGS. 5, 6,
and 7. For the purpose of illustration, let it be assumed that the
trunnion 25 is initially in the position illustrated in FIG. 5,
wherein the roller 24 is essentially centered relative to axis of
rotation of the illustrated disk 21 and engages the disk 21 at a
predetermined radius from the center thereof. To accomplish this, a
certain amount of pressurized fluid is supplied from the source of
pressurized fluid 106 through the trunnion control valve 105 to the
associated trunnion cylinder 107. The trunnion cylinder 107
includes a control piston 107a that is connected to the trunnion 25
for axial movement therewith. The control piston 107a is disposed
within the trunnion cylinder 107 so as to divide the interior
thereof into two chambers. When the magnitude of pressurized fluid
supplied to the two chambers is different, a differential force is
exerted-against the control piston 107a, urging to move axially
relative to the trunnion cylinder 107.
[0046] As a practical matter, in order to maintain the control
piston 107a in a static position within the trunnion cylinder 107,
such as in the centered position illustrated in FIG. 5, fluid
pressure of a predetermined magnitude is supplied to the first
chamber, while the second chamber is essentially vented. The
resultant pressure differential created across the control piston
107a causes a first axial force to be exerted against the control
piston 107a, urging it to move axially in a first direction
relative to the trunnion cylinder 107. This first axially directed
force is counteracted by a second similar, but oppositely directed
reaction force that is exerted against the trunnion 25 as a result
of the engagement of the roller 24 between one of the outboard
disks 21 and 22 and the inboard disk 23. Because these first and
second axially directed forces are essentially identical in
magnitude, but opposite in direction, the trunnion 25 is maintained
in a static position, such as the centered position illustrated in
FIG. 5. In this position, the roller 24 effects a predetermined
ratio between the outboard disks 21 and 22 and the inboard disk 23
of the continuously variable drive section 20.
[0047] To vary the ratio of the continuously variable drive section
20, the magnitudes of fluid pressure in either or both of the first
and second chambers of the trunnion cylinder 107 are increased or
decreased (depending upon the desired direction of movement) to
vary the differential pressure across the control piston 107a. To
accomplish this, the trunnion actuator 103 is operated by the
transmission control unit 101 to move the output member 103a. This
causes the mechanical link 104 to be moved, resulting in
corresponding movement of the input member 105a of the trunnion
control valve 105. As a result, the amount of pressurized fluid
supplied from the source of pressurized fluid 106 through the
trunnion control valve 105 to the first and second chambers of the
trunnion cylinder 107 is varied. In response to such fluid pressure
changes, the magnitude of the differential pressure across the
control piston 107a (and, therefore, the magnitude of the first
axial force exerted against such control piston 107a) is changed to
be different from the second reaction force described above. In the
illustrated embodiment, the operation of the source of pressurized
fluid 106 is directly controlled by the transmission control unit
101, although such is not necessary. For the sake of economy, a
feedback loop 106a may be provided for further controlling the
operation of the source of pressurized fluid 106 in response to the
magnitude of the pressurized fluid in the hydraulic line extending
from the trunnion control valve 105 to the trunnion cylinders
107.
[0048] Consequently, the control piston 107a, the trunnion 25, and
the roller 24 are axially displaced relative to the disk 21, such
as shown by the arrow in FIG. 6. The roller 24 now engages the disk
21 at a different radius from the center thereof than as shown in
FIG. 5. In compliance with a steering vector force that is now
generated by the engagement of the axially displaced roller 24 with
the disk 21, the roller 24 and the trunnion 25 then rotate relative
to the axis of movement, as shown by the arrow in FIG. 7. In this
manner, axial movement of the trunnion 25 causes the roller 24 to
effect a different ratio between the outboard disks 21 and 22 and
the inboard disk 23 than as shown in FIG. 5.
[0049] Referring back to FIG. 4, a sensor 108 is connected to the
trunnion 25 for axial and rotational movement therewith. The sensor
108 is responsive to these axial and rotational movements for
generating electrical signals that are representative thereof back
to the transmission control unit 101. The sensor 108 can be
embodied as any conventional sensor or plurality of sensors. The
purpose for the sensor 108 will be explained below.
