U.S. patent application number 10/371378 was filed with the patent office on 2004-08-26 for method of controlling a dual clutch transmission.
Invention is credited to Buchanan, Mark, Koenig, Melissa, Lemon, Russell.
Application Number | 20040166992 10/371378 |
Document ID | / |
Family ID | 32736456 |
Filed Date | 2004-08-26 |
United States Patent
Application |
20040166992 |
Kind Code |
A1 |
Buchanan, Mark ; et
al. |
August 26, 2004 |
METHOD OF CONTROLLING A DUAL CLUTCH TRANSMISSION
Abstract
A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission to
provide predetermined engine acceleration for each gear based on
the engine throttle position and subsequent changes of clutch speed
that are in response to operatively varying the torque transferred
across the engaged clutch as the engine accelerates to a
predetermined speed based on the throttle position. The method
includes the steps of determining the engine throttle position,
determining the currently engaged gear of the transmission, and
sensing the speed of the driven member of the engaged clutch. Then,
a target engine speed is determined based on the engine throttle
position, the currently engaged gear, and the clutch speed. The
torque transferred across the engaged clutch is then continuously
varied to cause the engine to accelerate toward the target engine
speed.
Inventors: |
Buchanan, Mark; (Rochester
Hills, MI) ; Lemon, Russell; (Illawong, AU) ;
Koenig, Melissa; (Howell, MI) |
Correspondence
Address: |
BorgWarner, Inc.
Powertain Technical Center
3800 Automation Avenue, Suite 100
Auburn Hills
MI
48326
US
|
Family ID: |
32736456 |
Appl. No.: |
10/371378 |
Filed: |
February 21, 2003 |
Current U.S.
Class: |
477/181 |
Current CPC
Class: |
B60W 2510/0604 20130101;
F16H 61/688 20130101; B60W 30/18027 20130101; B60W 2540/10
20130101; F16D 2500/5048 20130101; F16H 3/0915 20130101; B60W
2710/0644 20130101; F16D 48/06 20130101; F16D 2500/7044 20130101;
F16D 2500/1026 20130101; B60W 2710/0661 20130101; F16D 2500/3061
20130101; F16H 2312/02 20130101; B60W 2510/1005 20130101; F16H
61/0437 20130101; F16H 2306/52 20130101; B60W 2510/0241 20130101;
F16D 2500/50224 20130101; F16D 2500/70454 20130101; B60W 2510/1015
20130101; F16D 2500/1086 20130101; F16H 3/006 20130101; F16D
2500/30806 20130101; F16D 2500/30426 20130101; F16D 2500/10412
20130101 |
Class at
Publication: |
477/181 |
International
Class: |
B60K 041/02 |
Claims
We claim:
1. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission to
provide a predetermined engine acceleration curve for each gear
based on the engine throttle position and subsequent changes of
clutch speed as the engine accelerates to a predetermined speed
based on the throttle position, said method includes the steps of:
sensing the engine throttle position; determining the currently
engaged gear of the transmission; sensing the speed of the driven
member of the engaged clutch; determining a target engine speed
based on the engine throttle position, the currently engaged gear,
and the clutch speed; and continuously varying the torque
transferred across the engaged clutch to cause the engine to
accelerate toward the target engine speed.
2. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 1 wherein the step of increasing the torque
transferred across the clutch further includes the steps of:
determining a dynamic torque signal that represents the amount of
torque that must be transferred across the engaged clutch to cause
the engine to track the target engine speed in response to the
throttle position; determining a steady-state torque signal that
represents an additional amount of torque that must be transferred
across the engaged clutch to compensate for changes in driving
conditions; summing the dynamic torque signal and the steady-state
torque signal to provide a target torque signal that represents the
total amount of torque that must be transferred across the engaged
clutch to cause the engine to track the target engine speed in
response to the throttle position and changes in the driving
conditions; and varying the torque transferred across the clutch to
cause the engine to track the target torque signal.
3. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 2 wherein the step of determining a dynamic
torque signal further includes the steps of: determining the actual
engine speed; comparing the target engine speed to the actual
engine speed and producing a speed error signal; and proportioning
the value of the speed error signal to a relative value within a
predetermined range to provide the dynamic torque signal.
4. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 2 wherein the step of determining a steady-state
torque signal further includes the steps of: determining the
acceleration of the engine; determining the acceleration of the
engaged clutch; summing the engine acceleration and the clutch
acceleration to determine the clutch slip acceleration; determining
the actual dynamic torque transferred across the clutch from the
clutch slip acceleration; predicting a dynamic torque value that
should be transferred across the engaged clutch for the given
target torque signal; determining a dynamic torque error by summing
the actual dynamic torque and the predicted dynamic torque value,
wherein this sum indicates that changes in driving conditions have
occurred that must be compensated for by making a correction to the
target torque signal; and determining the steady-state torque
signal by integrating the dynamic torque error, wherein this
integration is the corrective value.
5. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 4 wherein the step of determining a steady-state
torque signal further includes the steps of: predicting the value
of the steady state torque signal based on inputs from an engine
controller when a change in throttle position results in a new
engine target speed; predicting the delay in the engine reaching
the new target engine based on known engine response times to
changes in the target torque signal; and summing the predicted
delay with the integrated dynamic torque error signal to provide
the steady-state torque signal.
6. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 5 wherein the step of predicting a dynamic
torque value further includes: predicting the value of the total
torque that should be transferred across the engaged clutch given
the target torque signal, based on known transmission responses to
the input of the target torque signal; subtracting a feedback
signal of the steady state torque value from the predicted total
torque value to provide a predicted dynamic torque value.
7. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 1 wherein said method further includes the steps
of: determining the currently engaged gear of the transmission and
sensing the clutch speed when a reduction in the engine throttle
position is detected; decreasing the pressure as applied to the
engaged clutch as the engine decelerates thereby controlling the
rate of vehicle deceleration; determining an initial decelerating
target engine speed based on the currently engaged gear and the
engine throttle position; varying the torque transferred across the
engaged clutch to cause the engine speed to decelerate toward the
initial decelerating target engine speed; incrementally
redetermining the target engine speed after the clutch speed
reaches a predetermined percentage of the initial target engine
speed due to the clutch response to the decreasing engine speed
thereby causing the engine speed to track the decreasing target
engine speed of said engine deceleration curve; and continuously
varying the torque transferred across the engaged clutch to cause
the engine to decelerate along said engine deceleration curve.
8. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission to
provide a predetermined engine acceleration curve for each gear
based on the engine throttle position and subsequent changes of
clutch speed in response to operatively increasing the torque
transferred across the engaged clutch as the engine accelerates to
a predetermined speed based on the throttle position, said method
includes the steps of: determining the engine throttle position;
determining the currently engaged gear of the transmission; sensing
the speed of the driven member of the engaged clutch; selecting an
engine stall speed for the current gear and engine throttle
position at which an increasing transfer of torque across the
engaged clutch will stop further engine acceleration and will hold
the engine to a constant speed, from a look-up table; determining a
target engine speed based on the engine stall speed; increasing the
torque transferred across the engaged clutch so as to cause the
actual engine speed to increase and approach the target engine
speed; continuously redetermining the target engine after the
clutch speed reaches a predetermined percentage of the engine stall
speed, the clutch speed increasing in response to the increasing
engine speed thereby causing the engine speed to track the rising
target engine speed of said engine acceleration curve; and
increasing the torque transferred across the engaged clutch so as
to operatively lock the clutch and engine together as the clutch
speed reaches a predetermined multiple of the engine stall
speed.
9. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 8 wherein the step of continually redetermining
the target engine speed further includes the steps of: determining
a dynamic torque signal that represents the amount of torque that
must be transferred across the engaged clutch to cause the engine
to track the target engine speed in response to the throttle
position; determining a steady-state torque signal that represents
an additional amount of torque that must be transferred across the
engaged clutch to compensate for changes in driving conditions;
summing the dynamic torque signal and the steady-state torque
signal to provide a target torque signal that represents the total
amount of torque that must be transferred across the engaged clutch
to cause the engine to track the target engine speed in response to
the throttle position and changes in the driving conditions; and
varying the torque transferred across the clutch to cause the
engine to track the target torque signal.
10. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 9 wherein the step of determining a dynamic
torque signal further includes the steps of: determining actual
engine speed; comparing the target engine speed to the actual
engine speed and producing a speed error signal; and proportioning
the value of the speed error signal to a relative value within a
predetermined range to provide the dynamic torque signal.
11. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 9 wherein the step of determining a steady-state
torque signal further includes the steps of: determining the
acceleration of the engine; determining the acceleration of the
engaged clutch; summing the engine acceleration and the clutch
acceleration to determine the clutch slip acceleration; determining
the actual dynamic torque transferred across the clutch from the
clutch slip acceleration; predicting a dynamic torque value that
should be transferred across the engaged clutch for the given
target torque signal; determining a dynamic torque error by summing
the actual dynamic torque and the predicted dynamic torque value,
wherein this sum indicates that changes in driving conditions have
occurred that must be compensated for by making a correction to the
target torque signal; and determining the steady-state torque
signal by integrating the dynamic torque error, wherein this
integration is the corrective value.
12. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 11 wherein the step of determining a
steady-state torque signal further includes the steps of:
predicting the value of the steady state torque signal based on
inputs from an engine controller when a change in throttle position
results in a new engine target speed; predicting the delay in the
engine reaching the new target engine, based on known engine
response times to changes in the target torque signal; and summing
the predicted delay with the integrated dynamic torque error signal
to provide the steady-state torque signal.
13. A method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission as
set forth in claim 11 wherein the step of predicting a dynamic
torque value further includes: predicting the value of the total
torque that should be transferred across the engaged clutch given
the target torque signal based on known transmission responses to
the input of the target torque signal; subtracting a feedback
signal of the steady state torque value from the predicted total
torque value to provide a predicted dynamic torque value.
14. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission to control
the torque transferred across the engaged clutch thereby providing
a predetermined engine acceleration curve for each gear based on
the engine throttle position and subsequent changes of clutch speed
in response to operatively varying the pressure on the engaged
clutch as the engine accelerates to a predetermined speed based on
the throttle position, said method includes the steps of:
determining the engine throttle position; determining the currently
engaged gear of the transmission; sensing the speed of the driven
member of the engaged clutch; selecting an engine stall speed for
the current gear and engine throttle position, at which the
increasing pressure applied to the engaged clutch will stop further
engine acceleration and will hold the engine to a constant speed,
from a look-up table; increasing the torque transferred across the
engaged clutch so as to cause the engine speed to increase and
approach the target engine speed thereby increasing the clutch
speed and accelerating the vehicle; reducing the pressure applied
to the engaged clutch allowing the engine to continue to accelerate
toward the target engine speed when the clutch speed reaches a
predetermined percentage of the engine stall speed thereby
continuing to increase the clutch speed and accelerating the
vehicle; continuously redetermining a target engine speed based on
the engine stall speed and the clutch speed using a target engine
speed equation that influences the rate of increase of engine and
clutch speed after the clutch speed reaches the predetermined
percentage of the engine stall speed, said target engine speed
equation defined as: 3 T = S + { [ C + [ ( K - 2 ) * S ] } 2 4 * (
K - 1 ) * S wherein T is the target engine speed, S is the stall
speed determined for the current gear and throttle position, C is
the clutch speed, and K is a control constant, the clutch speed
increasing in response to the increasing engine speed thereby
causing the engine speed to track the rising target engine speed;
and increasing the applied pressure on the engaged clutch so as to
operatively lock the clutch and engine together as the clutch speed
reaches a predetermined multiple of the engine stall speed.
15. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission as set forth
in claim 14 wherein the step of reducing the pressure applied to
the engaged clutch when the clutch speed reaches a predetermined
percentage of the engine stall speed further includes the defining
the predetermined percentage as the relationship ([2-K]*S), where K
is the chosen control constant and S is the stall speed for the
given throttle position.
16. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission as set forth
in claim 16 wherein the step of increasing the applied pressure on
the engaged clutch so as to operatively lock the clutch and engine
together as the clutch speed reaches a predetermined multiple of
the engine stall speed further includes defining the predetermined
multiple of the engine stall speed as the relationship (K*S), where
K is the chosen control constant and S is the stall speed for the
given throttle position.
17. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission as set forth
in claim 14 wherein the method further includes the step of
increasing the applied pressure on the engaged clutch as the clutch
speed reaches a predetermined multiple of the engine stall speed so
as to prevent locking of the clutch and engine together but allow a
small predetermined amount of slip to occur to provide for a smooth
transfer of motive force between the clutch and engine.
18. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission as set forth
in claim 14 further includes a the steps of: performing a
preparatory clutch pressure fill without a transfer of torque to
the clutch that will be the engaged clutch for either first gear,
or the gear immediately above or below the currently engaged gear,
said preparatory clutch pressure fill performed to first gear
clutch when the vehicle is stationary and performed to any other
gear immediately prior to a shift; and determining a standby
portion of clutch actuation where no engine acceleration has yet
occurred but said standby portion is representative of a period
prior to the transfer of torque across the clutch that will drive
the particular gear.
19. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission as set forth
in claim 14 further includes a the steps of: performing a
predetermined nominal increase to the pressure applied to the
clutch that drives first gear after the standby portion for first
gear when first gear is engaged, said predetermined nominal
pressure increase performed so as to cause a slight forward creep
of the vehicle in anticipation of increased torque transfer across
the clutch to drive the vehicle forward; and determining a creep
portion for first gear that is representative of said predetermined
nominal pressure increase to the clutch, where only slight engine
acceleration has occurred and said creep portion is representative
of the period immediately prior to the transfer of torque across
the clutch that will drive the first gear.
20. A method of controlling the pressure applied to the engaged
clutch of a vehicle having a dual clutch transmission as set forth
in claim 14 wherein said method further includes the steps of:
determining the currently engaged gear of the transmission and
sensing the clutch speed when a reduction in the engine throttle
position is detected; decreasing the pressure as applied to the
engaged clutch initially in a linear manner as the engine
decelerates thereby controlling the rate of vehicle deceleration;
selecting an engine stall speed for the current gear and engine
throttle position at which the increasing pressure applied to the
engaged clutch will stop further engine acceleration and will hold
the engine to a constant speed; determining a target engine speed
based on the engine stall speed and the clutch speed using a target
engine speed equation that influences the rate of decrease of
clutch engagement pressure, said target engine speed equation
defined as: 4 T = S + { [ C + [ ( K - 2 ) * S ] } 2 4 * ( K - 1 ) *
S wherein T is the target engine speed, S is the stall speed
determined for the current gear and throttle position, C is the
clutch speed, and K is a control constant; controlling the
reduction in the pressure applied to the engaged clutch causing the
engine speed to decelerate toward the target engine speed when the
clutch speed reaches a predetermined percentage of the engine stall
speed thereby continuing to decrease the clutch speed and
decelerating the vehicle; and continuously redetermining the target
engine speed to create said engine deceleration curve after the
clutch speed reaches the predetermined percentage of the engine
stall speed due to the decrease in clutch speed in response to the
decreasing engine speed thereby causing the engine speed to track
the decreasing target engine speed of said engine deceleration
curve.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] The present invention relates, generally to the control of a
dual clutch transmission and, more specifically, to a method for
controlling the acceleration of an engine of a motor vehicle by
controlling the torque transfer of the clutches of a dual clutch
transmission.
[0003] 2. Description of the Related Art
[0004] Generally speaking, land vehicles require a powertrain
consisting of three basic components. These components include a
power plant (such as an internal combustion engine), a power
transmission, and wheels. The power transmission component is
typically referred to simply as the "transmission." Engine torque
and speed are converted in the transmission in accordance with the
tractive-power demand of the vehicle. Presently, there are two
typical transmissions widely available for use in conventional
motor vehicles. The first, and oldest type is the manually operated
transmission. These transmissions include a foot operated start-up
or launch clutch to engage and disengage the driveline with the
power plant and a gearshift lever to selectively change the gear
ratios within the transmission. When driving a vehicle having a
manual transmission, the driver must coordinate the operation of
the clutch pedal, the gearshift lever and the accelerator pedal to
achieve a smooth and efficient shift from one gear to the next. The
structure of a manual transmission is simple and robust and
provides good fuel economy by having a direct power connection from
the engine to the final drive wheels of the vehicle. Additionally,
since the operator is given complete control over the timing of the
shifts, the operator is able to dynamically adjust the shifting
process so that the vehicle can be driven most efficiently.
However, one disadvantage of the manual transmission is that there
is an interruption in the drive connection during gear shifting. In
addition, there is a great deal of physical interaction required on
the part of the operator to shift gears in a manually operated
transmission.
[0005] The second type of transmission employed in a conventional
motor vehicle is an automatic transmission. Automatic transmissions
offer ease of operation. The driver of a vehicle having an
automatic transmission is not required to use both hands, one for
the steering wheel and one for the gearshift, and both feet, one
for the clutch and one for the accelerator and brake pedal in order
to safely operate the vehicle. In addition, an automatic
transmission provides greater convenience in stop and go
situations, because the driver is not concerned about continuously
shifting gears to adjust to the ever-changing speed of traffic.
Although conventional automatic transmissions avoid an interruption
in the drive connection during gear shifting, they suffer from the
disadvantage of reduced efficiency because of the need for
hydrokinetic devices, such as torque converters, interposed between
the output of the engine and the input of the transmission for
transferring kinetic energy therebetween.
[0006] At low speed ratios, RPM output/RPM input, torque converters
multiply or increase the torque translation from the engine. During
torque multiplication, the output torque is greater than the input
torque for the torque converter. However, at high speed ratios
there is no torque multiplication and the torque converter becomes
a fluid coupling. Fluid couplings have inherent slip. Torque
converter slip exists when the speed ratio is less than 1.0 (RPM
input>than RPM output of the torque converter). The inherent
slip reduces the efficiency of the torque converter.
[0007] While torque converters provide a smooth coupling between
the engine and the transmission, the slippage of the torque
converter results in a parasitic loss, thereby decreasing the
efficiency of the entire powertrain. Further, the torque converter
itself requires pressurized hydraulic fluid in addition to any
pressurized fluid requirements for the actuation of the gear
shifting operations. This means that an automatic transmission must
have a large capacity pump to provide the necessary hydraulic
pressure for both converter engagement and shift changes. The power
required to drive the pump and pressurize the fluid introduces
additional parasitic losses of efficiency in the automatic
transmission.
[0008] In an ongoing attempt to provide a vehicle transmission that
has the advantages of both types of transmissions with fewer of the
drawbacks, combinations of the traditional "manual" and "automatic"
transmissions have evolved. Most recently, "automated" variants of
conventional manual transmissions have been developed which shift
automatically without any input from the vehicle operator. Such
automated manual transmissions typically include a plurality of
power-operated actuators that are controlled by a transmission
controller or some type of electronic control unit (ECU) to
automatically shift synchronized clutches that control the
engagement of meshed gear wheels traditionally found in manual
transmissions. The design variants have included either
electrically or hydraulically powered actuators to affect the gear
changes. However, even with the inherent improvements of these
newer automated transmissions, they still have the disadvantage of
a power interruption in the drive connection between the input
shaft and the output shaft during sequential gear shifting. Power
interrupted shifting results in a harsh shift feel that is
generally considered to be unacceptable when compared to the smooth
shift feel associated with most conventional automatic
transmissions.
[0009] To overcome this problem, other automated manual type
transmissions have been developed that can be power-shifted to
permit gearshifts to be made under load. Examples of such
power-shifted automated manual transmissions are shown in U.S. Pat.
No. 5,711,409 issued on Jan. 27, 1998 to Murata for a Twin-Clutch
Type Transmission, and U.S. Pat. No. 5,966,989 issued on Apr. 04,
2000 to Reed, Jr. et al for an Electro-mechanical Automatic
Transmission having Dual Input Shafts. These particular variant
types of automated manual transmissions have two clutches and are
generally referred to simply as dual, or twin, clutch
transmissions. The dual clutch structure is most often coaxially
and cooperatively configured so as to derive power input from a
single engine flywheel arrangement. However, some designs have a
dual clutch assembly that is coaxial but with the clutches located
on opposite sides of the transmissions body and having different
input sources. Regardless, the layout is the equivalent of having
two transmissions in one housing, namely one power transmission
assembly on each of two input shafts concomitantly driving one
output shaft. Each transmission can be shifted and clutched
independently. In this manner, uninterrupted power upshifting and
downshifting between gears, along with the high mechanical
efficiency of a manual transmission is available in an automatic
transmission form. Thus, significant increases in fuel economy and
vehicle performance may be achieved through the effective use of
certain automated manual transmissions.
[0010] The dual clutch transmission structure may include two disc
clutches each with its own clutch actuator to control the
engagement and disengagement of the two clutches independently of
one another. While the clutch actuators may be of the
electromechanical type, since a lubrication system within the
transmission is still a necessity and therefore require a pump,
some dual clutch transmissions utilize hydraulic shifting and
clutch control. These pumps are most often gerotor types, and are
much smaller than those used in conventional automatic
transmissions because they typically do not have to supply a torque
converter. Thus, parasitic losses are kept small. Shifts are
accomplished by engaging the desired gear prior to a shift event
and subsequently engaging the corresponding clutch. With two
clutches and two inputs shafts, at certain times, the dual clutch
transmission may be in two different gear ratios at once, but only
one clutch will be engaged and transmitting power at any given
moment. To shift to the next higher gear, first the desired gears
on the input shaft of the non-driven clutch assembly are engaged,
then the driven clutch is released and the non-driven clutch is
engaged.
[0011] This requires that the dual clutch transmission be
configured to have the forward gear ratios alternatingly arranged
on their respective input shafts. In other words, to perform
up-shifts from first to second gear, the first and second gears
must be on different input shafts. Therefore, the odd gears will be
associated with one input shaft and the even gears will be
associated with the other input shaft. In view of this convention,
the input shafts are generally referred to as the odd and even
shafts. Typically, the input shafts transfer the applied torque to
a single counter shaft, which includes mating gears to the input
shaft gears. The mating gears of the counter shaft are in constant
mesh with the gears on the input shafts. The counter shaft also
includes an output gear that is meshingly engaged to a gear on the
output shaft. Thus, the input torque from the engine is transferred
from one of the clutches to an input shaft, through a gear set to
the counter shaft and from the counter shaft to the output
shaft.
