Feedback ratio change control for transmission

A ratio change control for a transmission, in which an increase in feedback control term is restrained when a gearing mechanism takes over a drive between an input shaft, and an output shaft or when a transition in drive from the gearing mechanism to a continuously variable transmission mechanism has not been completed. Specifically, the feedback control term is set to zero.

RELATED APPLICATIONS AND PATENTS 
U.S. patent application Ser. No. 07/278,887, filed Dec. 2, 1988 by Wataru 
ISHIMARU for a V-belt type continuously variable transmission (now U.S. 
Pat. No. 4,907,471 issued on Mar. 13, 1990). This Application corresponds 
to European Patent Application No. 88 120 186.7 
U.S. Pat. No. 4,955,260, filed on Mar. 31, 1989 for Toshikazu OSHIDARI for 
a hydraulic control system for a transmission. 
U.S. patent application Ser. No. 07/330,918, filed on Mar. 31, 1989 by 
Keiju ABO et al. for a control system for a transmission (now U.S. Pat. 
No. 4,895,552 issued on Jan. 23, 1990). This Application corresponds to 
European Patent Application No. 89 105 734.1. 
U.S. patent application Ser. No. 07/336,422, filed on Apr. 11, 1989 by 
Toshikazu OSHIDARI for a line pressure control for a V-belt type 
continuously variable transmission. 
U.S. patent application Ser. No. 07/348,837, filed on May 8, 1989 by Wataru 
ISHIMARU for a system for controlling hydraulic fluid pressure for a 
V-belt type automatic transmission. 
U.S. patent application Ser. No. 07/489,058, filed on Mar. 7, 1990 by 
Toshifumi HIBI for a control system for transmission. This application 
corresponds to European Patent Application No. 90104376.0. 
U.S. Pat. No. 5,021,031, claiming priority on Japanese Patent Application 
No. 1-192566. 
BACKGROUND OF THE INVENTION 
The present invention relates to a ratio change control for a transmission, 
and more particularly to a ratio change control for a hybrid continuously 
variable transmission including a gearing mechanism and a continuously 
variable transmission mechanism which are selectively rendered operabble 
to take over a drive from a transmission input shaft to a transmission 
output shaft. 
U.S. Pat. No. 4,735,113 discloses a V-belt type continuously variable 
transmission including a driver pulley, a follower pulley, and a V-belt 
drivingly interconnecting these pulleys. Laid-open Japanese Patent 
Application No. 63-176862 and U.S. Pat. No. 4,907,471 disclose a hybrid 
continuously variable transmission in which the above-mentioned 
continuously variable transmission mechanism is combined with a gearing 
mechanism such that the gearing mechanism is put into operation to provide 
a reduction ratio between a transmission input shaft and a transmission 
output shaft for start-up operation. The setting is such that this 
reduction ratio provided by the gearing mechanism is larger than the 
maximum or largest reduction ratio provided by the continuously variable 
transmission mechanism. For high speed operation where a relatively small 
drive force is required, the continuously variable transmission mechanism 
is put into operation to take over the drive from the transmission input 
shaft to the transmission output shaft owing to engagement of a power 
interruption device, such as a clutch. 
Referring to FIG. 9, consider the case where the control strategy shown in 
U.S. Pat. No. 4,735,113 is applied to the hybrid continuously variable 
transmission of the type mentioned above. FIG. 9 is a graph wherein a line 
G shows target driver pulley revolution speed values versus varying 
vehicle speed values when a gearing mechanism is put into operation to 
take over a drive. In FIG. 9, line L shows target driver pulley revolution 
speed values versus vehicle speed values when a continuously variable 
transmission mechanism provides its maximum or largest reduction ratio, 
while line S shows target driver pulley revolution speed values versus 
vehicle speed values when the continuously variable transmission mechanism 
provides its minimum or smallest reduction ratio. Let it be assumed that a 
vehicle starts moving with a 3/8 throttle opening degree. In this case, 
the driver pulley revolution speed increases along the line G from a point 
O to a point A with the reduction ratio provided by the gearing mechanism. 