[0050] As shown in FIG. 3, a feedback mechanism 109 is provided for
enhancing the stability of the movement of the trunnion 25
described above. One exemplary structure of the feedback mechanism
109 is shown in FIG. 4 as an annular cam that is secured to the
trunnion 25 for axial and rotational movement therewith. The
illustrated cam 109 has a curved ramp surface 109a formed on one
side thereof. As mentioned above, the output member 103a of the
trunnion actuator 103 is connected to a first end of the mechanical
link 104, while the input member 105a of the trunnion control valve
105 is connected to a central portion of the mechanical link 104. A
second end of the mechanical link 104 bears upon the curved ramp
surface 109a formed on the cam 109. As a result, when the trunnion
25 is displaced axially and rotationally as described above, the
second end of the mechanical link 104 is moved therewith. Such
movement of the second end of the mechanical link 104 is designed
to limit the ability of the trunnion actuator 103 to effect
movement of the input member 105a of the trunnion control valve
105.
[0051] For example, assume that the transmission control unit 101
generates an output signal to the trunnion actuator 103 to effect
movement of the trunnion 25. In response to that output signal, the
trunnion actuator 103 extends the output member 103a, causing
movement of the mechanical link 104. Such movement causes the input
member 105a of the trunnion control valve 105 to be moved in such a
manner as to change the amount of pressurized fluid supplied to the
trunnion cylinder 107. This change in the amount of pressurized
fluid causes the trunnion 25 to be initially axially displaced in
the manner described above. However, such axial displacement of the
trunnion 25 causes the cam 109 to move axially as well. Because the
second end of the mechanical link 104 bears upon the cam 109, such
axial displacement of the trunnion 25 causes movement of the input
member 105a of the trunnion control valve 105. Thus, the axial
displacement of the trunnion 25 functions to provide negative
feedback to the input member 105a of the trunnion control valve 105
which is sufficient to limit the magnitude of the axial
displacement of the trunnion 25. As a result, the response rate of
the trunnion control valve 105 is reduced in accordance with the
axial displacement of the trunnion 25.
[0052] Additionally, as described above, axial displacement of the
trunnion 25 causes rotational displacement as well. As mentioned
above, the cam 109 is connected to the trunnion 25 for rotational
movement therewith. Because the second end of the mechanical link
104 bears upon the curved ram surface 109a of the cam 109, such
rotational displacement of the trunnion 25 causes further movement
of the second end of the mechanical link 104 and, therefore, the
input member 105a of the trunnion control valve 105. Thus, the
rotational displacement of the trunnion 25 functions to provide
further negative feedback to the input member 105a of the trunnion
control valve 105. The negative feedback from the rotational
displacement of the trunnion 25 is sufficient to cause the trunnion
control valve 105 to restore the trunnion 25 to the centered or
equilibrium position illustrated in FIG. 5. Thus, it can be seen
that the negative feedback generated by the axial displacement of
the trunnion 25 is effective to slow or otherwise limit the rate of
change of such axial movement, while the negative feedback
generated by the rotation displacement of the trunnion 25 is
effective to limit the magnitude of such change.
[0053] It will be appreciated that the shape of the curved ram
surface 109a of the cam 109 can be varied as desired to provide a
desired amount of negative feedback. Thus, the ratio of rise per
unit rotation of the curved ram surface 109a of the cam 109 can be
linear or non-linear as desired. Such a non-linear rate of change
can be used to customize the magnitude of the negative feedback for
the particular rotational displacement of the trunnion 25 and the
roller 24 to account for varying degrees of sensitivity of the
structure at varying rotational positions. Similarly, if desired,
the negative feedback resulting from the axial displacement of the
trunnion 25 can be varied in a non-linear manner. This can be
accomplished by feeding the axial displacement from the sensor 108
back to the transmission control unit 101, where it can be used to
modify the initial control signal generated to the trunnion
actuator 103. Alternatively, a mechanical system (not shown) can be
employed to vary the negative feedback resulting from the axial
displacement of the trunnion 25 in a non-linear manner.