[0012] Gear engagement in a dual clutch transmission is similar to
that in a conventional manual transmission. One of the gears in
each of the gear sets is disposed on its respective shaft in such a
manner so that it can freewheel about the shaft. A synchronizer is
also disposed on the shaft next to the freewheeling gear so that
the synchronizer can selectively engage the gear to the shaft. To
automate the transmission, the mechanical selection of each of the
gear sets is typically performed by some type of actuator that
moves the synchronizers. A reverse gear set includes a gear on one
of the input shafts, a gear on the counter shaft, and an
intermediate gear mounted on a separate counter shaft meshingly
disposed between the two so that reverse movement of the output
shaft may be achieved.
[0013] While these power-shift dual clutch transmissions overcome
several drawbacks associated with conventional transmissions and
the newer automated manual transmissions, it has been found that
controlling and regulating the automatically actuated dual clutch
transmissions is a complicated matter and that the desired vehicle
occupant comfort goals have not been achievable in the past. There
are a large number of events that must be properly timed and
executed within the transmission to achieve smooth and efficient
operation, not only during the power-shifting events but also
throughout the entire operating range of the transmission as well.
To this point, conventional control schemes and methods have
generally failed to provide this capability. Accordingly, there
exists a need in the related art for better methods of controlling
the operation of dual clutch transmissions.
[0014] One particular area of improvement that is needed is in the
control of engine acceleration through the control of the torque
transferred across of the clutches of the transmission. The nature
of the dual clutch transmission, that is, the manual style
configuration discussed above that employs automatically actuated
disc type clutches, requires accurate control of the clutch
engagement and thus the torque transferred across them. More
specifically, it is desirable to operate the clutches of the dual
clutch transmission so that acceleration of the engine, and thus
the vehicle, through each of its gears is controlled by varying the
amount of torque transferred across the clutch, or in other words
to induce certain amounts of clutch slip in certain parts of the
vehicle's operating range.
[0015] Control of the torque transferred across the clutches, and
thus the control of the engine acceleration, is required to provide
smooth operation, avoid hard or noticeable lockup of the clutch to
the transmission, and to provide efficient engine-to-transmission
interaction. The control schemes for dual clutch transmissions
known in the related art are incapable of adequately providing for
fine control of engine acceleration to satisfy this need.
Specifically, they lack the ability to finely control the torque
transferred across the clutches to achieve the high degree of
accuracy needed for smooth transmission and engine operation.
Additionally, current control methods for the clutches of a dual
clutch transmission generally concern themselves with simple
engagement and disengagement of the clutch assemblies and fail to
adequately provide for the corresponding control of all aspects and
phases of engine acceleration control, vehicle movement, and the
preparation for movement.
[0016] In that regard, there are some prior art clutch control
methods for dual clutch transmissions of a more specialized nature
that are referred to as launch methods or strategies. Typically,
these methods are directed to limited control over engine
acceleration and vehicle movement from a standing start to initial
portion of the vehicle's operating range. These conventional launch
strategies are moderately successful at getting the vehicle moving
in an acceptable manner. However, the term launch may also be
extended, particularly in the case of the dual clutch transmission,
to encompass not only the movement of the vehicle from a standing
start but also to refer to all clutch activity (i.e. torque
transfer) within each of the available gears, outside of a
gearshift operation. Thus, in this greater context the term
"launch" as used here means the clutch control of the torque
transfer in each gear, so that the control of the engine
acceleration (and deceleration) through each gear of the
transmission (not including the shifts) follows a launch strategy.
This highlights a further inadequacy in the prior art dual
transmission clutch control schemes. Specifically, the clutch
control schemes generally known in the related art are primarily
designed to move a stationary vehicle and are incapable of
adequately providing for control of the torque transfer for all
activities of the clutches and are not designed to "launch" each of
the gears of the transmission. This inadequacy, like others
discussed above, also directly effects the smoothness and
efficiency of the entire powertrain of the vehicle.
[0017] For example, some current dual clutch transmission control
methods cannot prevent adverse engine responses to clutch
engagement, such as engine lugging or over-revving. The lugging
effect occurs when a clutch is heavily engaged without adequate
engine speed, and the engine becomes excessively loaded, causing
surges and roughness. The over-revving effect occurs when the
clutch engagement is slow and behind the acceleration of the engine
so that the slip is excessive and power is lost. Other current
control methods that overcome lugging and over-revving for first
gear and the initial start of the vehicle do not prevent these
problems in any other gear or in any other part of the vehicle
operating range. Accordingly, there remains a need in the art for a
method to operatively and actively control the launch of each of
the gears of the dual clutch transmission by providing control over
all the torque transfer activities of the clutches.
SUMMARY OF THE INVENTION
[0018] The disadvantages of the related art are overcome by the
method of the present invention for controlling the engine speed of
a vehicle having a dual clutch transmission. More specifically, the
present invention is directed to a method of controlling the torque
transferred across the engaged clutch of a vehicle having a dual
clutch transmission to provide a predetermined engine acceleration
curve for each gear based on the engine throttle position and
subsequent changes of clutch speed that are in response to
operatively varying the torque transferred across the engaged
clutch as the engine accelerates to a predetermined speed based on
the throttle position. The method includes the steps of determining
the engine throttle position, determining the currently engaged
gear of the transmission, and sensing the speed of the driven
member of the engaged clutch. Then, a target engine speed is
determined based on the engine throttle position, the currently
engaged gear, and the clutch speed. The torque transferred across
the engaged clutch is then continuously varied to cause the engine
to accelerate toward the target engine speed. Thus, the engaged
clutch is operatively slipped to smoothly cause the engine to track
a predetermined target speed, providing efficient operation with
the desired smooth driving feel. In this way, hard clutch lock-ups
and engine lugging are avoided.
[0019] Other objects, features and advantages of the present
invention will be readily appreciated as the same becomes better
understood after reading the subsequent description taken in
connection with the accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
[0020] FIG. 1 is a generalized schematic illustration of a dual
clutch transmission as controlled by the method steps of the
present invention;
[0021] FIG. 2 is a schematic illustration of the electro-hydraulic
control circuit for the clutch actuators of a dual clutch
transmission controlled by the method steps of the present
invention;
[0022] FIG. 3 is a block diagram flowchart of the method of the
present invention for controlling a dual clutch transmission;
[0023] FIG. 4 is a block diagram flowchart of the method of the
present invention for controlling a dual clutch transmission;
[0024] FIG. 5 is a block diagram flowchart of an additional
embodiment of the method of the present invention for controlling a
dual clutch transmission; and
[0025] FIG. 6 is a graph of the acceleration curve of the method of
the present invention for controlling a dual clutch
transmission.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT(S)
[0026] A representative dual clutch transmission that may be
controlled by the present invention is generally indicated at 10 in
the schematic illustrated in FIG. 1. Specifically, as shown in FIG.
1, the dual clutch transmission 10 includes a dual, coaxial clutch
assembly generally indicated at 12, a first input shaft, generally
indicated at 14, a second input shaft, generally indicated at 16,
that is coaxial to the first, a counter shaft, generally indicated
at 18, an output shaft 20, a reverse counter shaft 22, and a
plurality of synchronizers, generally indicated at 24.
[0027] The dual clutch transmission 10 forms a portion of a vehicle
powertrain and is responsible for taking a torque input from a
prime mover, such as an internal combustion engine and transmitting
the torque through selectable gear ratios to the vehicle drive
wheels. The dual clutch transmission 10 operatively routes the
applied torque from the engine through the dual, coaxial clutch
assembly 12 to either the first input shaft 14 or the second input
shaft 16. The input shafts 14 and 16 include a first series of
gears, which are in constant mesh with a second series of gears
disposed on the counter shaft 18. Each one of the first series of
gears interacting with one of the second series of gears to provide
the different gear ratios sets used for transferring torque. The
counter shaft 18 also includes a first output gear that is in
constant mesh with a second output gear disposed on the output
shaft 20. The plurality of synchronizers 24 are disposed on the two
input shafts 14, 16 and on the counter shaft 18 and are operatively
controlled by the plurality of shift actuators (not shown) to
selectively engage one of the gear ratio sets. Thus, torque is
transferred from the engine to the dual, coaxial clutch assembly
12, to one of the input shafts 14 or 16, to the counter shaft 18
through one of the gear ratio sets, and to the output shaft 20. The
output shaft 20 further provides the output torque to the remainder
of the powertrain. Additionally, the reverse counter shaft 22
includes an intermediate gear that is disposed between one of the
first series of gears and one of the second series of gears, which
allows for a reverse rotation of the counter shaft 18 and the
output shaft 20. Each of these components will be discussed in
greater detail below.
[0028] Specifically, the dual, coaxial clutch assembly 12 includes
a first clutch mechanism 32 and a second clutch mechanism 34. The
first clutch mechanism 32 is, in part, physically connected to a
portion of the engine flywheel (not shown) and is, in part,
physically attached to the first input shaft 14, such that the
first clutch mechanism 32 can operatively and selectively engage or
disengage the first input shaft 14 to and from the flywheel.
Similarly, the second clutch mechanism 34 is, in part, physically
connected to a portion of the flywheel and is, in part, physically
attached to the second input shaft 16, such that the second clutch
mechanism 34 can operatively and selectively engage or disengage
the second input shaft 16 to and from the flywheel. As can be seen
from FIG. 1, the first and second clutch mechanisms 32, 34 are
coaxial and co-centric such that the outer case 28 of the first
clutch mechanism 32 fits inside of the outer case 36 of the second
clutch mechanism 34. Similarly, the first and second input shafts
14, 16 are also coaxial and co-centric such that the second input
shaft 16 is hollow having an inside diameter sufficient to allow
the first input shaft 14 to pass through and be partially supported
by the second input shaft 16. The first input shaft 14 includes a
first input gear 38 and a third input gear 42. The first input
shaft 14 is longer in length than the second input shaft 16 so that
the first input gear 38 and a third input gear 42 are disposed on
the portion of the first input shaft 14 that extends beyond the
second input shaft 16. The second input shaft 16 includes a second
input gear 40, a fourth input gear 44, a sixth input gear 46, and a
reverse input gear 48. As shown in FIG. 1, the second input gear 40
and the reverse input gear 48 are fixedly disposed on the second
input shaft 16 and the fourth input gear 44 and sixth input gear 46
are rotatably supported about the second input shaft 16 upon
bearing assemblies 50 so that their rotation is unrestrained unless
the accompanying synchronizer is engaged, as will be discussed in
greater detail below.
[0029] In the preferred embodiment, the counter shaft 18 is a
single, one-piece shaft that includes the opposing, or counter,
gears to those on the inputs shafts 14, 16. As shown in FIG. 1, the
counter shaft 18 includes a first counter gear 52, a second counter
gear 54, a third counter gear 56, a fourth counter gear 58, a sixth
counter gear 60, and a reverse counter gear 62. The counter shaft
18 fixedly retains the fourth counter gear 58 and counter gear 60,
while first, second, third, and reverse counter gears 52, 54, 56,
62 are supported about the counter shaft 18 by bearing assemblies
50 so that their rotation is unrestrained unless the accompanying
synchronizer is engaged as will be discussed in greater detail
below. The counter shaft 18 also fixedly retains a first drive gear
64 that meshingly engages the corresponding second driven gear 66
on the output shaft 20. The second driven gear 66 is fixedly
retained on the output shaft 20. The output shaft 20 extends
outward from the transmission 10 to provide an attachment for the
remainder of the powertrain.