Upon arriving at point A, a transition is made from the drive owing to the 
gearing mechanism to a drive owing to the continuously variable 
transmission. This causes a drop in driver pulley revolution speed without 
any substantial change in vehicle speed, resulting in a transfer from the 
point A to a point B on the line L. Then, with the maximum reduction ratio 
provided by the continuously variable transmission mechanism, the vehicle 
increases its speed till a point C on the line L. As the vehicle increases 
its speed further from a vehicle speed value corresponding to the point C, 
the reduction ratio decreases continuously toward the minimum reduction 
ratio provided by the continuously variable transmission mechanism along 
an operation line I. According to this control strategy, the continuously 
variable transmission mechanism effects a ratio change from the maximum 
reduction ratio to a relatively smaller reduction ratio on a broken line 
M, since a target driver pulley revolution speed corresponding to the 
point C' is set if the vehicle speed increases to a vehicle speed value 
corresponding to the point C'. A problem arises if, during transition from 
the drive owing to the gearing mechanism to the drive owing to the 
continuously variable transmission mechanism, the vehicle speed increases 
beyond the vehicle speed value corresponding to the point C and the 
continuously variable transmission has shifted off the maximum reduction 
ratio thereof. The transition in drive takes a relatively long time, since 
there is a delay in the hydraulic system for activating a clutch 
contributing to this transition. In this case, the vehicle starts running 
through the drive owing to the continuously variable transmission 
mechanism with a reduction ratio smaller than the maximum reduction ratio 
thereof. This transition inherently causes a substantial shock, which 
causes an insufficient driving force and thus poor acceleration. 
Another hybrid continuously variable transmission is known. This known 
hybrid continuously variable transmission has a gearing mechanism which 
provides a reduction ratio smaller than the minimum reduction ratio 
provided by a continuously variable transmission mechanism. Such a hybrid 
continuously variable transmission is disclosed in Laid-open Japanese 
Patent Application No. 58-156764. If the continuously variable 
transmission mechanism is controlled according to the above-mentioned 
control strategy, there occurs a case where the continuously variable 
transmission mechanism shifts down from the minimum reduction ratio 
thereof during a transition from a drive owing to the gearing mechanism to 
a drive owing to the continuously variable transmission mechanism. This 
causes a substantial increase in engine speed, inducing a substantial 
shock. 
An object of the present invention is to provide a ratio change control for 
a transmission which assures smooth and shockless transition between a 
drive owing to a gearing mechanism and a drive owing to a continuously 
variable transmission mechanism. 
More specifically, the present invention provides a ratio change control 
employing a control action including a feedback control term in addition 
to a feedforward control term. 
SUMMARY OF THE INVENTION 
According to one aspect of the present invention, there is provided a ratio 
change control for a transmission, the transmission including an input 
shaft, an output shaft, a gearing mechanism and a continuously variable 
transmission mechanism, the gearing mechanism and continuously variable 
transmission mechanism being operatively combined with each other such 
that they alternatively take over a drive between the input and output 
shafts, the ratio change control comprising: 
means for determining an actual value of a predetermined operating variable 
related to a revolution speed of the input shaft and generating an actual 
value indicative signal indicative of said actual value detected; 
a control unit including 
means for determining a target value of said predetermined operating 
variable and generating a target value indicative signal indicative of 
said target value determined, 
means for calculating a deviation between said actual value indicative 
signal and said target value indicative signal and generating a deviation 
indicative signal indicative of said deviation calculated, 
means for calculating a feedback control term as a function of said 
deviation indicative signal and generating a feedback control term 
indicative signal indicative of said feedback control term calculated, and 
means for calculating a feedforward control term as a function of said 
target value indicative signal and generating a feedforward control term 
indicative signal indicative of said feedforward control term calculated; 
and 
means for controlling a change in reduction ratio in the continuously 
variable transmission in response to said feedback control term indicative 
signal and said feedforward control term indicative signal; 
said control unit also including means for limiting an increase in said 
feedback control term during periods when the gearing mechanism takes over 
a drive between the input and output shafts and when a transition in drive 
from the gearing mechanism to the continuously variable transmission 
mechanism has not been completed yet. 