[0054] The above-described feedback mechanism 109 is a purely
mechanical system. However, it will be appreciated that the same
result can be achieved using an electronic feedback mechanism. Such
a mechanism can be provided by employing an electrically actuated
trunnion control valve, instead of the illustrated mechanically
actuated trunnion control valve 105. The input signal from the
transmission control unit 101 to such an electrically actuated
trunnion control valve would be summed with electrical feedback
position signals that are representative of the axial and
rotational displacement of the trunnion 25, such as from the sensor
108. In a manner similar to that described above, the electrical
negative feedback functions to limit the rate of change and
magnitude of the displacement of the trunnion 25. An electronic
feedback mechanism could additionally be responsive to a variety of
other operating conditions, including speed, ratio, temperature,
mode of operation (i.e., whether the Mode One planetary gear
assembly 41 or the Mode Two planetary gear assembly 51 is clutched
to the output shaft 57), and whether the CVT 10 is being operated
in torque or ratio control, as described below.
[0055] Lastly, the transmission control unit 101 further generates
output signals to control the magnitude of the clamping force
exerted by traction rollers 24 against the disks 21, 22, and 23. To
accomplish this, a clamping force control valve 110 receives output
signals from the transmission control unit 101 to control the
operation thereof. The clamping force control valve 110 is
generally conventional in the art and may, for example, be embodied
as any known hydraulic valve that is capable of controlling the
flow of fluid therethrough in response to the output signals from
the transmission control unit 101. Typically, the clamping force
control valve 110 is connected between the trunnion control valve
105 and a plurality of clamping force cylinder assemblies 110 that
are typically located in the trunnions 25 or behind the rollers 24
shown in FIG. 1. The operation of the illustrated clamping force
control valve 110 is controlled directly by the transmission
control unit 101, although such is not necessary. By supplying
pressurized fluid to these clamping force cylinder assemblies 111,
the magnitude of the normal force exerted by traction rollers 24
against the disks 21, 22, and 23 can be varied as desired. The
operation of the clamping force cylinder assemblies 111 in this
manner is generally conventional in the art.
[0056] An additional aspect of this invention relates to the manner
in which pressurized fluid is supplied to the chambers of the
trunnion cylinder 107 during a mode shift (i.e., when the Mode One
clutch 46 and the Mode Two clutch 56 are actuated to reverse the
connection of the two planetary gear assemblies 41 and 51 to the
output shaft 57). As mentioned above, pressurized fluid is supplied
to the two chambers of the trunnion cylinder 107 so as to position
the control piston 107a as desired therein. Also, in the manner
described above, the negative feedback generated by the interaction
of the mechanical link 104 with the cam 109 limits the rate of
change and magnitude of such movement by re-establishing the
equilibrium across the control piston 107a after a certain amount
of axial and rotational displacement.
[0057] It is known that when a mode shift is effected in the output
gear section 40, a reversal of torque occurs within the
continuously variable drive section 20 of the CVT 10. In order to
maintain the control piston 107a in the same position within the
trunnion cylinder 107 during this reversal of torque, the
magnitudes of fluid pressure in the first and second chambers of
the trunnion cylinder 107 must be reversed and altered in
accordance with the new reaction force, which may be different from
the original reaction force. In other words, the magnitude of fluid
pressure within the first chamber (which was relatively high before
the mode shift) must be decreased, and the magnitude of fluid
pressure within the second chamber (which was relatively low before
the mode shift) must be increased. When this is done, the control
piston 107a is maintained in the same position within the trunnion
cylinder 107 after the reversal of torque occurs during the mode
shift.
[0058] In the past, these reversing adjustments of the magnitudes
of fluid pressure within the first and second chambers of the
trunnion cylinder 107 occurred simultaneously when the mode shift
was initiated. In other words, the relatively high fluid pressure
in the first chamber was decreased simultaneously as the relatively
low fluid pressure in the second chamber was increased. Although
effective, it has been found that such simultaneous adjustments of
the magnitudes of fluid pressures within the first and second
chambers of the trunnion cylinder 107 can result in undesirable
transient movements of the control piston 107a and, therefore, the
trunnions 25 during the mode shift.
