Note: Descriptions are shown in the official language in which they were submitted.
1 337827
METHOD OF SPEED REDUCTION RATIO CONTROL IN CONTINUOUSLY
VA~TARTT" SPEED TRANSMISSION
This invention relates to a method for controlling a speed
reduction ratio in a continuously variable speed transmission
for a vehicle.
The following methods are known for speed control of a
continuously variable speed transmission: selecting as a
reference value either (a) the engine speed, (b) the rate of
change of engine speed, or (c) the speed reduction ratio, and
controlling the speed reduction ratio so as to follow the
reference value.
These control methods, however, do not take the acceleration
caused by the reserved power of the engine into account. As
a result, the speed reduction ratio according to the known
control methods tends to be larger or smaller than expected,
thus at a low vehicle speed causing: (a) time-lag and
unfavorable feeling produced by the time-lag (poor response)
because of a small change rate of the speed reduction ratio
during ratio control toward greater speed reduction ratio,
(b) poor fuel consumption or unfavorable feelings
accompanying an excessive increase of engine speed during
ratio control toward smaller speed reduction ratio, (c)
hunting of the engine speed because of a small change rate of
the speed reduction ratio during ratio control toward
1 337827
greater speed reduction ratio, and (d) poor fuel consumption
associated with poor efficiency due to excessive reduction of
the speed reduction ratio during deceleration.
Therefore, the applicant proposed a method for solving these
problems which comprises; calculating a change rate of speed
reduction ratio by summing up a component corresponding to a
predicted acceleration and a component corresponding to a
reference change rate of the engine speed, and controlling
speed reduction ratio based on the calculated change rate of
speed reduction ratio (as exemplified by Japanese Patent
laid-open Publications No. 63(1988)-53343 and No. 63(1988)-
53344)-
In the above control method, however, if the change rate ofspeed reduction ratio is calculated based on level road
running, the actual engine speed will deviate from the
reference engine speed because of a difference in running
resistances during running on an ascending slope situation
the actual engine speed tends to be less than the reference
speed, while in the descending slope situation the actual
engine speed tends to be greater than the reference speed.
Many continuously variable speed transmissions employ clutch
means which can control the transmission of engine power. In
the partial engaging state of the clutch means, since the
whole engine power is not directly transmitted to the
transmission, such speed control as described above cannot be
applied. For this reason, in the prior art, the speed
control is divided into one for use "during clutch
engagement" and the other for use "after clutch engagement
completion". As an example of the clutch control, for
instance, as disclosed in Japanese Patent laid-open
1 337827
Publication No. 56 (1981)-95722, clutch opening is controlled
based on engine throttle opening, engine speed and vehicle
speed.
But, in the case where the speed reduction ratio is
controlled in the two stages, "during clutch engagement" and
"after clutch engagement completion", it is difficult to
smoothly shift the clutch engagement control stage to the
post clutch engagement completion control stage and to
smoothly start the speed reduction ratio control, because it
is not easy to identify the exact time of clutch engagement
completion, thus deteriorating driver's feeling. And, for
the speed reduction ratio control, such a control has been
employed as to set a reference engine speed corresponding to
an accelerator opening and to make the engine speed follow
the reference engine speed. But, if the determination of
clutch engagement completion is inaccurate, the start of the
speed reduction ratio control based on the reference engine
speed may be delayed, thus-resulting in momentary abrupt
increase of the engine speed or insufficient feeling of
acceleration.
It is an object of the invention to provide an improved
method for controlling a speed reduction ratio of a
continuously variable speed transmission.
According to one aspect of the invention there is provided a
method of controlling a speed reduction ratio of a
continuously variable transmission in a vehicle, said
transmission being operatively connected to an engine,
comprising the steps of determining a predicted acceleration
of said vehicle based on a reserved power of said engine,
determining a reference change rate of the speed of said
engine, calculating a change rate of speed reduction ratio of
said transmission as the sum of a first component
corresponding to said predicted acceleration of said vehicle
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and a second component corresponding to said reference change
rate of the speed of said engine, and controlling the speed
reduction ratio using said calculated change rate of speed
ratio as a control value.
It is an advantage of the preferred embodiment of the present
invention that it provides a method for controlling speed
reduction ratio in which a change rate of speed reduction
ratio is calculated with such correction as to minimize the
effect of the running resistance variations due to the
ascending or descending slopes.
It is another advantage of the preferred embodiment of the
present invention that it provides a method for controlling
speed reduction ratio in which the control in the two stages
(the stage of "during clutch engagement control" the stage of
"after clutch engagement completion") is continuously and
smoothly performed.