According to another aspect of the present invention, there is provided a 
method of ratio change control for a transmission, the transmission 
including an input shaft, an output shaft, a gearing mechanism and a 
continuously variable transmission mechanism, the gearing mechanism and 
continuously variable transmission mechanism being operatively combined 
with each other such that they alternatively take over a drive between the 
input and output shafts, the method comprising the steps of: 
determining an actual value of a predetermined operating variable related 
to a revolution speed of the input shaft and generating an actual value 
indicative signal indicative of said actual value detected; 
determining a target value of said predetermined operating variable and 
generating a target value indicative signal indicative of said target 
value determined; 
calculating a deviation between said actual value indicative signal and 
said target value indicative signal; and generating a deviation indicative 
signal indicative of said deviation calculated; 
calculating a feedback control term as a function of said deviation 
indicative signal and generating a feedback control term indicative signal 
indicative of said feedback control term calculated; 
calculating a feedforward control term as a function of said target value 
indicative signal and generating a feedforward control term indicative 
signal indicative of said feedforward control term calculated; 
controlling a change in reduction ratio in the continuously variable 
transmission in response to said feedback control term indicative signal 
and said feedforward control term indicative signal; 
controlling a transition in drive between the gearing mechanism and the 
continuously variable transmission mechanism; and 
limiting an increase in said feedback control term during periods when the 
gearing mechanism takes over a drive between the input and output shafts 
and when a transition in drive from the gearing mechanism to the 
continuously variable transmission mechanism has not been completed yet. 
According to further aspect of the present invention, there is provided a 
method of ratio change control for a transmission, the transmission 
including an input shaft, an output shaft, a gearing mechanism and a 
continuously variable transmission mechanism, the gearing mechanism and 
continuously variable transmission mechanism being operatively combined 
with each other such that they alternatively take over a drive between the 
input and output shafts, the method comprising the steps of: 
determining an actual value of a predetermined operating variable related 
to a revolution speed of the input shaft and generating an actual value 
indicative signal indicative of said actual value detected; 
determining a target value of said predetermined operating variable and 
generating a target value indicative signal indicative of said target 
value determined; 
calculating a deviation between said actual value indicative signal and 
said target value indicative signal and generating a deviation indicative 
signal indicative of said deviation calculated; 
calculating a feedback control term as a function of said deviation 
indicative signal and generating a feedback control term indicative signal 
indicative of said feedback control term calculated; 
calculating, as a feedforward control term, a target reduction ratio as a 
function of said target value indicative signal and generating a target 
reduction ratio indicative signal indicative of said target reduction 
ratio calculated; 
controlling a change in reduction ratio in the continuously variable 
transmission in response to said feedback control term indicative signal 
and said feedforward control term indicative signal; 
determining whether said target reduction ratio falls in a predetermined 
range of reduction ratios, and generating a determined result indicative 
signal when said target reduction ratio falls in said predetermined range 
of reduction ratios; and 
setting said feedback control term to zero in response to said determined 
result indicative signal. 
According to still further aspects of the present invention, there is 
provided a method of a ratio change control for a transmission, the 
transmission including an input shaft, an output shaft, a gearing 
mechanism and a continuously variable transmission mechanism, the gearing 
mechanism and continuously variable transmission mechanism being 
operatively combined with each other such that they alternatively take 
over a drive between the input and output shafts, the method comprising 
the steps of: 
determining an actual value of a predetermined operating variable related 
to a revolution speed of the input shaft and generating an actual value 
indicative signal indicative of said actual value detected; 
determining a target value of said predetermined operating variable and 
generating a target value indicative signal indicative of said target 
value determined; 
calculating a deviation between said actual value indicative signal and 
said target value indicative signal and generating a deviation indicative 
signal indicative of said deviation calculated; 
calculating a feedback control term as a function of said deviation 
indicative signal and generating a feedback control term indicative signal 
indicative of said feedback control term calculated; 
calculating, as a feedforward control term, a target reduction ratio as a 
function of said target value indicative signal and generating a target 
reduction ratio indicative signal indicative of said target reduction 
ratio calculated; 
controlling a change in reduction ratio in the continuously variable 
transmission in response to said feedback control term indicative signal 
and said feedforward control term indicative signal; 
calculating an actual reduction ratio based on said actual value indicative 
signal and generating an actual reduction ratio indicative signal 
indicative of said actual reduction ratio calculated; 
determining whether said actual reduction ratio indicative signal falls in 
a predetermined range of reduction ratio, and generating a determined 
result indicative signal when said actual reduction ratio indicative 
signal falls within in said predetermined range of reduction ratio; and 
setting said feedback control term equal to zero in response to said 
determined result indicative signal.