[0030] In the preferred embodiment, the reverse counter shaft 22 is
a relatively short shaft having a single reverse intermediate gear
72 that is disposed between, and meshingly engaged with, the
reverse input gear 48 on the second input shaft 16 and the reverse
counter gear 62 on the counter shaft 18. Thus, when the reverse
gear 48, 62, and 72 are engaged, the reverse intermediate gear 72
on the reverse counter shaft 22 causes the counter shaft 18 to turn
in the opposite rotational direction from the forward gears thereby
providing a reverse rotation of the output shaft 20. It should be
appreciated that all of the shafts of the dual clutch transmission
10 are disposed and rotationally secured within the transmission 10
by some manner of bearing assembly such as roller bearings, for
example, shown at 68 in FIG. 1.
[0031] The engagement and disengagement of the various forward and
reverse gears is accomplished by the actuation of the synchronizers
24 within the transmission. As shown in FIG. 1 in this example of a
dual clutch transmission 10, there are four synchronizers 74,
76,78, and 80 that are utilized to shift through the six forward
gears and reverse. It should be appreciated that they are a variety
of known types of synchronizers that are capable of engaging a gear
to a shaft and that the particular type employed for the purposes
of this discussion is beyond the scope of the present invention.
Generally speaking, any type of synchronizer that is movable by a
shift fork or like device may be employed. As shown in the
representative example of FIG. 1, the synchronizers are two sided,
dual actuated synchronizers, such that they engage one gear to its
shaft when moved off of a center neutralized position to the right
and engage another gear to its shaft when moved to the left.
[0032] It should be appreciated that the operation of the dual
clutch transmission 10 is managed by some type of control device
such as an electronic control unit (ECU) that oversees the
functioning of the transmission 10, or by an electronic control
unit for the vehicle in which the dual clutch transmission 10 may
be installed. Regardless, there exists a control device, beyond the
scope of this invention, that controls and operates the dual clutch
transmission through a stored control scheme or series of control
schemes of which the present invention is merely a part. The
control device having the capability of providing the proper
voltages, signals, and/or hydraulic pressures to operate the
transmission 10 and particularly the clutch engagement functions.
Thus, the control method of the present invention as described
below may be a standalone process or merely a portion, such as a
sub-routine, or series of sub-routines, of a larger control scheme
within the ECU.
[0033] The first and second clutch mechanisms 32 and 34 of the
dual, coaxial clutch assembly 12 are operatively engaged and
disengaged in a coordinated manner relative to the actuator of the
various gear sets by the synchronizer 24 to selectively transfer
torque to the output shaft 20. By way of example, if torque is
being transferred to the drive wheels of the vehicle to initiate
movement from a standing start, the lowest, or first, gear ratio of
the dual clutch transmission 10 will likely be engaged. Therefore,
as seen in FIG. 1, synchronizer 78 will be driven to the left to
engage the first counter gear 52 to the counter shaft 18 and the
first clutch mechanism 32 will be engaged to transfer torque from
the engine to the output shaft 20 through the first gear set. When
vehicle speed increases and the ECU determines that the conditions
require a shift to the second gear set, synchronizer 80 will first
be driven to the right to engage the second counter gear 54 to the
counter shaft 18. Then the second clutch mechanism 34 will be
engaged as the first clutch mechanism 32 is disengaged. In this
manner, a powershift, where no power interruption occurs, is
effected. Additionally, while engaged and driving a particular
gear, the first and second clutch mechanisms 32 and 34 are
controlled by certain stored routines that provide varying amounts
of engagement force to the clutch discs and thereby operatively
control the amount of torque transferred across the clutches and
the resultant engine speed. Of particular concern to this
application is the speed control routine that causes the engine
speed to track a predetermined target speed for given input
parameters by varying the applied engagement pressure across the
clutch discs. In that regard, the actuating components of the first
and second clutch mechanisms 32 and 34 are not shown and it should
be appreciated there may be of any number of suitable known devices
that are capable of selectively varying the applied engagement
pressure between the clutch discs, such as, but not limited to
mechanical actuators, hydro-mechanical actuators, electromechanical
actuators, or fully electrical actuators.
[0034] For example, in one embodiment of the dual clutch
transmission 10, the first and second clutch mechanisms 32 and 34
of the dual, coaxial clutch assembly 12 are actuated by hydraulic
pressure supplied by the first and second clutch actuator
solenoids, respectively. The clutch actuator solenoids are
schematically represented, and generally indicated at 120 and 122
in FIG. 2, and as shown, are supplied with pressurized hydraulic
fluid by a regulating circuit generally indicated at 82. It should
be appreciated that, as previously mentioned, the actuation of the
components of the dual clutch transmission 10 may be electrical
rather than electro-hydraulic, and in that case, the first and
second clutch actuator solenoids 120, 122 would be replaced by some
type of physical drive devices to operatively engage the first and
second clutch mechanisms 32 and 34.
[0035] As shown in FIG. 2, for this example of a dual clutch
transmission 10, there are two on/off solenoids, generally
indicated at 124 and 126, and two enable valves, generally
indicated at 128 and 130 that provide the operative hydraulic
pressure to the clutch actuator solenoids 120 and 122. A main
pressure supply line 92 that is operatively connected to a source
of pressurized hydraulic fluid from a pump within the transmission
10 (not shown) provides the two on/off solenoids 124 and 126 with
pressurized hydraulic fluid. The on/off solenoids 124 and 126 each
have a selectively movable valve member 134 disposed within a valve
body 136 that has internal hydraulic flow passages 138 and 140.
When energized, the valve members 134 of the on/off solenoids 124
and 126 are driven to the left, as illustrated, by actuators 142
and 144 respectively. The on/off solenoids 124 and 126 then
selectively provide hydraulic pressure though pressure lines 148
and 150 to act upon the right sides of enable valves 128 and 130,
as illustrated in FIG. 2. In their normally de-energized state,
biasing member 152 causes the valve member 134 to be driven back to
the right and any residual pressure in pressure lines 148 or 150 is
bled off and routed back to the fluid sump, shown at 90.
[0036] The enable valves 128 and 130 also each have a selectively
movable valve member 154 disposed within a valve body 156 that has
internal hydraulic flow passages 158 and 160. The applied hydraulic
pressure from the on/off solenoids 124 and 126 act to push the
valve members 154 of the enable valves 128 and 130 to the left to
open the internal hydraulic passage 158 and provide hydraulic
pressure to clutch actuator solenoid 120 and 122 through the
pressure supply lines 162 and 164. In their normally de-energized
state biasing member 166 causes the valve member 154 to be driven
back to the right and any residual pressure in pressure lines 162
or 164 is bled off and routed back to the fluid sump, shown at
90.
[0037] Though beyond the scope of this invention and not shown
here, the two enable valves 128 and 130 are also in fluid
communication with, and hydraulically feed, the synchronizer
actuator solenoids that drive the synchronizers 24 of the
transmission 10 between their engaged and neutralized positions.
Thus, it should be appreciated that two on/off solenoids 124 and
126, and two enable valves 128 and 130 also have other hydraulic
switching functions within the transmission 10, such that the
on/off solenoids 124 and 126 are selectively operable to provide
and remove hydraulic actuating pressure and prevent uncontrolled
actuation of the mechanisms within the transmission 10.
[0038] When the on/off solenoids 124 and 126 are actuated and the
enable valves 128 and 130 have charged the pressure supply lines
162 and 164 to the clutch actuator solenoids 120 and 122, the first
and second clutch mechanisms, generally indicated at 32 and 34, are
controllable. The clutch actuator solenoids 120 and 122 are in
fluid communication with the clutch mechanisms 32 and 34 through
clutch pressure lines 170 and 172 respectively. Each of the clutch
actuator solenoids 120 and 122 have a selectively movable valve
member 176 disposed within a valve body 178 that has internal
hydraulic flow passages 180 and 182. The clutch actuator solenoids
120 and 122 also have external hydraulic feedback passages 184. A
solenoid 188 selectively drives the valve member 176 operatively
from its de-energized position biased to the left as illustrated in
FIG. 2 to its energized position which allows the flow of
pressurized hydraulic fluid to flow through internal passage 182
out the clutch pressure line 170, 172 to the clutch 32, 34.
[0039] The clutch actuator solenoids 120 and 122 are current
controlled, variable regulating valves, such that a given control
current applied to solenoids 188 will result in a particular
pressure output in the clutch pressure lines 170, 172. Regulation
of the clutch actuator solenoids 120, 122 is further provided by
the pressure feedback through passages 184. Similar to the on/off
solenoids 124 and 126 and the enable valves 128 and 130, the clutch
actuator solenoids 120 and 122 have internal passages 180 to send
residual pressure from the clutch pressure lines 170 and 172 back
to the sump 90 when the solenoid is de-energized.
[0040] The method of controlling the dual clutch transmission from
launch and operation in each gear will now be discussed in greater
detail with reference to FIGS. 3-6. The method of the present
invention is initialized at some point in the operation of the
vehicle when it is determined that the engine should be under the
acceleration control of the launch strategy for the currently
engaged gear. It should be appreciated that the method of the
present invention is designed for the control of all clutch
activity, with the exception of gear shifting events. However, the
present invention, as disclosed herein, may be equally employed in
only portions of the engine and clutch operating ranges for each
gear of the transmission 10, if so desired. In that regard, as will
be discussed in greater detail below, the method of the present
invention employs an engine acceleration curve that has a number of
sections separated by predetermined threshold points in which the
different sections determine the manner of the control torque
transfer across the engaged clutch. Thus, the use of the method of
the present invention, its initialization, and its duration of
operation within the operating range of the engine and motor
vehicle is dependant on a higher order control program or control
unit that dictates when the present invention is to be applied.
[0041] The method of the present invention, as generally indicated
at 200 in FIG. 3, controls the torque transferred across the
engaged clutch 32, 34 of a vehicle having a dual clutch
transmission 10 to provide a predetermined engine acceleration
curve shown in FIG. 6 for each gear based on the engine throttle
position and subsequent changes of clutch speed that are in
response to operatively varying the torque transferred across the
engaged clutch as the engine accelerates to a predetermined speed
based on the throttle position. The method begins at the start
entry block 202 and includes the steps of determining the engine
throttle position as generally indicated at 208, determining the
currently engaged gear of the transmission 10 as generally
indicated at 210, and sensing the speed of the driven member of the
clutch as generally indicated at 212. A target engine speed based
on the engine throttle position, the currently engaged gear, and
the clutch speed is then determined, as generally indicated at 214.
The method then continuously varies the torque transferred across
the engaged clutch 32, 34 to cause the engine to track the target
engine speed, as generally indicated at 218 thus creating the
acceleration curve shown in FIG. 6. In this manner, the curve is
created based on the changes of clutch speed that are in response
to operatively varying the torque transferred across the engaged
clutch as the engine accelerates to a predetermined speed based on
the throttle position. When the method of the present invention is
disabled or otherwise turned off by the ECU or a higher level
command function, the method exits at step 204.
[0042] Once the ECU, or other control device, makes a determination
that (by some other set of control parameters beyond the scope of
this invention) engine acceleration control is required, the method
of the present invention is initiated. More specifically, and by
way of non-limiting example, the method steps generally indicated
in FIG. 3 may include those generally indicated at 220 and
described with reference to FIG. 4. Thus in this example of the
present invention, once initialized at the start block 222, the
method steps move to process block 224, which determines the engine
throttle position, then to process block 226 to determine the
currently engaged gear, and to process block 228 to sense the speed
of the driven member of the engaged clutch. These values are used
at process block 230 to determine a target engine speed based on
the engine throttle position (224), the currently engaged gear
(226), and the clutch speed (228). It should be appreciated that
the reference to clutch speed throughout this specification is
taken to simply mean the rotational speed of the driven member of
the particular clutch assembly being discussed.