The advantage of the present invention are fulfilled in the
preferred embodiment by providing the following methods:
The first method includes; determining a predicted
acceleration dVo/dt calculated from reserved power of engine,
a reference change rate dNo/dt of the engine speed obtained
based on a parameter representing driver's intention to
accelerate or decelerate, an actual vehicle speed V, an
actual acceleration dV/dt and an actual engine speed N,
-
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computing a change rate di/dt of speed reduction ratio by the
following equation using the above values.
di/dt = -C~*(N/~2)*(dVo/dt)+C2*(1/V)*(dNo/dt)
+C3*(N/V2)*(dVo/dt~dV/dt)
where, Cl, C2 and C3 are constants, and controlling the speed
reduction ratio of a continuously variable speed transmission
based on the above computed change rate di/dt. (As used in
the specification an asterisk "*" in a formula represents
multiplication.)
The second method includes; determining a predicted
acceleration dVo/dt calculated from reserved power of the
engine, a reference change rate dNo/dt of engine speed
obtained based on a parameter representing driver's intention
to accelerate or decelerate, an actual vehicle speed V, an
actual engine speed N, and a transmission rate KcL of engine
power through a clutch, computing a change rate di/dt of
speed reduction ratio by the following equation using the
above values
0
d i /d t = -C.,* (N/V2 ) * (dVo/d t ) *Kc 1
+C,s*( l/Y)*(dNo/dt)
where, C4 and C5 are constants, and controlling the speed
reduction ratio of a continuously variable speed transmission
based on the above computed change rate di/dt.
Reference is now made to the accompanying drawings in which:
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FIG. 1 is a hydraulic circuit diagram of a continuously
variable speed transmission of which speed reduction ratio is
controlled by a method according to the invention;
FIG. 2 is a cross sectional view of ratio control servo units
of aforesaid continuously variable speed transmission;
FIG. 3 is a cross sectional view of a clutch-control servo
unit of aforesaid continuously variable speed transmission;
FIG. 4 is a flow chart showing clutch control;
FIG. 5A and 5B are graphs for obtaining reference clutch
opening and clutch opening control speed respectively;
- 5a -
. -
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FIG. 6 is a flow chart showing speed reduction ratio controlassociated with the first embodiment of the instant
invention;
FIG. 7 and 8 are graphs showing maps for obtaining engine - -
power and reference change rate of the engine speed
respectively;
FIG. 9 is a flow chart showing a subroutine for obtaining the
corrective component D(di/dt) in the above speed reduction
ratio control;
FIG. 10 is a flow chart illustrating speed reduction ratio
control in the second embodiment of the instant invention;
and
Fig. 1 shows a hydraulic circuit diagram of a continuously
variable speed transmission provided with a controller,
according to the first embodiment of the present invention.
The continuously variable speed transmission has a constant
displacement hydraulic pump P driven by the engine E through
the input shaft 1 and a variable displacement hydraulic motor
M connected to the output shaft 2 driving the wheels W. The
hydraulic pump P and motor M constitute a closed hydraulic
circuit along with two hydraulic lines: the first hydraulic
line La by which the delivery port of the pump P communicates
with the suction port of the motor M, and the second
hydraulic line Lb by which the suction port of the pump P
does with the delivery port of the motor M.
A charge pump 10 driven by the engine E is connected to the
closed circuit through a charge hydraulic line Lh having a
check
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valve 11 and through a third hydraulic line Lc having a pair of
check valves 3 and 3. Hydraulic oil pumped up by the charge
pump 10 from an oil sump 15 and regulated in pressure by a
charge pressure relief valve 12 is supplied to either of the
two lines La or Lb which has lower pressure through the
check valves 3 and 3. And, a fourth hydraulic line Ld having a
shuttle valve 4 is connected to the closed circuit. To the
shuttle valve 4 is connected a fifth and a sixth hydraulic line
which respectively have a high pressure relief valve 6 and a low
pressure relief valve 7 and are connected to the oil sump 15.
The shuttle valve 4 is a 2-port 3-position selector valve, which
is operated in response to a hydraulic pressur~ difference of the
first and second hydraulic lines to connect either of the first
or second hydraulic lines La having higher pressure with the
fifth hydraulic line Le as well as to connect the other having
lower pressure to the sixth hydraulic line Lf. Therefore, the
relief hydraulic pressure of a higher pressure-side line is
regulated by the high pressure relief valve 6, and the relief
hydraulic pressure of the other lower pressure-side line is
regulated by the low pressure relief valve 7.
Between the first and the second hydraulic lines La and Lb
is provided a seventh hydraulic line Lg through which the first
and second hydraulic lines can be communicated with each other.
The seventh hydraulic line Lg is provided with a clutch valve
5, a flow metering valve to control the opening degree of the
line Lg. The clutch valve 5 is actuated by a clutch servo unit
80 which is connected thereto through a link 88. Therefore,
the flow metering control of the clutch valve 5 by the
actuation of the clutch servo valve 80 can
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accomplish a clutch control for controlling transmission of
driving power from the hydraulic pump P to the hydraulic motor
M.