DETAILED DESCRIPTION OF THE EMBODIMENT 
Referring to the accompanying drawings, a plurality of preferred 
embodiments of a control system, according to the present invention, are 
described. 
Referring now to FIG. 1, a power train of a motor vehicle is described. In 
FIG. 1, an engine 10 is shown as having an output shaft 10a to which a 
torque converter 12 is coupled in the conventional manner. The torque 
converter 12 includes a pump impeller 12a, a turbine runner 12b and a 
stator 12c. It also includes a lock-up clutch 12d with which the pump 
impeller 12a and turbine runner 12b are selectively interconnected with. 
The turbine runner 12b of the torque converter 12 is drivingly connected 
to a turbine shaft or a driver shaft 14. On the driver shaft 14 is a 
driver pulley 16. The driver pulley 16 includes an axially stationary 
conical member 18 fixedly connected to the driver shaft 14, and an axially 
moveable conical member 22 connected to the driver shaft 14 in opposed 
spaced relationship with the stationary conical member 18. The conical 
members 18 and 22 define therebetween a V-shaped pulley groove. The driver 
pulley 16 includes a driver pulley cylinder chamber 20. The moveable 
conical member 22 is axially moveable toward the stationary conical member 
18 in response to hydraulic pressure developed in the driver pulley 
cylinder chamber 20, causing the V-shaped pulley groove to decrease its 
width. The driver pulley 16 is drivingly connected via a V-belt 24 to a 
follower pulley 26. The follower pulley 26 includes an axially stationary 
conical member 30 fixedly connected to a follower shaft 28, and an axially 
moveable conical member 34 connected to the follower shaft 28 for axial 
movement. The conical members 30 and 34 define a V-shaped pulley groove 
therebetween. The follower pulley 26 includes a follower pulley cylinder 
chamber 32. The moveable conical member 34 is axially moveable toward the 
stationary conical member 30 in response to hydraulic pressure developed 
in the follower pulley cylinder chamber 32. The driver pulley 16, V-belt 
24, and follower pulley 26 form a continuously variable transmission 
mechanism. The maximum reduction ratio provided by this continuously 
variable transmission mechanism is smaller than a reduction ratio provided 
by a gearing mechanism including a driver shaft side forward gear 42 and 
an output shaft side forward gear 48, which are described later. The 
driver shaft 14 extends through a hollow shaft 36. The hollow shaft 36 is 
rotatably supported on the driver shaft 14. Rotatably supported on the 
hollow shaft 36 are a reverse gear 38 and a forward gear 42. The forward 
gear 42 is connectable to the hollow shaft 36 by means of a hydraulic 
fluid operated forward clutch 52, while the reverse gear 38 is connectable 
to the hollow shaft 36 by means of hydraulic fluid operated reverse clutch 
53. With a hydraulic fluid operated low clutch 44 engaged, the driver 
shaft 14 is connected to the hollow shaft 36. Extending in parallel to the 
driver shaft 14 is an output shaft 46. The forward gear 14 is mounted via 
a one-way clutch 40 to the output shaft 46, while a reverse gear 50 is 
mounted for unitary rotation with the output shaft 46. The forward gear 48 
is in constant mesh with the forward gear 42. The reverse gear 50 is in 
constant mesh with a reverse idler gear 56, rotatable with an idler shaft 
54. The reverse idler gear 56 is also in constant mesh with the reverse 
gear 38 too. In FIG. 1, since it is impossible to illustrate them in the 
cross sectional plane, the reverse idler shaft 54 and reverse idler gear 
56 are illustrated by the broken line. Actually, they are arranged as 
illustrated in FIG. 2. In FIG. 1, the distance between the shafts and the 
diameter of each of the gears do not reflect the actual dimension. Thus, 
in order to know the actual relationship, reference should be made to FIG. 