[0059] To prevent this from occurring, this invention contemplates
that prior to the commencement of the mode shift, the magnitude of
fluid pressure within the first chamber (which was initially
relatively high) is further increased above the originally relative
high level, and the magnitude of fluid pressure within the second
chamber (which was initially relatively low) is increased by a
similar magnitude. This simultaneous increase in fluid pressures in
both of the first and second chambers of the trunnion cylinder 107
maintains the same pressure differential across the control piston
107a. Thus, no axial movement of the control piston 107a occurs
during this initial pressurization phase. Accordingly, when the
mode shift occurs, all that is required is to vent the first
chamber of the trunnion cylinder 107 to effect the reversing
adjustments of the magnitudes of fluid pressure within the first
and second chambers of the trunnion cylinder 107. It has been found
that the venting of the first chamber can occur more rapidly than
the pressurization of the second chamber, as previously performed.
Consequently, because all that is required is to vent the first
chamber, undesirable transient movements of the control piston 107a
during the mode shift are effectively avoided.
[0060] 3. Dual Strategy Control
[0061] As discussed above, the transmission control unit 101 of the
control system 100 is responsive to the various input signals
supplied thereto for generating output signals to actuate the
trunnion actuator 103. In response thereto, the trunnion actuator
103 is effective to operate the trinnion control valve 105, which
regulates the magnitude of pressurized fluid that is provided to
the first and second chambers of the trunnion cylinder 107. For
toroidal drive type continuously variable transmissions that
utilize pressurized fluid to effect movement of the trunnions 25,
such as the illustrated CVT 10 discussed above, the magnitude of
this pressurized fluid (which is referred to as the control
pressure) is the root action of any control strategy used to
control the operation of the CVT 10. Thus, if the control pressure
in the trunnion cylinder 107 is increased above an equilibrium
point, then the ratio of the continuously variable drive section 20
will increase. Similarly, if the control pressure in the trunnion
cylinder 107 is decreased below the equilibrium point, then the
ratio of the continuously variable drive section 20 will decrease.
Finally, if the control pressure in the trunnion cylinder 107 is
maintained at the equilibrium point, then the ratio of the
continuously variable drive section 20 will not change.
[0062] There are two commonly known control strategies of operation
for operating the CVT 10. The first control strategy is referred to
as the torque control strategy. When the CVT 10 is operated in the
torque control strategy, the goal of the control system 100 is to
correlate the throttle pedal position of the engine (as set by the
driver of the vehicle using the accelerator pedal) with a desired
amount of thrust for the vehicle. Thus, when the throttle pedal
position is relatively low, it can be inferred that the driver of
the vehicle desires a relatively small amount of such thrust as a
response. Conversely, when the throttle pedal position is
relatively high, it can be inferred that the driver of the vehicle
desires a relatively large amount of such thrust as a response.
[0063] Torque control operation of the CVT 10 is achieved by
correlating the throttle pedal position to a predetermined
magnitude of pressure at the trunnion cylinders 107. As discussed
above, the magnitude of this control pressure is directly related
to the axial displacement of the trunnions 25 and, therefore, the
angle of the traction rollers 24. Accordingly, as also discussed
above, the magnitude of this control pressure is also directly
related to the ratio provided by the continuously variable drive
section 20 of the CVT 10. For example, let it be assumed that the
throttle is moved to and maintained at a single predetermined
position. The transmission control unit 101 is responsive to that
throttle pedal position for determining a single predetermined
magnitude of the control pressure that is to be supplied to the
trunnion cylinders 107. Accordingly, the transmission control unit
101 will actuate the trunnion actuator 103 (and, thus, the trunnion
control valve 105) to achieve the desired control pressure in the
trunnion cylinder 107. Consequently, the trunnions 25 are axially
displaced in the manner described above, resulting in the
continuously variable drive section 20 being operated at a ratio
angle of the trunnions 25 that allows a balance between the torque
from the engine and the desired control pressure. The vehicle
responds, at least under normal circumstances, by gradually
accelerating at a rate that is related to the single predetermined
ratio as determined by the throttle pedal position.
[0064] As a practical matter, however, the throttle pedal position
often varies considerably when the vehicle is operated at a
relatively low speed, wherein acceleration and deceleration of the
vehicle is relatively frequent. For each throttle pedal position,
the control system 100 determines a corresponding control pressure.