Actuators for controlling the speed reduction ratio of the
continuously variable speed transmission T by way of the
displacement control of the hydraulic motor M are the first and
the second ratio control servo units 30 and 50 connected to one
another by a link mechanism 40. The hydraulic motor M is a swash
plate type axial piston motor whose displacement is varied by the
control of swash plate angle by the ratio control servo units 30
and 50.
The actuations of the ratio control servo valves 30 and ~0
as well and the clutch servo valve 80 are respectivelY controlled
by pairs of solenoid valves 151, 152 and 155, 156 which are duty-
ratio-controlled by signals from a controller 100. The
controller 100 receives signals corresponding to such factors as
vehicle speed V, engine speed Ne, throttle opening ~th, swash plate
angle etr of the hydraulic motor M, accelerator pedal opening eacc
operated by a driver, atmospheric pressure Pat, oil temperature To,
water temperature Tw and clutch opening ecl. The controller
lO0, based on the above signals, outputs signals for controlling
the above solenoid valves to effectuate desirable traveling
control. Here, the engine throttle valve opening eth and the
accelerator pedal opening eacc to actuate the engine throttle
valve are parameters representing the driver's intention to
accelerate or decelerate. The accelerator pedal opening eacc is
"full-open" when the accelerator pedal is fully depressed, and
"full-closed" when fully released.
- 1 3 3 7 8 2 7
The structures and operations of the above servo units 30,
50 and 80 described in detai1 hereinafter.
Referring first to the ratio control servo unit 30, 50 shown
in FIGS. 1 and 2, the first ratio control servo unit 30 controls
the swash plate angle of the hydraulic motor M by the help of the
high hydraulic pressure fed from the closed hydraulic circuit of
the transmission T through the shuttle valve 4, the fifth line Le
and a high pressure line 120. The second ratio control servo
unit 50 is connected to the first ratio control servo unit 30 by
a link mechanism 40 and controls the operation of the first ratio
control servo unit 30.
The first ratio control servo unit 30 comprises a housing 31
having a connection port 31a connected to the hi~h pressure line
120, a piston member 32 slidably inserted into the housing 31,
and a spool member 34 slidably and coaxially inserted into the
piston member 32. The piston member 32 consists of a piston
portion 32a formed at its right end and a rod portion 32b
coaxially extending leftward. The piston portion 32a is fitted
into a cylinder hole 31c of the housing 31 and divides the space
inslde the cylinder hoie 31c inio two chambers defining .wo
cylinder chambers 35, 36. The rod portion 32b having a smaller
diameter than that of the cylinder hole 31c is inserted into a
rod hole 31d which is concentric with the cylinder hole 31c. The
right cylinder chamber 36 is covered by a plug member 33a and
cover 33b throu~h which the right end of the spool member 34
protrudes.
The high pressure line 120 connected to the port 31a is
communicated with the left cylinder chamber 35 through a
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hydraulic line 31b. The piston member 32 is pushed rightward by
the hydraulic pressure fed in the left cylinder chamber 35
throu~h the high pressure line 120.
A land portion 34a which is inserted in a spool hole 32d is
formed at the left end of the spool member 34. A pair of dents
34b having diagonal planes with fixed axial widths is formed at
the right side of the land portion 34a. A stop ring 37 mounted
on the spool member 34 hits against a stop ring 38 mounted on the
inside surface of the piston member 32 before the spool member 34
comes out.
A drain passage 32e which can connect the right cylinder
chamber 36 to the oil sump (not shown) through the spool hole 32d
responding to the rightward motion of the spool member 34 and a
connection passage 32c which can connects the left cylinder
chamber 35 to the right cylinder chamber 36 through the dents 34b
responding to the leftward motion of the spool member 34 are
formed in the piston member 32.
When the spool member 34 is moved rightward, the land
portion 34a blocks the connection passage 32c and opens the drain
passage ~2e. Accordingly the hydraulic pressure fed throu~h the
high pressure line 120 is led in the left cylinder chamber 35 and
pushes the piston member 32 rightward so that the piston member
32 follows the spool member 34. When the spool member 34 is
moved leftward, the connection passage 32c communicates with the
right cylinder chamber 36 through the dents 34b and the drain
passage 32e is bloc~ed by the land portion 34a. Accordingly the
high hydraulic pressure is fed to both the left and right
cylinder chambers 35, 36. The piston member 32 is pushed
- 10 -
1 337827
_,
leftward because of the difference in areas where pressurei~
applied and thérefore the piston member 32 is moved so as to
follow the spool member 34.
When the spool member 34 is held still, the piston member 32
is also held still creating a hydraulic floating state because of
pressure balance between the left and right cylinder chambers 35,
36.