2. 
The follower shaft 28 has a forward gear 58. Via a hydraulic fluid operated 
high clutch 60, the forward gear 58 is connectable to the follower shaft 
28. As best seen in FIG. 2, the forward gear 58 is in constant mesh with 
the reverse gear 50. The forward gear 58 and reverse gear 50 have the same 
diameter. The output shaft 46 has a reduction gear 62 for rotation 
therewith. The reduction gear 62 is in constant mesh with a final gear 64 
of a differential 66. The differential 66 includes a pair of pinion gears 
68 and 70, which are rotatable with the final gear 64. A pair of side 
gears 72 and 74 mesh with the pinion gears 68 and 70. The side gears 72 
and 74 are coupled with drive axles 76 and 78, respectively, for rotation 
therewith. 
The neutral state is provided when the low clutch 44 and the high clutch 60 
are both released. In this state, the transmission of rotational power 
from the driver shaft 14 to the output shaft 46 is interrupted. 
On start-up or hill-climbing where a relatively large driving force is 
required, the forward clutch 52 is engaged and the low clutch 44 engaged. 
The high clutch 60 is released. In this state, the rotational power of the 
output shaft 10a of the engine 10 is transmitted via the torque converter 
12 to the driver shaft 14, and further to the hollow shaft 36 via the 
engaged low clutch 44. The torque of the hollow shaft 36 is transmitted 
via the forward clutch 52 to the forward gear 42, and further to the 
forward gear 48 with which the gear 42 meshes with. Owing to the fact that 
the forward gear 48 is drivingly connected via the one-way clutch 40 to 
the output shaft 46, the rotational power is transmitted to the output 
shaft 46. Thereafter, the rotational power is transmitted via the 
reduction gear 62, and the final gear 64, to the differential 66 where it 
is distributed between the drive axles 76 and 78, causing the drive wheels 
of the vehicle, not illustrated, to rotate. During the transmission of 
rotational power mentioned above, the rotational power is not transmitted 
through the continuously variable transmission mechanism, but through the 
gearing mechanism. With the reduction ratio provided by the intermeshed 
forward gears 42 and 48, the rotational power is transmitted to the output 
shaft 46, thus providing a relatively large driving force. 
When the operating condition progresses and now demands less driving force, 
the high clutch 60 is engaged with the above described state maintained. 
This causes the rotational power to be transmitted through the 
continuously variable transmission. The rotational power of the driver 
shaft 14 is transmitted, via the V-belt 24 and the follower pulley 26, to 
the follower shaft 28, and further to the forward gear 58 via the high 
clutch 60 that is engaged. Since the forward clutch 58 meshes with the 
reverse gear 50, the rotational power is transmitted to the output shaft 
46, and further to the drive axles 76 and 78 via the same power delivery 
path as previously described. In this case, the output shaft 46 rotates at 
a higher speed than the forward gear 48 does, and thus the one-way clutch 
40 idles. This allows the low clutch 44 to be kept engaged. In the manner 
as described above, the rotational power is transmitted through the 
continuously variable transmission mechanism. Thus, the reduction ratio 
can be varied continuously by varying the width of the V-groove of the 
driver pulley 26 which in turn induces variation in the width of the 
V-shaped groove of the follower pulley 26. 
For reverse drive, the reverse clutch 53 is engaged, the low clutch 44 is 
engaged, and the high clutch 60 is released. The engagement of the reverse 
clutch 53 causes the reverse gear 38 to be connected to the hollow shaft 
36 for unitary rotation. In this state, the rotational power of the driver 
shaft 14 is transmitted via the low clutch 44, the hollow shaft 36, the 
reverse clutch 53, the reverse gear 38, the reverse idler gear 56 and the 
reverse gear 50 to the output shaft 46. Since the reverse idler gear 56 is 
operatively disposed in the power delivery path, the direction of rotation 
of the output shaft 46 is opposite the direction of rotation of the output 
shaft 46. Thus, the vehicle can move in the reverse direction. 