Accordingly, during such relatively low speed operation, the
transmission control unit 101 will continually adjust the control
pressure in the trunnion cylinders 107 in accordance with the
throttle pedal position. As a result, the magnitude of the thrust
supplied by the CVT 10 to the wheels of the vehicle will vary as
well. The torque control strategy is well suited for operating the
CVT 10 at relatively low vehicle speeds because it can provide for
very smooth acceleration of the vehicle, particularly when the CVT
10 is operated in the lower ratios, similar to a torque converter
or similar fluid coupling structure. When the vehicle is operated
at relatively high speeds, the throttle pedal position usually
varies by a lesser amount. When the throttle pedal position is
relatively constant (at any vehicle speed), the transmission
control unit 101 will cause a relatively constant amount of thrust
to be generated. As a result, the vehicle will be operated, at
least under normal circumstances, at a relatively constant
speed.
[0065] Unfortunately, the torque control strategy is not well
suited for facilitating the performance of the mode shift described
above, such as when the CVT 10 is shifted from the Mode One
planetary gear assembly 41 to the Mode Two planetary gear assembly
51 by disengaging the Mode One clutch 46 and engaging the Mode Two
clutch 56. As mentioned above, the mode shift preferably occurs at
or near the mode point, wherein the Mode One carrier 43 rotates at
the same speed as the Mode Two ring gear 45. When operated in the
torque control strategy, the CVT 10 is unable to precisely control
when the mode point occurs because it is focused on obtaining and
maintaining the desired control pressure at the trunnion cylinder
107, regardless of the rotational speeds of the components therein.
In the torque control strategy, the mode shift occurs when the mode
point is sensed by the transmission control unit 101. However, the
transmission control unit 101 is unable to direct the components of
the CVT 10 to the mode point or hold them at the mode point until
the mode shift is completed. Consequently, it has been found that
operation of the CVT 10 in the torque control strategy can, at
least in some instances, result in the mode shift occurring away
from the mode point. This can result in the generation of an
undesirable transient torque in the CVT 10 as a result of the
unsynchronized mode shift.
[0066] Ratio control operation of the CVT 10, on the other hand, is
achieved by correlating the throttle pedal position to a
predetermined desired rotational speed for the input shaft 12
(i.e., the output rotational speed of the engine). In response to
the throttle pedal position signal, the transmission control unit
101, either alone or in combination with the engine control unit
102, decides what rotational speed for the input shaft 12 is most
appropriate for the current conditions. For example, let it be
assumed that the throttle is moved to and maintained at a single
predetermined position. The transmission control unit 101 is
responsive to that throttle pedal position (and, in some instances,
a signal from the engine control unit 102) for determining a single
predetermined desired rotational speed for the input shaft 12.
Accordingly, the transmission control unit 101 will actuate the
trunnion actuator 103 (and, thus, the trunnion control valve 105)
to effect axial displacement of the trunnions 25, resulting in a
change in the ratio of the continuously variable drive section 20.
The transmission control unit 101 will continue to change the ratio
of the continuously variable drive section 20 until the
predetermined desired rotational speed for the input shaft 12 is
achieved.
[0067] As discussed above, the throttle pedal position often varies
considerably when the vehicle is operated at a relatively low
speed, wherein acceleration and deceleration of the vehicle is
relatively frequent. For each throttle pedal position under current
conditions, the control system 100 determines a corresponding
predetermined desired rotational speed for the input shaft 12.
Accordingly, during such relatively low speed operation, the
transmission control unit 101 will continually adjust the control
pressure in the trunnion cylinders 107. As a result, the rotational
speed of the input shaft 12 will be guided toward and maintained at
the predetermined desired rotational speed. When the vehicle is
operated at relatively high speeds, the throttle pedal position
usually varies by a lesser amount. When the throttle pedal position
is relatively constant (at any vehicle speed), the transmission
control unit 101 will maintain the same ratio in the continuously
variable drive section 20 to maintain a relatively constant desired
rotational speed for the input shaft 12. As a result, the vehicle
will be operated, at least under normal circumstances, at a
relatively constant speed.
[0068] Unlike the torque control strategy described above, the
ratio control strategy is very well suited for effecting the mode
shift described above. As mentioned above, the mode shift
preferably occurs at or near the mode point, wherein the Mode One
carrier 43 rotates at the same speed as the Mode Two ring gear 45.