As discussed, when the spool member 34 is moved leftward or
rightward, the piston member 32 is moved laterally so as to follow
the spool member 34 by the help of the high hydraulic pressure fed
though the high pressure line 120. Accordingly the variable
displacement of the motor M is controlled by the motion of the spool
member 34 since the piston member 32 is connected to the swash plate
73 of the motor M by means of a link member 39.
The spool member 34 is linked to the second servo unit 50 by
means of a link mechanism 40. The link mechanism 40 includes a
first link member 42 being swingable around an axis 42c and
having two arms 42a and 42b perpendicular to each other, and a
second link member 48 pivotally connected to the arm 42b. The
upper end of the arm 42a is pivotally connected to the righi end
of the spool member 34. The bottom end of the second link member
48 is pivotally connected to a spool member 54 of the second
servo unit 50. Therefore when the spool member 54 of the second
servo unit 50 is moved up or down, the spool member 34 of the
first servo unit 30 is moved rightward or leftward.
The second servo unit 50 comprises a housing 51 having ports
51a, 51b to which hydraulic lines 102, 104 are connected
respectively, and the spool member 54 vertically slidably fitted
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in the housin~ 51. The spool member 54 consists of a piston
portion 54a, an end spool portion 54b coaxiallY extending
downward and a rod portion 54c coaxiallY extending upward,
The piston portion 54a is inserted into a cylinder
hole 51c of the housing 51 and divides the space inside the
cylinder hole 51c covered by a cover 55 into two chambers
defining a rod-side ~upper) and a head-side (lower) cylinder
chamber 52, 53. The end spool portion 54b is fitted into a rod
hole 51d which is concentric with the cylinder hole 51c and
extends downward.
A spool 58a of a top position detecting switch 58 is
projected into a recess 54e formed on the end spool portion 54b.
The spool 58a is pushed up along the tapered surface of the
recess 54e when the spool member 54 is moved up. Therefore it
can be found by the top position detecting switch 58a if the
speed reduction ratio has become minimum since the pushed-up
spool 58a turns the switch 58 on.
Further, the hydraulic lines 102, 104 are communicated with
the rod-side and head-side cylinder chambers 52, 53 through the
ports 51a, 5lb. ~.e s~oGl member 5A ~5 moved up or do-~n ~y the
difference of hydraulic forces applied to the piston portion 54a
which are determined based on the differences of hydraulic
pressures and of areas where the hydraulic pressures in the
cylinder chambers 52, 53 are applied. The up and down motions of
the spool member 54 are transmitted to the spool member 34 of the
first servo unit 30 by the link mechanism 40 causing left and
right motions of the spool member 34. In other words, the
control of the hydraulic pressures supplied through the hydraulic
- 12 -
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lines 102, 104 enables control of the motion of the spool member 34
and the piston member 32 in the first servo unit 30 and also enables
control of the swash plate angle of the hydraulic motor M and the
displacement thereof. Hence, when the spool member 54 of the second
servo unit 50 is moved up, the piston member 32 of the first servo
unit 30 is moved rightward lessening the swash plate angle, the
displacement of the hydraulic motor M and the speed reduction ratio.
The pressurized oil in the hydraulic line 102 connecting the
port 51a to the rod-side cylinder chamber 52 is sent through a
hydraulic line 101 and 102 from the delivery line of the charge
pump 10 after its pressure is regulated by the charge pressure
relief valve 12. The oil pressure in the hydraulic line 104
connecting the port 51b to the head-side!cylinder chamber 53 is
obtained by regulating the oil pressure in a hydraulic line 103
(includlng an orifice 103a therein), which is branched from the
hydraulic line 102, by the first and second duty-ratio-controlled
solenoid valves 151 and 152. The first solenoid valve 151 duty-
ratio-controls the flow rate of the oil flowing from the
hydraulic lir.e 103 'having the orifice 10~a therein) to the
hydraulic line 104. The second solenoid valve 152 is disposed
between a hydraulic line 105 branched from the line 104 and a
hydraulic line 106 communicating with the drain through an
orifice 106a, and duty-ratio-controls drain-flow of the hydraulic
oil from the line 104 in accordance with a given duty ratio.
As a result, to the rod-side cylinder chamber 52 a charge
pressure regulated by the charge pressure relief valve 12 is
applied through the line 102, while to the head-side cylinder
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1 337827
chamber 53 is supplied from the line 104 a lower pressure than
the charge pressure which is regulated by the first and second
solenoid valves 151 and 152. Since the
pressure applied area of the rod-side cylinder chamber 52 is
smaller than that of the head-side cylinder chamber 53, the
forces of oil pressures in the cylinder chambers 52 and 53 acting
on the spool member 54 keep their balance when the oil pressure
in the head-side cylinder chamber 53 is a specified value P1
which is smaller than the oil pressure Pu in the rod-side
cylinder chamber 52 (Pu > P1). Therefore, when the oil pressure
supplied into the head-side cylinder chamber 53 from the line 104
is controlled by the first and second solenoid valves 151 and 152
so as to be higher than the above pressure P1, the spool member
54 is moved upward to have a small swash plate angle of the
hydraulic motor M, i.e., to have a small speed reduction ratio,
while when the oil pressure supplied into the head-side cylinder
chamber 53 from the line 104 is controlled so as to be smaller
than the above pressure P1, the spool member 54 is moved downward
to have a big swash plate angle of the hydraulic motor M, i.e.,
to have a big speed reduction ratio.