A hydraulic control system for the hybrid continuously variable 
transmission is now described. As shown in FIG. 3, the control system 
comprises an oil pump 101, a line pressure regulator valve 102, a manual 
valve 104, a shift control valve 105, a direction control valve 108 (which 
may be called a shift command valve) for controlling the direction of 
adjustment pressure within a hydraulic fluid line 190, a shift operating 
mechanism 112, a throttle valve 114, a constant pressure regulating valve 
116, a solenoid valve 118, a torque converter pressure regulating valve 
120, and a lock-up control valve 122. 
The shift control valve 106 has a valve bore 172 provided with five ports 
172a, 172b, 172c, 172d and 172e, a spool 174 having three axially spaced 
lands 174a, 174b, and 174c slidably fit in the valve bore 172, and a 
spring 175 biasing the spool 174 to the left as viewed in FIG. 3. The port 
172b communicates via a hydraulic fluid conduit 176 with the driver pulley 
cylinder chamber 20, and this conduit 176 communicates with the high 
clutch 60 at its servo chamber. The port 172a and the port 172e are drain 
ports, respectively. An orifice 177 is provided at the drain port 172a. 
The port 172d communicates via a hydraulic fluid conduit 179 with the 
follower pulley cylinder chamber 32. The port 172c communicates with a 
hydraulic fluid conduit 132 that serves as a line pressure circuit and, 
thus, is supplied with the line pressure. The spool 174 has a lefthand 
end, as viewed in FIG. 3, rotatably linked via a pin 181 to a middle 
portion of a lever 178 of the shift operating mechanism 112, later 
described in detail. The land 174b has an axial section with a curved 
contour. This allows a portion of hydraulic fluid supplied from the line 
pressure port 172c to flow into the port 172a. Thus, the pressure at the 
port 172b is determined by a ratio of the amount of hydraulic fluid 
flowing from the port 172c toward the port 172b to the amount of hydraulic 
fluid discharged out of the drain port 172a. If the spool 174 moves to the 
left as viewed in FIG. 3, this leftward movement of the spool 174 causes 
the degree of opening of a clearance on the line pressure side of the port 
172b to increase, and the degree of opening of a clearance on the 
discharge side of the port 172b to decrease. This results in an increase 
in pressure at the port 172b. The port 172d is always supplied with the 
line pressure from the port 172c. The hydraulic presure developed at the 
port 172b is supplied via the conduit 176 to the driver pulley cylinder 
chamber 20, while the hydraulic pressure developed at the port 172d is 
supplied to the follower pulley cylinder chamber 32. 
Therefore, leftward movement of the spool 174, as viewed in FIG. 3, causes 
an increase in the hydraulic pressure developed in the driver pulley 
cylinder chamber 20, resulting in a decrease in the width of the V-shaped 
pulley groove of the driver pulley 16. This also results in an increase in 
the width of the V-shaped pulley groove of the follower pulley 26, since 
the V-belt 26 is wedged into the V-shaped groove of the follower pulley 
26. Therefore, the reduction ratio becomes small since the radius of the 
running diameter of the V-belt on the driver pulley 16 increases, but the 
radius of the running diameter of the V-belt 24 on the follower pulley 26 
decreases. The reduction ratio becomes large when the spool 174 is urged 
to move to the right as viewed in FIG. 3. 
The lever 178 of the shift operating mechanism 112 has its middle portion 
linked via a pin 181 to the spool 174 of the shift control valve 106. The 
lever 178 has one or lower end, as viewed in FIG. 3, linked via a pin 183 
to a reduction ratio transmission member 158 and the opposite or an upper 
end linked via a pin 185 to the rod 182 of the direction control valve 
108. The rod 182 is formed with a rack 182c with which a pinion gear 110a 
of a shift motor 110 in the form of a stepper motor meshes. 