When operated in the ratio control strategy, the CVT 10 is able to
precisely control when the mode point occurs because it is focused
on obtaining and maintaining the predetermined desired rotational
speed for the input shaft 12. That predetermined desired rotational
speed for the input shaft 12 can be precisely guided to the
synchronized rotational speeds of the Mode One carrier 43 and the
Mode Two ring gear 45. Thus, the mode shift will always be
performed at the mode point.
[0069] Unfortunately, however, it has been found that the ratio
control strategy can result in somewhat rough acceleration of the
vehicle, particularly when the CVT 10 is operated in the relatively
low ratios. This is because the predetermined desired rotational
speed for the input shaft 12 may change rapidly in this situation,
such as when the vehicle is being initially accelerated from a
stationary position. By focusing on ratio and ratio angle in a
system whose sensitivity is not constant, it is difficult to stay
"in phase" with the system. As a result, undesirable transient
torque can be generated in the CVT 10 as a result of the
insufficient response of the control system 100.
[0070] Traditionally, the control system 100 illustrated in FIG. 3
has been programmed to operate the CVT 10 illustrated in FIG. 1
either in the torque control strategy or in the ratio control
strategy. As discussed above, both of such control strategies have
disadvantages that can occur under certain operating conditions. To
address this shortcoming, this invention contemplates that the CVT
10 be operated in the torque control strategy during certain
operating conditions and in the ratio control strategy during other
operating conditions. In particular, with reference to FIG. 8, it
has been found to be desirable to operate the CVT 10 in the torque
control strategy at relatively low vehicle speeds, such as shown by
the box identified as the torque control region. At such low
vehicle speeds, the CVT 10 is effective as described above to
provide a very smooth acceleration, particularly when operated in
the low ratios. As indicated by the dotted line 60, the torque
control region encompasses operation of the CVT 10 only when the
Mode One clutch 46 is engaged. Thus, as described above, the CVT 10
can be operated to effect either forward, stationary, or reverse
movement of the vehicle at relatively low speeds under torque
control. Such operation is consistent with the advantageous aspects
of the torque control strategy described above.
[0071] However, as also discussed above, the ratio control strategy
is well suited for facilitating the performance of the mode shift
because of the affirmative speed control. Thus, in accordance with
this invention, as the vehicle accelerates and approaches the mode
point 62, a transition is made from the torque control strategy to
the ratio control strategy, such as shown by the box identified as
the ratio control region in FIG. 8. Preferably, the transition from
the torque control strategy to the ratio control strategy is
performed somewhat before the mode point is reached. This is done
to allow the CVT 10 to utilize the ratio control strategy to
precisely achieve the mode point, wherein the rotational speeds of
the Mode One carrier 43 and the Mode Two ring gear 45 are
synchronized. When the mode point is achieved, the mode shift is
effected, and further operation of the CVT 10 continues under ratio
control, as indicated by the solid line 61. Thus, it can be seen
that the CVT 10 is operated in such a manner as to benefit from the
advantageous aspects of both the torque and ratio control
strategies, while avoiding the disadvantageous aspect of both
strategies.
[0072] The transition from the torque control strategy to the ratio
control strategy (and vice versa) can be accomplished simply by
programming the transmission control unit 101 to switch abruptly
from one control strategy to the other control strategy at any
desired predetermined point of operation. However, to insure a
smooth transition occurs that is transparent to the driver of the
vehicle, it is desirable that this transition occur gradually over
a predetermined range of such operation, such as indicated by the
small gap between the torque control region box and the ratio
control region box in FIG. 8. To accomplish this, the transmission
control unit 101 can be programmed to simultaneously calculate the
control valve settings that would result from operation in both the
torque and ratio control strategies, and further to assign a
weighted factor or value to each of such calculated control valve
settings based upon the current operating conditions.
[0073] This method is illustrated graphically in the chart of FIG.