The solenoid valves 151 and 152 are controlled by
signals from the controller 100: only by controlling the
operations of the two solenoid valves 151 and 152 based on the
signals from the controller 100, the actuations of the first and
second ratio control servo units 30 and S0 are controlled, which
results is the control of the displacement of the hydraulic motor
M, in other words the control of speed reduction ratio.
The following is a detailed description of the construction
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of the clutch servo unit 80 based on FIG. 3. The clutch servo
unit 80 consists of a cylinder member 81, a piston member 82
inserted in the cylinder member 81 slidably to the right and left
in Fig. 3, a cover member 85 fixed to cover the cylinder chamber
into which the piston member 82 is inserted, and a spring 87
pushing the piston member 82 to the leftin Fig. 3. A piston 82a
on the piston member 82 divides aforesaid cylinder member 81
into a head-side (left) cylinder chamber 83 and a rod-side
(right) cylinder chamber 84 to which hydraulic lines 112
and 110 are respectively connected via respective ports 86a and
8~b.
The hydraulic oil in the hydraulic line 110 is transferred
from the charge pump 10 (whose delivery pressure is regulated by
the charge pressure relief valve 12) through a hydraulic line
101, while the hydraulic oil in the hydraulic line 112 is
transferred from the hydraulic line 101 through a hydraulic line
111. When the hydraulic oil is diverged into the line 111 and
passes through an orifice 111a in the line 111, hydraulic oil
pressure is controlled by the two duty-ratio-controlled solenoid
valves 155 and 15~. Here, the soienoid valve ;~6 is provided to
control the flow rate of the hydraulic oil flowing from the line
111 (having the orifice 111a therein) to the line 112 based on
the duty ratio signals, while the solenoid valve 155 is disposed
between a hydraulic line 113 diverging from the line 112 and a
hydraulic line 114 communicating with the drain throu~h an
orifice 114a, to control the flow of the hydraulic oil from the
line 113 to the drain based on the duty signals.
Therefore, to the rod-side cylinder chamber 84 via the line
- 15 -
- 1 3 3 7 8 27
110 is applied the charge pressure regulated by the charge
pressure relief valve 12, while to the head-side cylinder chamber
83 is applied a pressure from the line 112 lower than the charge
pressure, by the action of the aforesaid two solenoid valves 155
and 156. In this connection, the force applied on the piston
member 82 from the right side (that is, a force due to the
hydraulic pressure P1 in the rod-side cylinder chamber 84 plus
the force of the spring 87) balances with the force applied on
the piston member 82 from the left side ~that is, a force due to
the hydraulic pressure P2 in the head-side cylinder chamber 83),
even when P2 is lower than P1, because the area of the rod-side
cylinder chamber 84 subject to oil pressure is designed to be
much smaller than that of the head-side cylinder chamber 83.
Therefore, if the solenoid valves 155! and 156 control the
hydraulic pressure (in the head-side cylinder chamber 83)
supplied from the line 112 so as to be larger than the pressure
P2, the piston member 82 will be moved to the right, while when
the solenoid valves 155 and 156 control the hydraulic pressure in
the head-side cylinder chamber 83 supplied from the line 112 so
as to be smaller than the pressure P2, the piston member 82 will
be moved to the left.
The movement of the piston member 82 to the left or right is
transmitted to the clutch valve 5 through a link 88. The clutch
valve 5 consists of a stationary member 5a having a f;rst valve
port 5b therein, and a rotary member 5c having a second valve
port 5d rotatably inserted in the stationary member 5a. An arm
5e engaging with the rotary member Sc is connected to the
aforesaid link 88, thus allowing a rotation of the rotary member
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5c in accordance with the movement of the aforesaid piston member
82. When the rotary member 5c is rotated, the communication
between the first and second valve ports 5b and 5d varies from
"fully open" to "fully closed". When the piston member 82 is
moved to the leftmost as shown in Fig. 3,the communication in the
clutch valve 5 is "fully open", while as the piston member 82
moves to the right, the communication varies gradually to "fully
closed"
Because the first valve port 5b communicates with the first
line ~a and the second valve port Sd communicates with the second
line Lb constituting the hydraulic closed-circuit, the variation
in the communication between the aforesaid first and second valve
ports 5b and 5d can change the opening degree of the seventh line
Lg or the short circuit line of the first and second lines La and
Lb, thus effecting clutch control. In other words, based on
signal from the controller 100, duty-ratio control of the
solenoid valves 155 and 156 can perform a successful clutch
control.