According to this shift operating mechanism 112, rotating the pinion gear 
110a of the shift motor 110 in such a direction as to displace the rod 182 
to the right, as viewed in FIG. 3, causes the lever 178 to swing clockwise 
about the pin 183. This clockwise movement of the lever 178 causes the 
spool 174 of the shift control valve 106 to move to the right, as viewed 
in FIG. 3. The rightward movement of the spool 174 of the shift control 
valve 106 causes a reduction in hydraulic pressure within the driver 
pulley chamber 20, causing the axially moveable conical member 22 of the 
driver pulley 16 in such a direction as to increase the width of the 
V-shaped pulley groove. Viewed in FIG. 3, the conical member 22 moves to 
the left. Since the end of the lever 178 is connected via the pin 183 to 
the reduction ratio transmission member 158, this leftward movement of the 
conical member 22, as viewed in FIG. 3, causes the lever 178 to swing 
clockwise about the pin 185. This clockwise movement of the lever 178 
displaces the spool 174 to the left, as viewed in FIG. 3, causing an 
increase in hydraulic pressure within the driver pulley chamber 20, thus 
causing the reduction ratio to become small. As a result, the driver 
pulley 16 and the follower pulley 26 assume one of stable states, 
providing one reduction ratio. Each stable state corresponds to one rotary 
position which the shift motor 110 takes. If the shift motor 110 rotates 
in the opposite or reverse direction, the above-mentioned process 
progresses. The rod 182 is moveable to the right, as viewed in FIG. 3, 
beyond a position at which the maximum and largest reduction ratio is 
induced toward an overstroke position. When the rod 182 has moved toward 
and stayed at the overstroke position, a change-over switch 298 is 
activated. The output signal of this switch 298 is fed to a control unit. 
It will now be recognized that when the shift motor 110 is operated in a 
predetermined shift pattern, the variation of the reduction ratio follows 
this pattern. Therefore, the variation in reduction ratio in the 
continuously variable transmission mechanism is controlled by controlling 
the shift motor 110. 
The rotary position which the shift motor 110 takes is determined by a 
number of pulses supplied to the shift motor 110 by a control unit 300 
shown in FIG. 4. The control unit 300 stores a plurality of shift patterns 
and generates the number of pulses in accordance with one shift pattern 
selected out of all. 
The direction control valve 108 includes a valve bore 186 provided with 
ports 186a, 186b, 186c and 186d, and a rod 182 with lands 182a and 182b 
received in the valve bore 186. The port 186a communicates with a 
hydraulic fluid conduit 188. The port 186b communicates via a hydraulic 
fluid conduit 190 with the solenoid valve 118. The port 186c communicates 
with a hydraulic fluid conduit 189. The port 186d is a drain port. 
Normally, the ports 186a and 186b communicate with each other via a space 
defined between the lands 182a and 182b. When the rod 182 moves beyond the 
position corresponding to the maximum reduction ratio toward the 
overstroke position, the port 186a is covered by the land 182a, while the 
port 186b is allowed to communicate with the port 186c. The 
above-mentioned hydraulic fluid conduit 189 communicates with the low 
clutch 44. 
The other valves illustrated in FIG. 3 are substantially the same as their 
counterparts disclosed in JP 61-105351. The hydraulic circuit except the 
low clutch 44 and the high clutch 60 is substantially the same as a 
hydraulic control system disclosed in European Patent Application 
published under publication number 0180209 on May 7, 1986 or U.S. Pat. No. 
4,735,113 issued to Yamamuro et al. on Apr. 5, 1988. 
Referring to FIG. 4, the control unit 300 comprises an input interface 311, 
a reference pulse generator 312, a central processor unit (CPU) 313, a 
read only memory (ROM) 314, a random access memory (RAM) 315 and an output 
interface 316. They are operatively interconnected by an address bus 319 
and a data bus 320. Output signals of an engine speed sensor 305, a 
vehicle speed sensor 302, a throttle sensor 303, a shift position switch 
304, a turbine speed sensor 305, an engine coolant temperature sensor 306, 
a brake sensor 307, and a change-over switch 298. These output signals are 
supplied directly or indirectly via wave shapers 308, 309, and 322, and an 
A/D converter 310. Output signals of the control unit 300 are supplied via 
an amplifier and leads 317a, 317b, 317c and 317d to the shift motor 110. 
The solenoid 224 is also under the control of the control unit 300. For 
further understanding of the control unit 300, reference should be made to 
U.S. Pat. No. 4,735,113, mentioned before. 