[0043] Once a target engine speed has been determined at process
block 230, process block 232 determines the difference between the
target engine speed and the actual (measured) engine speed by
making a summation of the two values to produce a speed error. The
speed error from process block 232 is fed to process block 234,
which determines a dynamic torque signal. Process block 234
represents a conversion circuit that changes the raw speed error
value into an error signal that is proportioned relative to a
predetermined range. The proportional error signal is the dynamic
torque signal that represents the amount of torque that must be
transferred across the engaged clutch to cause the engine to track
the target engine speed in response to the throttle position. In
other words, the actual engine speed is determined, the target
engine speed is compared to the actual engine speed to produce a
speed error signal, and the value of the speed error signal is
proportioned to a relative value within a predetermined range to
provide the dynamic torque signal. The dynamic torque signal is
then used to form the basis for a target torque signal that is then
used to vary the torque transfer across the engaged clutch to cause
the engine to track the target engine speed.
[0044] It should be appreciated that the target torque signal is a
composite signal that comprises the dynamic torque signal discussed
above and a steady state torque signal. The term "dynamic" is used
in this context to signify that the torque transfer across the
engaged clutch is being varied according to the difference between
the target engine speed and the actual engine speed. Thus, the
speed error is reduced as the actual engine speed closes on the
target engine speed. The "steady state" torque signal represents an
additional amount of torque that must be transferred across the
engaged clutch to compensate for changes in driving conditions that
would otherwise disrupt the speed control of the engine using the
dynamic signal alone. As will be discussed in greater detail below,
the steady state torque signal is derived as the difference between
predicted changes to the engine speed and the actual change based
on known engine and transmission response times to changes in the
target torque signal. Thus, if the engine speed is not changing in
accordance with known response times, then external forces have
occurred than must be accounted for. For example, if the vehicle
were to encounter a hill while the dynamic torque signal was acting
alone to control the engine speed, some of the change in torque
transfer commanded by the dynamic torque signal would be lost to
the increased load and the engine speed would not change as
expected. Thus, the steady state torque signal is determined and
added to the dynamic torque signal so that the dynamic torque
signal will control the engine speed in a "steady state" manner
without any detrimental influences from changing driving
conditions.
[0045] Therefore, process block 236 sums the dynamic torque signal
and the steady-state torque signal to provide a target torque
signal that represents the total amount of torque that must be
transferred across the engaged clutch to cause the engine to track
the target engine speed in response to the throttle position and
changes in the driving conditions. Then, the target torque signal
is used to vary the torque transferred across the engaged clutch to
cause the engine to track the target torque signal. This is shown
in FIG. 4 by the flow path from process block 236 to the
transmission and engine assembly representatively illustrated in
block 260. It should be appreciated that as the actual engine speed
reaches, or is about to reach, the target engine speed for the
particular clutch speed from process block 228, that a new target
engine speed is determined within process block 230 to cause
another speed error at process block 232 so that the engine is
caused to accelerate and track the new target engine speed. This is
a repetitive process so that an engine speed loop is determined by
each new target engine speed in response to the changes in clutch
speed as the engine reaches the current target engine speed.
[0046] To obtain a steady state torque signal that is summed with
the dynamic torque signal at process block 236, a measured value
representing the actual dynamic torque produced must be derived
from the engine and clutch (FIG. 4, block 260). The measured
dynamic torque is generally derived by calculations that use engine
and transmission output measurements. For example, the acceleration
of the engine is determined at block 238, the acceleration of the
engaged clutch is determined at block 240, and then the engine
acceleration and the clutch acceleration are summed to determine
the clutch slip acceleration at block 242. Finally, the measured
dynamic torque transferred across the engaged clutch is determined
from the clutch slip acceleration. It should be appreciated that
there are a number of ways to derive this value and that other
terms, such as the known slip inertia for example, may have to be
involved in the calculation. However, the manner in which the
measured dynamic torque value is reached is not critical to this
control method and is beyond the scope of this invention. The
measured dynamic torque value is passed to process block 244 where
it is summed with a predicted dynamic torque value.
[0047] The predicted dynamic torque value is derived at process
block 246 by first predicting the value of the total torque that
should be transferred across the engaged clutch given the target
torque signal and based on known transmission responses to the
input of that particular target torque signal. It should be
appreciated that this predicted value may be retrieved from a
stored look-up table or calculated directly. Then, at process block
248, a feedback signal of the steady state torque value that has
been determined from the previous iteration through the engine
speed control loop is subtracted from this predicted total torque
value (from process block 246) to provide a predicted dynamic
torque value. In other words, taking the target torque signal,
which is composed of the dynamic torque signal and the steady-state
torque signal, then predicting a total torque value that should be
transferred across the engaged clutch for that target torque signal
and summing it to (i.e., subtracting) the previously derived
compensation for driving conditions (feedback of prior steady-state
torque value) will leave as a remainder a value that represents the
predicted dynamic torque signal portion of the target torque
signal.
[0048] The measured dynamic torque is summed with the predicted
dynamic torque value at process block 244 to determine if a dynamic
torque error exists. A dynamic torque error represents the
difference between the measured engine/clutch output and the
predicted output that is expected for the given target torque input
signal. This error indicates that changes in driving conditions
have occurred that must be compensated for by making a correction
to the target torque signal. More specifically, if a predicted, or
expected output value of torque transfer should occur across the
clutch for a particular target torque signal and the actual, or
measured torque that occurs is different, then external factors
(i.e., road and/or driving conditions) have occurred that are
negatively influencing the torque transfer of the clutch.
[0049] Once the dynamic torque error has been determined at process
block 244, it is integrated at process block 250 to provide the
steady-state torque signal. As previously mentioned, the
steady-state torque signal is summed with the dynamic torque signal
at process block 236 to provide the target torque signal that
represents the total amount of torque that must be transferred
across the engaged clutch to cause the engine to track the target
engine speed in response to both the throttle position and changes
in driving conditions. It should be appreciated that since the
steady-state torque value is a compensation value that is added to
the dynamic torque signal when the measured torque output value
differs from the predicted torque value, any other event that would
also cause a difference in this value will trigger an errant
steady-state compensation. To prevent this occurrence and to assist
in maintaining the correct level of steady-state torque signal, an
additional compensating value is added to the steady-state torque
signal at process block 252.
[0050] In practical terms, the only event that would cause an
errant correction by the steady-state torque compensation is a
change in the throttle position. More specifically, if a change in
throttle position is commanded, the target engine speed and thus
the target torque signal will change instantaneously in response to
the dynamic torque signal portion of the flowchart. However, since
the steady-state torque signal is derived from the measured engine
and clutch outputs, and there is a delay in the time it takes the
engine to respond to the engine controller inputs based on throttle
position changes, an errant steady-state torque compensation is
produced. More precisely, due to the delay in measured engine
response compared to the instantaneous changes in the predicted
total torque and predicted dynamic torque values the steady-state
torque signal is unable to change with equal speed thus causing an
erroneous change in the steady-state torque value in effort to
compensate. Therefore, at process block 252 a steady-state delay
signal is added to the integrated dynamic torque error (from
process block 250) to assist the steady-state torque value to
quickly compensate for the change in commanded torque.
[0051] Starting at process block 254, the method of the present
invention determines the steady-state delay by first predicting the
value of the steady state torque signal based on inputs from an
engine controller when a change in throttle position results in a
new engine target speed. This predicted value corresponds to the
correct steady-state value that should be produced in response to
the new throttle position. Then, at process block 256, based on
known engine response times to changes in the target torque signal,
the delay in the time required for the engine to reach the new
target engine speed and thus the delay in the steady-state torque
reaching its proper level is predicted. It should be appreciated,
as with the aforementioned predicted values, that this delay value
may be retrieved from a stored look-up table or calculated
directly. Finally, at process block 252, the predicted steady-state
torque value (process block 256)is summed with the integrated
dynamic torque error signal to provide the proper steady-state
torque signal. In this manner, if an incorrect dynamic torque error
is produced at process block 244, in response to the throttle
position change, the integrated dynamic torque error is enhanced,
by the summation of the steady-state delay signal at process block
252, so that the errant compensation is accounted for. This
continues until expiration of the predicted delay in engine
response (process block 258), so that the steady-state torque
compensation (from process block 252) is reinstated once the
effects from the delay in response to the change in throttle
position pass. The method of the present invention as represented
in FIGS. 3 and 4 provides an engine acceleration curve by raising
the target engine speed as the clutch speed increases in response
to the engine speed increasing as well as providing a compensation
to the changing engine target speed for efficient and smooth
operation as the driving conditions vary.
[0052] Those having ordinary skill in the art will appreciate that
when using a target engine speed to develop an acceleration curve
that is used to control the clutch engagement to regulate the speed
of an accelerating engine, the term "target engine speed" and its
associated concepts may involve a number of meanings in common
practice and the terminology used here should be clearly
understood. In general use, the phrase "target engine speed" is
described herein in connection with an "engine speed control"
scheme or strategy. As used herein, the term "engine speed control"
may mean holding the engine to a specific speed (RPM), or limiting
the engine to a specific speed, or controlling the engine speed
(and thus, its acceleration) over its operating range. Thus, engine
speed control using a target engine speed may use its target as
either a static point, or as a dynamic control.
[0053] More specifically, as an engine is called upon to accelerate
with an applied load (vehicle mass and inertia) that is connected
though a drivetrain and a selected gear to the clutch of a
transmission (in this case, a dual clutch transmission), a static
target engine speed may be determined. The static target engine
speed is a particular RPM value chosen that the engine speed will
be allowed to accelerate to by the control of the torque transfer
across the engaged clutch. As previously mentioned, the use of a
static target engine speed alone in a method of controlling engine
speed is not desirable for the dual clutch transmission. Thus, as
discussed above, the method of the present invention provides an
acceleration curve that is in actuality a constantly changing, or
dynamic, target engine speed for the given throttle position and
the selected gear based on the responsive changes in clutch speed.
It may also be regarded as a curve formed from a series of
continuously redetermined target engine speeds that the engine
acceleration is allowed to follow, or track.
[0054] From the description that follows, it will be appreciated
that the acceleration curve may be separated into portions having
predetermined thresholds. The initial portion of the acceleration
curve is first set using a static target engine speed, but as the
engine speed changes the curve moves into a continuously
redetermined target engine speed portion as the clutch speed rises.
In this regard, another non-limiting embodiment of the present
invention is generally indicated at 280 in FIG. 5. This embodiment
includes a method of controlling the torque transferred across the
engaged clutch of a vehicle having a dual clutch transmission to
provide a predetermined engine acceleration curve for each gear.
The method is based on the engine throttle position and subsequent
changes of clutch speed in response to operatively increasing the
torque transferred across the engaged clutch as the engine
accelerates to a predetermined speed based on the throttle position
is provided.