With a continuously variable speed transmission having the
above configuration, methods of ratio control and clutch contrGl
by way of the operational control of the solenoid valves are
described.
First, the clutch control is performed as shown in the flow
chart in FrG. 4. This control starts with inputting accelerator
pedal opening eacc (or engine throttle opening eth) and vehicle
speed V. Then a reference clutch opening ecln is calculated
based on the vehicle speed V and the accelerator pedal opening e
acc. This calculation is made based on the graph shown in FIG.
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5A representing reference clutch opening ecln which is preset
corresponding to vehicle speed V for each accelerator pedal
opening eacc (1), eacc(2)~ .... eacc(n~ (where, eacc(l)
corresponds to the throttle "full-closed"and eacc(n) corresponds
to throttle "full-open"). The reference clutch opening ecln
varies in a range between O degreeSand 90 degrees.
Then, the difference decl ( = eclo - ecl)between the
reference clutch opening eclo and the actual clutch opening ecl
is calculated, to obtain a clutch opening control speed Scl using
the graph in FIG. 5B. The control speed Scl is, as apparent in
FIG. 5B, given in the "clutch-engagement" direction (ON
direction) to close the clutch valve when the difference decl is
positive, while it is given in the "clutch-disengagement"
direction (OFF direction) to open the clutch valve when the
difference decl is negative. The speed Scl increases in
proportion to the difference d~cl, but when the difference
exceeds a given value, the control speed Scl tends essentially to
a constant.
After the control speed Scl is obtained as described above,
i"e controller 100 Gutputs cGmmand signals to the sclenoi~ valves
155 and 156 so that the clutch valve 5 is actuated in accordance
with the control speed Scl.
Next, ratio control associated with the first embodiment is
described. First, the speed reduction ratio i ( = input
speed/output speed) is represented by equation (1):
i = N / ( C' * V ) .............................. (1)
where, N = engine speed, V = vehicle speed and C' is a
constant. Differentiation of the equation (1) by time t gives
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the equation (2) for change rate of speed reduction ratio di/dt:
di/dt = 1/(C'*V)*(dN/dt-N/(C'*V)*C'*dV/dt) ... (2)
In the equation (2), substitutions of a reference change
rate dNo/dt of engine speed for the change rate dN/dt of engine
speed, a predicted acceleration dVo/dt for the acceleration dV/dt
and 1/C for C' gives the e~uation (3):
di/dt = -C*(N/V2)*dVo/dt+ *(1/V)*dNo/dt .... (3)
As shown in the equation (3), the change rate di/dt of
speed reduction ratio can be interpreted as the sum of a
component dia/dt ( = -C * N/~2 * dVo/dt) corresponding to the
predicted acceleration dVo/dt and a component di~/dt ( = C * 1/V2
* dNo/dt) corresponding to the reference change rate dNo/dt of
engine speed. The predicted acceleration dVo/dt is obtained from
the following equations (4) to (7):
Output power Pe of the engine E (when transmission
efficiency is assumed 100~) is given by the following equation
(4);
Pe = Ru + Ra + Pa ............................... (4)
where, Ru = running resistance, Ra = air resistance, Pa
reserved power o engir.e ~.
From the equation (4), Pa is given by the following equation
(5);
Pa = Pe - (Ru + Ra) ............................. (5)
And, the reserved power can be given by the followin~
equation (6) too;
Pa = (W+dW)*(1/g)*(dVo/dt)*(V*103)/60Z*(1/75) .. (6)
where, W is total vehicle weight and dW is total engine
rotational mass.
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From the equations (63 and (5), the following equation
(7) is given:
dVo/dt = Pa*g*602/~(w+dW)*(V*103)3*75 (7)
Therefore, the predicted acceleration dVo/dt can be
calculated from reserved power Pa of engine E, and the
reserved power Pa can be obtained from the equation (5). On
the other hand, the reference change rate dNo/dt of engine
speed is determined by the following steps: calculating the
difference dN between the reference engine speed No~ which is
determined based on a parameter representing the driver's
intention to accelerate or decelerate, and the actual engine
speed N; and reading the reference change rate dN J dt from a
table in which the reference change rate dNJ dt is defined
corresponding to aforesaid difference dN so as to obtain
favourable running feeling and fuel consumption.
However, with regard to the change rate di/dt of speed
reduction ratio thus obtained from the equation (3), the
component (dia/dt) corresponding to the predicted
acceleration dVo/dt is determined based on only the reserved
power of engine E. The influences of running resistance in
the ascending or descending slope are not taken into account.
For this reason when the speed reduction ratio is controlled
based on the change rate di/dt of speed reduction ratio
calculated by the equation (3), in the ascending slope the
engine speed tends to fall under the reference speed because
of increased running resistance, while in the descending
slope it tends to rise over the reference speed because of
decreased running resistance.