9. As shown therein, when the vehicle is operated at relatively low
speeds, the torque control weighted factor is high. The magnitude
of the influence of the calculated control valve setting based upon
the torque control strategy can be determined by multiplying the
calculated control valve setting based upon the torque control
strategy by its associated torque control weighted factor. Because
the torque control weighted factor is high at such low speeds, the
influence of the calculated control valve setting based upon the
torque control strategy is relatively large. Conversely, when the
vehicle is operated at relatively low speeds, the ratio control
weighted factor is low. The magnitude of the influence of the
calculated control valve setting based upon the ratio control
strategy can be determined by multiplying the calculated control
valve setting based upon the ratio control strategy by its
associated ratio control weighted factor. Because the ratio control
weighted factor is low at such low speeds, the influence of the
calculated control valve setting based upon the ratio control
strategy is relatively small. Thus, during such low speed
operation, the CVT 10 is essentially operated strictly in the
torque control strategy.
[0074] On the other hand, when the vehicle is operated at
relatively high speeds, the torque control weighted factor is low.
Because the torque control weighted factor is low at such high
speeds, the influence of the calculated control valve setting based
upon the torque control strategy is relatively small. Conversely,
when the vehicle is operated at relatively high speeds, the ratio
control weighted factor is high. Because the ratio control weighted
factor is high at such high speeds, the influence of the calculated
control valve setting based upon the ratio control strategy is
relatively large. Thus, during such high speed operation, the CVT
10 is essentially operated strictly in the ratio control
strategy.
[0075] Between these two extremes, however, a transition zone is
established, wherein the influence of the calculated control valve
setting based upon the torque control strategy (i.e., the torque
control weighted factor) is progressively decreased, while the
influence of the calculated control valve setting based upon the
ratio control strategy (i.e., the ratio control weighted factor) is
progressively increased. Thus, the CVT 10 is operated in a blended
torque/ratio control strategy during this transition, wherein the
effective control valve setting is equal to the sum of (1) the
calculated control valve setting based upon the torque control
strategy multiplied by its associated torque control weighted
factor and (2) the calculated control valve setting based upon the
ratio control strategy multiplied by its associated ratio control
weighted factor. The net result of this blended, weighted
transition is further assurance of a smooth and transparent
transition from the torque control strategy to the ratio control
strategy.
[0076] The transition between the torque and ratio control
strategies can occur at other desired points of operation beyond
the mode point. For example, if the driver rapidly depresses the
accelerator pedal while the vehicle is being operated at a
relatively high, constant speed, the ratio control strategy would
be effective to accelerate the vehicle as desired. However, it may
be desirable under such circumstances to transition back to the
torque control strategy to insure a smooth acceleration. As another
example, if the CVT 10 becomes disconnected while the vehicle is in
motion, such as might occur if both of the clutches 46 and 56 were
simultaneously disengaged, the CVT 10 can be transitioned to ratio
control strategy, even at low vehicle speeds, to match the
rotational speed of the appropriate component within the CVT 10
with the rotational speed of the output shaft 57 as one or the
other of the clutches 46 and 56 are engaged. This can prevent an
undesirable transient torque from occurring during such engagement.
An associated condition can occur at start-up of the vehicle when
the gear selector is in park or neutral. In such a situation, the
use of ratio control can insure that a transient torque is not
generated when the gear selector is moved out of park or neutral to
a ratio engaging position. The ratio control strategy may further
be implemented to maintain the operation of the CVT 10 within
certain safety limits.
[0077] Thus, the dual control strategy of the CVT 10 provides
several advantages over either of the individual single control
strategies. First, the dual control strategy provides for excellent
low speed control and acceleration of the vehicle, particularly in
the low ratios provided by the continuously variable drive section
20. Second, the dual control strategy achieves accurate speed
management to facilitate the mode shift precisely at the mode
point. Third, the dual control strategy results in superior engine
management, such as by raising or lowering the engine speed to
adjust for operating temperatures and to obtain optimum operating
conditions. Fourth, the characteristic "feel" of a torque converter
during launch of the vehicle from a stop can be replicated by
increasing the differential pressure across the control piston
107a, resulting in a slight output torque against which the driver
must brake. Lastly, the dual control strategy provides a measure of
fault protection when unusual events occur.
[0078] In accordance with the provisions of the patent statutes,
the principle and mode of operation of this invention have been
explained and illustrated in its preferred embodiment. However, it
must be understood that this invention may be practiced otherwise
than as specifically explained and illustrated without departing
from its spirit or scope.
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