[0055] This method includes the steps of determining the engine
throttle position, determining the currently engaged gear of the
transmission, and sensing the speed of the driven member of the
engaged clutch. Once these values are determined, the method
selects an engine stall speed (from a look-up table) for the
current gear and engine throttle position. The stall speed is the
point at which an increasing transfer of torque across the engaged
clutch will stop further engine acceleration and will hold the
engine to a constant speed. The stall speed will be discussed in
greater detail below. Using the stall speed, the method determines
a target engine speed based on the engine stall speed and the
clutch speed, then increases the torque transferred across the
engaged clutch so as to cause the actual engine speed to increase
and approach the target engine speed. As the engine accelerates,
the method continuously redetermines the target engine speed to
create the engine acceleration curve after the clutch speed reaches
a predetermined percentage of the engine stall speed. The clutch
speed increases in response to the increasing engine speed, thereby
causing the engine speed to track the rising target engine speed of
the engine acceleration curve. Finally, the method increases the
torque transferred across the engaged clutch so as to operatively
lock the clutch and engine together as the clutch speed reaches a
predetermined multiple of the engine stall speed.
[0056] The method illustrated with reference to FIG. 5 allows the
engine to accelerate while causing the increasing engine speed to
follow an acceleration curve that is formed of separate control
portions. The first portion of the acceleration curve involves a
determination of an initial static target engine speed followed by
a portion having continuously redetermined (dynamic) target engine
speeds, followed finally by a portion of the curve in which the
clutch speed has met the engine speed and the clutch is locked to
the engine.
[0057] Referring to FIG. 5, once the ECU, or other control device,
makes a determination that engine speed control is required, this
embodiment of the method of the present invention is initiated at
the start entry block 282. The flow path of the method steps then
moves to process block 284, which determines the position of the
engine throttle. Then, process block 286 determines the currently
engaged gear of the transmission, which provides a reference as to
the current operating range of the vehicle and to ascertain which
of the two clutches is currently engaged and transferring torque.
Process block 288 then determines the clutch speed and process
block 290 determines the stall speed for this engine and clutch
configuration given the throttle position (284).
[0058] The step of determining engine stall speeds is important to
developing the acceleration curve in connection with the embodiment
of the present invention illustrated in FIG. 5. In connection with
an accelerating engine that is speed controlled by varying the
torque transfer across a clutch, the term "engine stall speed"
refers to the point at which the increasing clutch pressure causes
the engine to no longer accelerate so that it holds to a constant
RPM. The stall speed of the engine is dependant on the commanded
throttle position, the currently engaged gear, the torque producing
capability of the particular engine, and the torque transfer
capability of the clutch assembly. It should be appreciated that
the stall speed of an engine and clutch assembly may be derived in
different manners, such as empirical testing or by mathematical
modeling.
[0059] The engine stall speed is influenced by the commanded
throttle position in the following manner. When the engine is
commanded to accelerate from a low or idle speed by a throttle
position change, the clutch engagement force, or pressure, and
thereby the torque transferred across the clutch, may be increased
to the point at which the clutch holds the engine to a constant
speed. This is the engine stall speed for that particular throttle
setting. The higher the commanded throttle position, the higher the
stall speed. This means that the stall speed varies over a range in
conjunction with the throttle position. Stall speeds essentially
represent how the clutch effects the engine speed, so the stall
speeds may be expressed in terms of the amount of clutch engagement
force, or the amount of torque transferred across the clutch, or
the clutch speed versus engine speed (i.e. clutch slip), at which
the engine stops accelerating.
[0060] By extension, the range of the stall speeds is influenced by
the selected gear and the particular engine used in the vehicle in
the following manner. First, the selected gear has a secondary
effect by the load that the particular gear places upon the driven
member of the clutch. Secondly, the torque producing capability of
the engine impacts the stall speed. For example, if two engines are
joined with clutches having comparable torque transfer
capabilities, the stall speeds of the engine that produces a given
amount of torque, with the majority of the torque produced in the
lower portion of the RPM range, will be lower than that of the
engine producing roughly the same amount of torque but producing it
higher in the RPM range. Additionally, it should be appreciated
that the torque transferring capability of the particular clutch
used will also have an influence on the stall speeds. Since stall
speeds are consistent for any given throttle position across the
range of throttle positions for each gear and for each particular
engine and transmission configuration, once they have been
determined, they are generally stored in a look-up or reference
table.
[0061] In the initial portion of the engine acceleration curve,
meaning as the engine is brought from the low end of its
acceleration range, either in first gear from a vehicle standing
start or from the acceleration following a gear change, the stall
speed determined from process block 290 is used to set a static
target engine speed. Since the stall speed is the point at which a
highest clutch pressure for a given throttle position can be
applied before the engine is caused to decelerate, it provides the
maximum transfer of torque for the engine speed and thereby maximum
vehicle acceleration for the engine at the particular throttle
setting. This occurs because as the engine speed, and thereby the
torque output, are held constant by the clutch engagement pressure
as the engine attempts to accelerate to the throttle position, the
clutch speed is still increasing and accelerating the vehicle.
Thus, in this example of the method of the present invention,
process block 292 sets the initial target engine speed to the
engine stall speed. However, it should be appreciated that this
approach of keeping the target engine speed at the stall speed has
a practical limitation in that, for the given engine speed only a
portion of the available torque output of the engine is reached.
Also, the clutch speed will reach a limit relatively far below the
engine speed, since the clutch pressure cannot be increased any
further beyond holding the engine at stall speed or the engine
speed and torque output will drop off. Thus, if only a single
static target engine speed were used, as in some conventional
methods known in the related art, the above mentioned drawbacks of
drivability harshness will arise as the operator must continually
increase the throttle position to command a higher engine speed,
which would cause the engine speed to be controlled in distinct and
noticeable stages. These control methods known in the related art
also cause the operator to put the throttle to its highest setting
each time an increase in engine speed is desired to achieve
continued engine acceleration, resulting in loss of efficiency.
However, the method of the present invention accounts for the
static target engine speed limitation by entering into a dynamic
portion of the acceleration curve at a predetermined threshold.
[0062] When employing a dynamic target engine speed, as in the
method of the present invention, even if using a static target
engine speed to initially accelerate the engine first toward the
stall speed, there is a transitional point at which the clutch
pressure, and thus the torque transfer, should be slightly reduced
to allow the engine speed to increase. This transition point can be
defined at a point where the clutch speed reaches some portion of
the stall speed of the engine. The transition point is the initial
control point of the engine speed control curve of the present
invention and is made to occur at such a time as the increasing
engine speed and clutch engagement will be smooth and efficient.
This transition point will be discussed in greater detail below.
However, when this first transition point is reached and the clutch
pressure is slightly reduced allowing the engine speed to increase,
the clutch speed will also increase until a second transitional
point is reached. This second transition point is where the clutch
speed meets the engine speed and the clutch and engine are
essentially locked. The determination of the target engine speed is
particularly important in this range, between the transition
points, where the clutch allows the engine speed to rise.
[0063] As shown in FIG. 5, once process block 292 determines the
stall speed and sets the target engine speed to that value,
decision block 294 determines if the clutch speed has reached a
transition point defined by a predetermined percentage of the
engine stall speed. If the clutch speed has not yet reached this
transition point, the static target engine speed is passed, as the
target engine speed, to the remainder of the process beginning at
process block 300. As will be discussed below, the remainder of the
process will increase the torque transferred across the engaged
clutch so as to cause the actual engine speed to increase and
approach the target engine speed. However, once the clutch speed
has reached, or exceeds, the transition point of the predetermined
percentage of the engine speed, then the flow path continues to
process block 296. The target engine speed is continuously
re-determined in process block 296 to create the dynamic portion of
the engine acceleration curve shown in FIG. 6. The clutch speed
increases in response to the increasing engine speed thereby
causing the engine speed to track the rising target engine speed of
the dynamic portion of the engine acceleration curve.
[0064] Thus, as process block 296 is continuously redetermining the
dynamic portion of the acceleration curve, decision block 297
determines if the clutch speed has reached another transition point
defined by a predetermined multiple of the engine stall speed. If
the clutch speed has not yet reached this transition point, the
continuously redetermined dynamic target engine speed is passed to
process block 300. However, once the clutch speed has reached, or
exceeds, the transition point of the predetermined multiple of the
engine speed, then the flow path continues to process block 298
which causes the ECU to increase the torque transferred across the
engaged clutch so as to operatively lock the clutch and engine
together. At this point, generally high in the engine RPM range,
the clutch speed and the engine speed are equal and thus the
acceleration curve will rise to the maximum for the given throttle
position or other conditions (beyond the scope of this invention)
are such that a gear shift will be performed. Process block 299
allows a return through the method steps or a departure from the
acceleration curve if other operations (e.g. shifting) are to
occur. It should be appreciated, however, that the throttle
position may be changed to reflect a lower engine speed setting
without a gear shift and the engine is similarly controlled in a
deceleration curve.
[0065] In this manner, the method of the present invention also
provides a predetermined engine deceleration curve for each gear
based on the engine throttle position and subsequent changes of
clutch speed that are in response to operatively varying the torque
transferred across the engaged clutch, as the engine decelerates to
a predetermined speed based on the throttle position. To accomplish
this, the method further includes the steps of determining the
currently engaged gear of the transmission and sensing the clutch
speed when a reduction in the engine throttle position is detected.
Then, the pressure applied to the engaged clutch is initially
decreased in a linear manner as the engine decelerates thereby
controlling the rate of vehicle deceleration. Further, an initial
decelerating target engine speed is determined based on the
currently engaged gear and the engine throttle position. Then, the
torque transferred across the engaged clutch is varied to cause the
engine speed to decelerate toward the initial decelerating target
engine speed. After the decelerating clutch speed reaches a
predetermined percentage of the initial decelerating target engine
speed, due to the clutch response to the decreasing engine speed,
the target engine speed is incrementally redetermined to create the
engine deceleration curve thereby causing the engine speed to track
the decreasing target engine speed of the engine deceleration
curve. Finally, the torque transferred across the engaged clutch is
continuously varied to cause the engine to decelerate along the
engine deceleration curve.
[0066] Referring again to FIG. 5, process block 300 and the
remainder of the blocks of FIG. 5 are functionally similar to those
of FIG. 4. Thus, process block 300 determines the difference
between the target engine speed and the actual engine speed, which
is essentially determining the difference between the target engine
acceleration curve and the actual (measured) engine speed by
summing the two values to produce a speed error. The speed error
from process block 300 is fed to process block 302, which
determines a dynamic torque signal. Process block 302 represents a
conversion circuit that changes the raw speed error value into an
error signal that is proportioned relative to a predetermined
range. The proportional error signal is the dynamic torque signal
that represents the amount of torque that must be transferred
across the engaged clutch to cause the engine to track the target
engine speed curve in response to the throttle position. In other
words, the actual engine speed is determined, the target engine
speed curve is compared to the actual engine speed to produce a
speed error signal, and the value of the speed error signal is
proportioned to a relative value within a predetermined range to
provide the dynamic torque signal. The dynamic torque signal is
then used to form the basis for a target torque signal that is then
used to vary the torque transfer across the engaged clutch to cause
the engine to track the target engine speed.