Such being the case, in this ratio control, aforesaid
deviation from the reference speed is corrected using the
difference between the predicted acceleration dvo/dt and the
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actual acceleration dV/dt: thus the change rate di/dt of speed
reduction ratio is given by the following equation (8).
di/dt = -C1*N/V2*dVo/dt~C2*1/V*dNo/dt
+C3*N/~2*(dVo/dt-d~/dt)
where, C1, C2 and C3 are constants for weighting.
More particularly, the change rate di/dt of speed reduction
ratio in the ratio control is calculated as the sum of the
component dia/dt ( = -C1 * (N/V2) * dVo~dt ) corresPonding to the
predicted acceleration dVo/dt, the component di~/dt ( = C2 *
(1/V)*dNo/dt) corresponding to the reference change rate dNo/dt
of engine speed, and the component d(di/dt)( = C3 * (N/V2) *
(dVo/dt - dv/dt)t) for correction of predicted acceleration
component corresponding to the variations of running resistance.
At this time, weights on respective components can be adjusted
for obtaining proper ratio control characteristics in accordance
with actual vehic-le running.
Such calculation of the change rate di/dt of speed reduction
ratio as described above and the control of speed reduction ratio
based on the computed change rate di/dt is made by the controller
lOO. ~ere, ~he algor.thm of the ^o~trcl is illustrated using the
flow chart in FIG. ~:
First, in the first step S1, the engine speed N and the
vehicle speed ~ are inputted into the controller 100. In the
second step S2 is calculated the reserved engine power Pa. The
calculation is performed based on the equation (5), while the
engine output power Pe can be obtained from the map shown
typ~cally in FIG. 7, the abscissa represents the engine speed N,
and the ordinate does the engine output power Pe, based on a
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plurality of vacuums P1 to P13 of intake air 13. In other words,
the engine power Pe is obtained from the engine speed N and the
vacuum pressure of intake air.
Here, the engine output power Pe thus obtained is determined
without considering transmission efficiency. To calculate actual
power transmitted to the wheels, it must be corrected by the
product Ef M of the speed reduction ratio coefficient (function of
the speed reduction ratio i) and the transmission efficiency
(function of the engine output power Pe and the engine speed N).
For this reason, the reserved power Pa used in the second step S2
employs the corrected power obtained in FIG. 7 by said efficiency
E~M .
As described above, the reserved power Pa of the engine E
actually applied to the wheels is calculated, which gives the
predicted acceleration dVo/dt using the equation ~7) in the third
stepS3. Then, in the fourth step S4, the predicted acceleration
component dia/dt of the speed reduction ratio change rate is
computed.
In the fifth step S5, the reference change rate dNo/dt of
the engine speed is obtained: as shown in FIG. 8, the reference
change rate dNo/dt is determined based on the speed difference dN
between the reference engine speed No and the actual engine speed
N. In the sixth step s6, based on the reference change rate
dNo/dt, the component diN/dt of the change rate di/dt of speed
reduction ratio is obtained.
Next, in the seventh step S7, the correction component
d(di/dt) of the predicted acceleration is obtained: subroutine in
FIG. 9 shows the calculation. In the subroutine, the actual
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acceleration dV/dt is calculated by differentiatin~ the vehicle
speed V, and the difference (dVo/dt - dV/dt) between the
predicted acceleration and the actual acceleration is calculated.
Then using the engine speed N, the vehicle speed V, and the
weighting constant C3 already established, the correction
component d(di/dt) is calculated from the above equation C3 *
-(N/~2~ * (d~o/dt - dV/dt).
As described above, when all components di~/dt, di~/dt and
d(di/dt) are calculated, in the eighth step S~, as shown in
equation (8), these components are added to compute the change
rate di/dt of speed reduction ratio. Using this calculated value
di/dt as a control value, the controller 100 actuates and
controls the solenoid valves 151 and 152.
The predicted acceleration dVo/dt obtained from the equation
(~) based on the reserved output Pa is determined based on level
road running. However, even when there are variations in running
resistance due to ascending / descending, or exposure to tail
wind / head wind, since the change rate di/dt of speed reduction
ration is corrected by the correction component d(di/dt), this
ratio control has hardly ar.y poss.bility that the actual engine
speed deviates from the reference engine speed.
Next, a ratio control as a second embodiment is described.
Regarding the change rate di/dt of speed reduction ratio
thus obtained from the equation (3), the component corresponding
to the predicted acceleration dVo/dt is calculated from the
reserved power of the engine, but the reserved power of engine is
the one obtained when all of the engine output is directly
transmitted to the transmission. Therefore, for example, when
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the engine power is partly transmitted through the clutch, the
equation ~3) cannot be used directly as it is.
Accordingly, in this control to calculate the component for
the predicted acceleration using the engine power actually
delivered to the transmission, the equation ~3) is corrected by
a transmission rate of the clutch Kc~ into the following equation
(9) to obtain the change rate di/dt of speed reduction ratio:
di/dt = -C4*N/V2*dVo/dt*Kc~+C~*1/V*dNo/dt .... (9)
where, C~ and C~ are constants for weighting the two
components.
The calculation of change rate di/dt of speed reduction
ratio described above and the control of speed reduction ratio
based on thus computed change rate di/dt of speed reduction ratio
are performed by the controller 100. The control is detailed
referring to the flow chart in FIG. 10:
First, in the first step Sl, the engine speed Ne and the
vehicle speed Y is inputted to the controller 100, and in the
second step S2 the reserved engine power Pa is calculated. The
calculation of the reserved power Pa is performed using the
equation (5). As in t~e ~irst embodiment, the power Pe o the
engine E itself is obtained from the predefined map for instance
typically shown in FIG. 7. Further, the reserved power Pa
calculated in the second step S2 is corrected by the aforesaid
efficiency Efm
The reserved power Pa of the engine E is obtained as
described above, and in the third step S~ predicted acceleration
d~o/dt is obtained from the equation (7).
Next, in the fourth stepS4, clutch opening ~cl is inputted,
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and in the fifth step it is judged whether the clutch opening ecl
is larger than an opening es predetermined as a value close to
"full-close" condition. In other words, whether the clutch
opening ecl is essentially fully closed is judged, and if ecl > e
i.e., the clutch valve 5 is substantially fully closed, the
transmission rate of the clutch Kc~ is set to 1. On the
contrary, if ecl < es, the step is advanced to the sixth step S6
to read the transmission rate Kc. using FIG. 11. The
transmission rate Kc~ has been predefined for the clutch opening
ecl corresponding to each engine speed (Ne , Ne2, .. Ne~) as
shown in the graph In FIG. 11. Therefore, using this graph, the
transmission rate is obtained based on the engine speed Ne and
the clutch opening ecl.
Next in the ei~hth step S8, using the transmission rate Kc~,
the predicted acceleration component dia/dt for the change rate
di/dt of speed reduction ratio is calculated from the following
equation:
dia/dt = -C~ * (N/V2) * (dVo/dt) * Kc~
In the 9th step S9, the reference change rate dNo/dt of
engine spced i~ obta ~.ed. As shown in FIG. 8, the reference
change rate dNo/dt is predetermined for the difference dN between
the reference engine speed No and the actual engine speed N, and
the reference change rate dNo/dt is calculated based on the
difference dN. In the tenth step S10, based on the reference
change rate dNo/dt, the component diN/dt for the change rate
di/dt of speed reduction ratio corresponding to the reference
change rate dNo/dt of engine speed is calculated from the
following equation:
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._
diN/dt = C~ * (1 / V) * (dNo/dt)
After the both components dia/dt and di~/dt are calculated
as described above, in the eleventh step S11, the both components
are added to calculate the change rate di/dt of speed reduction
ratio as shown in the equation (9), and the controller 100
actuates and controls the solenoid valves 151 and 152 using the
calculated value di/dt as control value.
In the control used in the second embodiment, the change
rate di/dt of speed reduction ratio is calculated as the sum of
the component dia/dt ( = C4 * N/~2 * d~o/dt * Kc~) corresponding
to predicted acceleration dYo/dt, which is obtained based on the
actual input power to the transmission and corrected by the
transmission rate Kc~, and the component di~/dt ( = C~ * 1/Y *
dNoJdt) corresponding to the reference change rate dNo/dt of
engine speed, and the ratio control is performed using the change
rate di/dt of speed reduction ratio calculated as described
above. Therefore, both during clutch engaging and after clutch
engagement completion, the ratio control can be performed based
only on aforesaid change rate di/dt of speed reduction ratio. As
a result, the clutch engagement control condition is smoothly
shifted to the clutch engagement completion condition, thus
eliminating causes of unfavorable feelings during speed change,
and a rapid increases of engine speed.
Here the embodiments are adapted for a continuously variable
speed transmission having a hydraulic pump and a hydraulic motor,
but the application of the control methods according to the
invention are not limited to such a continuously variable speed
transmission, but may include other types of continuously
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variable speed transmission. Further, as a control device of
speed reduction ratio, not only such electrical-hYdraulic type
device as in the embodiments for controlling solenoid valves by
an electrical controller but also a hydraulic device for
producing oil pressure corresponding to throttle opening and a
actuating servo units by the oil pressure may be used.
The invention being thus described, it will be obvious
that the same may be varied in many ways. Such variations
are not to be regarded as a departure from the spirit and
scope of the invention, and all such modifications as would
be apparent to one skilled in the art are intended to be
included within the scope of the following claims.