[0067] The target torque signal is a composite signal that
comprises the dynamic torque signal discussed above and a steady
state torque signal. The term "dynamic" is used in this context to
signify that the torque transfer across the engaged clutch is
varied according to the difference between the target engine speed
and the actual engine speed. Thus, the speed error is reduced as
the actual engine speed closes on the target engine speed. The
"steady state" torque signal represents an additional amount of
torque that must be transferred across the engaged clutch to
compensate for changes in driving conditions that would otherwise
disrupt the speed control of the engine using the dynamic signal
alone. The steady state torque signal is derived as the difference
between predicted changes to the engine speed and the actual change
based on known engine and transmission response times to changes in
the target torque signal. Thus, the steady state torque signal is
determined and added to the dynamic torque signal so that the
dynamic torque signal will control the engine speed in a "steady
state" manner without any detrimental influences from changing
driving conditions.
[0068] Therefore, process block 304 sums the dynamic torque signal
and the steady-state torque signal to provide a target torque
signal that represents the total amount of torque that must be
transferred across the engaged clutch to cause the engine to track
the target engine speed in response to the throttle position and
changes in the driving conditions. Then, the target torque signal
is used to vary the torque transferred across the engaged clutch to
cause the engine to track the target torque signal. This is shown
in FIG. 5 by the flow path from process block 304 to the
transmission and engine assembly represented by block 306. It
should be appreciated that as the actual engine speed reaches, or
is about to reach, a point on the target engine speed curve for the
particular clutch speed from process block 294, another speed error
is determined at process block 300 so that the engine is caused to
accelerate and track the target engine speed curve, or that the
engine is merely allowed to accelerate but be held to the target
engine speed curve in relation to the clutch speed by process block
298. Regardless, a target engine speed curve is established that
will provide control of the torque transferred across the engaged
clutch for the given gear and throttle position.
[0069] To obtain a steady state torque signal that is summed with
the dynamic torque signal at process block 304, a measured value
representing the actual dynamic torque produced must be derived
from the engine and clutch (block 306, FIG. 5). The measured
dynamic torque is generally derived by calculations that use engine
and transmission output measurements. For example, the acceleration
of the engine is determined at block 308, the acceleration of the
engaged clutch is determined at block 310, and then the engine
acceleration and the clutch acceleration are summed to determine
the clutch slip acceleration at block 312. Finally, the measured
dynamic torque transferred across the engaged clutch is determined
from the clutch slip acceleration. It should be appreciated that
there are a number of ways to derive this value and that other
terms, such as the known slip inertia for example, may have to be
involved in the calculation. However, the manner in which the
actual dynamic torque value is reached is not critical to this
control method and is beyond the scope of this invention. The
measured dynamic torque value is passed to process block 314 where
it is summed with a predicted dynamic torque value.
[0070] The predicted dynamic torque value is derived at process
block 316 by first predicting the value of the total torque that
should be transferred across the engaged clutch given the target
torque signal and based on known transmission responses to the
input of that particular target torque signal. It should be
appreciated that this predicted value may be retrieved from a
stored look-up table or calculated directly. Then, at process block
318, a feedback signal of the steady state torque value that has
been determined from the previous iteration through the engine
speed control loop is subtracted from this predicted total torque
value (from process block 316) to provide a predicted dynamic
torque value. In other words, taking the target torque signal,
which is composed of the dynamic torque signal and the steady-state
torque signal, then predicting a total torque value that should be
transferred across the engaged clutch for that target torque signal
and summing it to (i.e., subtracting) the previously derived
compensation for driving conditions (feedback of prior steady-state
torque value) will leave as a remainder a value that represents the
predicted dynamic torque signal portion of the target torque
signal.
[0071] The measured dynamic torque is summed with the predicted
dynamic torque value at process block 314 to determine if a dynamic
torque error exists. A dynamic torque error represents the
difference between the measured engine/clutch output and the
predicted output that is expected for the given target torque input
signal. This error indicates that changes in driving conditions
have occurred that must be compensated for by making a correction
to the target torque signal. More specifically, if a predicted, or
expected output value of torque transfer should occur across the
clutch for a particular target torque signal and the actual, or
measured torque is different, then external factors (i.e., road
and/or driving conditions) have occurred are negatively influencing
the torque transfer of the clutch.
[0072] Once the dynamic torque error has been determined at process
block 314, it is integrated at process block 320 to provide the
steady-state torque signal. As previously mentioned, the
steady-state torque signal is summed with the dynamic torque signal
at process block 304 to provide the target torque signal that
represents the total amount of torque that must be transferred
across the engaged clutch to cause the engine to track the target
engine speed in response to both the throttle position and changes
in driving conditions. It should be appreciated that since the
steady-state torque value is a compensation value that is added to
the dynamic torque signal when the measured torque output value
differs from the predicted torque value, any other event that would
also cause a difference in this value will trigger an errant
steady-state compensation. To prevent this occurrence and to assist
in maintaining the correct level of steady-state torque signal, an
additional compensating value is added to the steady-state torque
signal at process block 322.
[0073] In practical terms, the only event that would cause an
errant correction by the steady-state torque compensation is a
change in the throttle position. More specifically, if a change in
throttle position is commanded, the target engine speed and thus
the target torque signal will change instantaneously in response to
the dynamic torque signal portion of the flowchart. However, since
the steady-state torque signal is derived from the measured engine
and clutch outputs, and there is a delay in the time it takes the
engine to respond to the engine controller inputs based on throttle
position changes, an errant steady-state torque compensation is
produced. More precisely, due to the delay in measured engine
response compared to the instantaneous changes in the predicted
total torque and predicted dynamic torque values the steady-state
torque signal is unable to change with equal speed thus causing an
erroneous change in the steady-state torque value in effort to
compensate. Therefore, at process block 322 a steady-state delay
signal is added to the integrated dynamic torque error (from
process block 320) to assist the steady-state torque value to
quickly compensate for the change in commanded torque.
[0074] Starting at process block 324, the method of the present
invention determines the steady-state delay by first predicting the
value of the steady state torque signal based on inputs from an
engine controller when a change in throttle position results in a
new engine target speed. This predicted value corresponds to the
correct steady-state value that should be produced in response to
the new throttle position. Then, at process block 326, based on
known engine response times to changes in the target torque signal,
the delay in the time required for the engine to reach the new
target engine speed and thus the delay in the steady-state torque
reaching its proper level is predicted. It should be appreciated,
as with the aforementioned predicted values, that this delay value
may be retrieved from a stored look-up table or calculated
directly. Finally, at process block 322, the predicted steady-state
torque value (process block 324)is summed with the integrated
dynamic torque error signal to provide the proper steady-state
torque signal. In this manner, if an incorrect dynamic torque error
is produced at process block 314, in response to the throttle
position change, the integrated dynamic torque error is enhanced by
the summation of the steady-state delay signal at process block
322, so that the errant compensation is accounted for. This
continues until expiration of the predicted delay in engine
response (process block 326), so that the steady-state torque
compensation (from process block 320) is reinstated once the
effects from delay in response to the change in the throttle
position pass.
[0075] To further refine the ability of the present invention to
account for the desired driveability and comfort issue relating to
the engagement of the clutches of the dual clutch transmission 10
and the rate of acceleration of the engine, and thus the vehicle,
the method of the present invention also provides for the use of an
equation in calculating the acceleration curve. The equation
includes an operatively selectable control constant that directly
influences the rate of increase of engine and clutch speed thus
providing operative control over the ride characteristics of the
vehicle.
[0076] Thus, the present invention also encompasses a method of
controlling the pressure applied to the engaged clutch of a vehicle
having a dual clutch transmission to control the torque transferred
across the engaged clutch by employing a predetermined engine
acceleration curve for each gear based on the engine throttle
position and subsequent changes of clutch speed in response to
operatively varying the pressure on the engaged clutch as the
engine accelerates to a predetermined speed based on the throttle
position. The method includes the steps of determining the engine
throttle position, determining the currently engaged gear of the
transmission, and sensing the speed of the driven member of the
engaged clutch. Then, an engine stall speed for the current gear
and engine throttle position is selected from a look-up table and a
target engine speed is determined based on the engine stall speed.
The torque transferred across the engaged clutch is increased so as
to cause the engine speed to increase and approach the target
engine speed thereby increasing the clutch speed and accelerating
the vehicle. The pressure applied to the engaged clutch is then
reduced to allow the engine to continue to accelerate toward the
target engine speed when the clutch speed reaches a predetermined
percentage of the engine stall speed thereby continuing to increase
the clutch speed while at the same time accelerating the
vehicle.
[0077] A target engine speed is continuously redetermining based on
the engine stall speed and the clutch speed using a target engine
speed equation that influences the rate of increase of engine and
clutch speed after the clutch speed reaches the predetermined
percentage of the engine stall speed. The target engine speed
equation is defined as: 1 T = S + { [ C + [ ( K - 2 ) * S ] } 2 4 *
( K - 1 ) * S ( 1 )
[0078] wherein T is the target engine speed, S is the stall speed
determined for the current gear and throttle position, C is the
clutch speed, and K is a control constant. Using this equation in
the method of the present invention, the clutch speed may be caused
to increase in response to the increasing engine speed thereby
causing the engine speed to track the rising target engine speed.
Ultimately, the applied pressure on the engaged clutch is increased
so as to operatively lock the clutch and engine together as the
clutch speed reaches a predetermined multiple of the engine stall
speed.
[0079] Essentially, the application of the equation provides a more
detailed manner of determining the target engine speed in process
block 294 through 297 of FIG. 5 as previously discussed to develop
the acceleration curve. The curve and its associated components are
graphically illustrated in FIG. 6. In FIG. 6, the relative throttle
position over time is represented graphically and is generally
indicated at 400. An increase in the throttle position is shown at
402. This change in throttle position may cause the ECU to initiate
the method of the present invention by commanding engine
acceleration. A decrease in throttle position is shown at 404. As
generally indicated at 420 in FIG. 6, the related activity of the
engine and the clutch are shown in reference to the determined
target engine speed. Specifically, the clutch speed is shown as a
solid line 406. Control of the torque transfer across the clutch
attempts to cause the engine to track the target engine speed shown
as a dotted line 408, thereby producing the resultant engine
acceleration curve, generally indicated as the solid line at 410.
To graph this interaction, speed as a relative value is designated
on the vertical axis and time is indicated along the horizontal
axis. The time reference for the lines at 420 corresponds with the
time shown for the throttle position at 400 and likewise for the
pressure curve, generally indicated at 440. The pressure curve 440
depicts the pressure applied to the engaged clutch to cause the
engine to track the acceleration curve 410.
[0080] More specifically, and referring back to process block 298
in FIG. 5, the target engine speed is determined based on the
currently engaged gear and the engine throttle position using an
equation that includes a predetermined control constant that
influences the rate of increase in engine and clutch speed, and
thereby vehicle acceleration. This is also effectively the rate at
which the clutch engages the engine, and corresponds to the
relative harshness of the clutch engagement. At decision block 294
it is determined whether the clutch speed has reached or exceeded
the first transition point. If decision block 294 determines that
the clutch speed has not yet reached the first transition point,
the "NO" path is followed and the engine stall speed, as the static
target engine speed, is passed to process block 300, as shown in
FIG. 5 and as graphically represented as the portion of the engine
acceleration curve designated 412 in FIG. 6. In FIG. 6, the
relative value of the engine stall speed for the given throttle
position is shown as dotted horizontal line 422. The first
transition point occurs along dotted vertical line 424 where the
increasing clutch speed meets the predetermined percentage of the
engine stall speed, shown as a horizontal dotted line 430.
[0081] If decision block 294 determines that the clutch speed has
reached the first transition point, the "YES" path is followed and
the above-mentioned equation is employed at process block 296 (FIG.
5) to continuously redetermine the dynamic target engine speed. The
first solution to the equation provides a target engine speed for
the particular instantaneous clutch speed, which is passed to
process block 300 to sum with the measured (actual) engine speed to
produce a difference, or error, between the target engine speed and
the measured engine speed. This event is followed by the repetitive
solutions to the equation that will provide the dynamic target
engine speed (the portion of the engine acceleration curve
designated 414 in FIG. 6) by repeatedly solving for the equation as
the clutch speed increases in response to engine acceleration.
Thus, not only is a single curve resolved for the particular
throttle position in the particular selected gear, but a whole
range of target engine curves are subsequently created for each
possible throttle position in each available gear. It should be
appreciated that the repetitive solution to the instantaneous
equation may be controlled by a sampling rate or other incremental
indexing rather than performing a non-stop refiguring of the
target. It should be further appreciated that once the appropriate
target engine speed curves are determined they may be stored and
referred to rather than requiring the equation step (in process
block 296) to continue as an on-going mathematical process.
[0082] With respect to the transition points that are determined by
decision block 294 and 297, the stall speed and the control
constant K are the defining factors. Generally speaking, from the
equation (1) and FIG. 6, it can be seen that once the engine speed
reaches and exceeds the first transition point and the equation is
used (in block 296) to determine the dynamic target engine speed,
the complex term added to the stall speed "S" causes the dynamic
target engine speed to be initially set slightly higher than the
stall speed. The clutch pressure is then slightly reduced to allow
the engine to accelerate in further response to the commanded
throttle position. This is graphically illustrated in FIG. 6 as the
engine speed line 410 passes through the first transition point at
dotted line 424 and begins its rise in the dynamic acceleration
portion 414. Then, throughout the continuation of the dynamic
acceleration portion 414 of the curve, as clutch speed (C)
increases (406, FIG. 6), the target engine speed (T) increases
(408, FIG. 6) at a lesser rate until the point at which the clutch
speed and the engine speed are the same (the second transition
point). As shown in FIG. 6, the second transition point occurs
along the vertical dotted line 426 where the increasing clutch
speed meets the predetermined multiple of the engine stall speed
shown as horizontal dotted line 432.
[0083] The K value is important to setting the transition points of
the target engine speed curve. The K value is a constant that
controls the rate at which the clutch speed approaches the engine
speed. Thus, the method of the present invention also provides
selective control over the first transition point wherein the
predetermined percentage of the engine stall speed is further
defined as the relationship ([2-K]*S), where K is the chosen
control constant and S is the stall speed for the given throttle
position. This particular relationship is shown in FIG. 6 as the
horizontal dotted line 430, which represents the predetermined
percentage of the engine stall speed that sets the first transition
point and is further indicated as the relative speed represented by
the value ([2-K]*S).
[0084] Additionally, the method of the present invention also
provides selective control over the second transition point that
operatively locks the clutch and engine together by setting the
predetermined multiple of the engine stall speed as the
relationship (K*S), where K is the chosen control constant and S is
the stall speed for the given throttle position. This particular
relationship is shown in FIG. 6 as the horizontal dotted line 432,
which represents the predetermined multiple of the engine stall
speed that sets the second transition point and is further
indicated as the relative speed represented by the value (K*S).
[0085] The goal of the engine acceleration curve of the present
invention is not only to cause the engine to track the target
engine speed, but also to cause the clutch to engage smoothly and
with the most efficient transfer of engine torque. The K value and
its relationship to the stall speed 422 determines the rate of the
rise in the target engine speed 408 and the rate at which the
clutch speed 406 meets the engine speed 410 thereby providing the
relative "feel" or smoothness of the clutch engagement. The
relative value of K controls the rate of clutch engagement and thus
the rate of vehicle acceleration. Mathematically, K must be greater
than 1 (one) and less than (two) 2. As the value for K is set
closer to 1, the clutch engagement, and thus the vehicle ride, will
be harsher with very rapid clutch engagement and sharper engine
acceleration from the static target engine speed to a clutch
locking event. In other words, setting the value of K closer to 1
brings the two transition points described by the vertical dotted
lines 424 and 426 closer to the stall speed, so that the dynamic
portion 141 of the acceleration curve 410 (between lines 424 and
426, FIG. 6) is much smaller. In this case, the clutch will lock up
very quickly with hardly any clutch slip.
[0086] When the K value is selected closer to 2, the clutch
engagement will be smoother and more prolonged, and thus the
vehicle ride will be smoother with slower engine acceleration to
the clutch lock point. The median K value of 1.5 is generally
considered the best compromise of acceleration and smoothness for
most current engines and vehicles. However, this is greatly
influenced by the engine and clutch capabilities, as discussed
above, as well as gear selection and vehicle weight. In this
manner, the K value controls the relative smoothness of the
vehicle's acceleration and harshness of the clutch engagement. It
is an adjustable value that may be selected to provide a
programmable variable based on a predetermined desired
"characteristic" driving feel for any given vehicle. It should be
appreciated that since the determination of target engine speed is
based, in part, on the gear selected, then a different K value can
be used for each gear, if so desired. It should be further
appreciated that, given the flexibility provided by the K term, K
can be adjusted within the framework of the present invention so as
to function with any engine and dual clutch transmission
combination, in any vehicle.
[0087] Those having ordinary skill in the art, should appreciate
that in certain applications it may be desirable to further
increase the smooth feel of the clutch to engine engagement by not
actually completing a full clutch to engine lock up. Thus, the
method of the present invention also provides the step of
increasing the applied pressure on the engaged clutch as the clutch
speed reaches a predetermined multiple of the engine stall speed so
as to not operatively lock the clutch and engine together but allow
a small predetermined amount of slip to occur to provide for a
smooth transfer of motive force between the clutch and engine. This
is graphically illustrated in FIG. 6 as the section located between
the vertical dotted lines 426 and 428, also indicated as "lock or
minimal slip." In this section, the clutch pressure indicted at 446
may either be a value high enough to cause the clutch and engine to
lock as discussed above or be of some lesser predetermined pressure
value which will allow for a minimal slip.
[0088] It should be still further appreciated that as each gear in
the dual clutch transmission is prepared to accept torque to
provide motive force to the vehicle, the on-coming clutch will
enter a "standby" portion of the acceleration curve shown in graph
420. Thus, the method of the present invention provides for
preparatory steps of performing a preparatory clutch pressure fill
without a transfer of torque to the clutch that will be the engaged
clutch for either first gear, or the gear immediately above or
below the currently engaged gear. This is illustrated in FIG. 6 at
444 as the pressure initially increases along the clutch pressure
line 442 in curve 440. The preparatory clutch pressure fill is
performed with respect to the first gear clutch when the vehicle is
stationary and the brake is applied and is performed with respect
to any other gear immediately prior to a shift. The standby portion
indicates that no engine acceleration has yet occurred but is
representative of a period prior to the transfer of torque across
the clutch that will drive the particular gear.
[0089] It should be still further appreciated that, while the
method of the present invention is applicable in each gear of the
dual clutch transmission, an additional requirement is necessary
for first gear to launch the vehicle from a standing start. Thus,
the method of the present invention also includes the steps of
performing a predetermined nominal increase to the pressure applied
to the clutch that drives first gear after the standby portion of
the engine acceleration curve for first gear when first gear is
engaged. This is illustrated in FIG. 6 at 448 as the pressure
increases along the clutch pressure line 442 in the curve 440. The
predetermined nominal pressure increase is performed so as to cause
a slight forward creep of the vehicle in anticipation of increased
torque transfer across the clutch to drive the vehicle forward. As
shown in the "Creep" portion of the engine acceleration curve in
FIG. 6, creep is determined for first gear that is representative
of the predetermined nominal pressure increase to the clutch
indicating that only slight engine acceleration has occurred that
is representative of the period immediately prior to the transfer
of torque across the clutch that will drive the first gear.
[0090] It should be still further appreciated that in a like
manner, the method of the present invention also provides a
predetermined engine deceleration curve for each gear based on the
engine throttle position and subsequent changes of clutch speed in
response to operatively varying the pressure on the engaged clutch,
as the engine decelerates to a predetermined speed based on the
throttle position. Thus, the method further includes the steps of
determining the currently engaged gear of the transmission, sensing
the clutch speed when a change in the engine throttle position
indicating a commanded deceleration is detected, and decreasing the
pressure as applied to the engaged clutch initially in a linear
manner as the engine decelerates thereby controlling the rate of
vehicle deceleration. Then, the method selects an engine stall
speed for the current gear and engine throttle position from a
look-up table. Using the engine stall speed, the method determines
a target engine speed based on the engine stall speed and the
clutch speed using a target engine speed equation that influences
the rate of decrease of clutch engagement pressure. The target
engine speed equation is defined as, 2 T = S + { [ C + [ ( K - 2 )
* S ] } 2 4 * ( K - 1 ) * S ( 2 )
[0091] wherein T is the target engine speed, S is the stall speed
determined for the current gear and throttle position, C is the
clutch speed, and K is a control constant. The method then controls
the reduction in the pressure applied to the engaged clutch causing
the engine speed to decelerate toward the target engine speed when
the clutch speed reaches a predetermined percentage of the engine
stall speed. The target engine speed is continuously redetermined
after the clutch speed reaches the predetermined percentage of the
engine stall thereby causing the engine speed to track the
decreasing target engine speed.
[0092] The deceleration curve is graphically illustrated in FIG. 6
as a continuation of the acceleration curve 410. Specifically, if a
condition exists where the transmission does not or cannot shift
into a higher gear, the engine will eventually begin to decelerate
and the curve 410 will pass the 428 line as illustrated in FIG. 6.
It should also be appreciated that the 410 curve may also be
entered into after a downshift in which further engine deceleration
is required. Regardless, in response to the reduced throttle
position that reflects a commanded downshift, a deceleration stall
speed is determined. The deceleration stall speed (S).sub.Decel is
shown in FIG. 6 as the dotted horizontal line 434. In a similar
manner to the first and second transition points which occur during
the engine acceleration, there are two deceleration transition
points that occur at dotted vertical lines 450 and 452. These
deceleration transition points are located at the points where the
(K*S).sub.Decel (at dotted line 436) and ([2-K]*S).sub.Decel (at
dotted line 438) relationships occur and are substantially
identical to those previously discussed above but utilize the
deceleration stall speed (S).sub.Decel for the given lowered
throttle position.
[0093] Therefore, the method of the present invention provides a
determination of a target engine speed that is based on the
currently engaged gear and the engine throttle position by
employing a repetitively determined control constant which
influences the rate at which the clutch engages the engine. The
control constant K providing the relative "feel" or smoothness of
the clutch engagement. In this manner, the present invention
provides an engine acceleration curve (FIG. 6) that allows for the
predetermined selection of the relative smoothness of the vehicle's
acceleration and harshness of the clutch engagement by the varying
of the control constant as applied in the equation (1). Likewise,
the method provides for an engine deceleration curve, which
accomplish these same goals. Thus, the method of the present
invention overcomes the drawbacks and disadvantages of all dual
clutch transmission engine speed control schemes of the prior
art.
[0094] The invention has been described in an illustrative manner.
It is to be understood that the terminology which has been used is
intended to be in the nature of words of description rather than of
limitation. Many modifications and variations of the invention are
possible in light of the above teachings. Therefore, within the
scope of the appended claims, the invention may be practiced other
than as specifically described.
* * * * *