The operation of this embodiment is described. When the rotation of motor 
110 toward the maximum and largest reduction ratio causes the rod 182 to 
move beyond the predetermined position corresponding to the maximum 
reduction ratio to the overstroke position, the direction control valve 
108 takes a position as illustrated by the lower half thereof, viewed in 
FIG. 3. In this state of the direction control valve 108, the conduit 190 
is allowed to communicate with the conduit 189, establishing a state where 
the hydraulic fluid pressure adjusted by the solenoid valve 118 is 
supplied to the low clutch 44. Thus, the torque capacity of the low clutch 
44 becomes adjustable by the solenoid valve 118. Upon the direction 
control valve 108 staying in the overstroke position, the spool 174 of the 
shift control valve 106 which is linked via the lever 178 to the rod 182 
is displaced to the right as viewed in FIG. 3 and, thus, the port 172b is 
allowed to communicate with the drain port 172a. Thus, the driver pulley 
cylinder chamber 20 and high clutch 60 are drained. As a result, the high 
clutch 60 is released with the low clutch 44 left engaged and thus the 
gearing mechanism takes over a drive between the driver shaft 14 and the 
output shaft 46. The torque is therefore multiplied at the reduction ratio 
determined by the forward gears 42 and 48. 
When the operating condition progresses and demands a less driving force, 
the shift motor 110 rotates toward the minimum and smallest reduction 
ratio, the rod 182 moves toward the minimum reduction ratio beyond the 
predetermined position corresponding to the maximum reduction ratio. This 
movement of the rod 182 causes the lever 178 to displace the spool 174 to 
the left, as viewed in FIG. 3. This causes the land 174b of the spool 174 
to close communication of the port 172b with the drain port 172a and open 
communication of the port 172b with the line pressure port 172c, allowing 
the supply of hydraulic fluid to the port 172b from the line pressure port 
172c. Thus, the driver pulley cylinder chamber 20 is supplied with the 
hydraulic fluid pressure and the high clutch 60 is also supplied with the 
hydraulic fluid pressure and engaged. Engagement of the high clutch 60 
causes a transition from a drive owing to the gearing mechanism to a drive 
owing to the continuously variable transmission mechanism. Upon completion 
of the transition, the continuously variable transmission takes over a 
drive between the shafts 14 and 46. Thereafter, the rotational power is 
transmitted by the continuously variable transmission mechanism and the 
reduction ratio is allowed to continuously vary. Under this operating 
condition, although the low clutch 40 is kept engaged, since the one-way 
clutch 40 idles, the rotational power is transmitted via the continuously 
variable transmission mechanism. The direction control valve 108 allows 
the conduit 188 to communicate with the conduit 190, rendering the 
solenoid valve 118 to control the lock-up control valve 122. Thus, the 
engagement of the lock-up clutch 12d is controlled by the solenoid valve 
118. The valve 108 assumes the overstroke position when the spool 136 of 
the manual valve 104 is placed at P or R or N position. Since the valve 
108 is in the overstroke position and the solenoid valve 118 drains the 
conduit 190 when the spool 136 of the manual valve 104 is placed at P or N 
position, no hydraulic fluid pressure is applied to the low clutch 44, 
leaving the same released. 
Referring to the flowchart of FIG. 5, it is now explained how the control 
unit 300 operates during transition from the drive owing to the gearing 
mechanism to the drive owing to the continuously variable transmission 
mechanism. The ROM 314 of the control unit 300 stores a program as 
illustrated by the flowchart in FIG. 5. 
In FIG. 5, at a step 400, a reading operation is performed, based on the 
output signal of the throttle sensor 303 to store the result as a throttle 
opening degree TH. At the subsequent step 402, a reading operation is 
performed, based on the output signal of the vehicle speed sensor 302 to 
store the result as a vehicle speed V. At a step 404, a reading operation 
is performed, based on the output signal of the turbine speed sensor 305 
to store the result as an actual driver pulley revolution speed N.sub.a. 
At a step 406, a table look-up operation of a shift point mapping data 
stored in the ROM 314 is performed, using the stored throttle opening 
degree TH and vehicle speed V in order to determine and store the result 
as a target driver pulley revolution speed N.sub.t. At a step 408, using 
the data N.sub.t and V stored at steps 406 and 402, an equation I.sub.t 
=N.sub.t /V is calculated to store the result as a feedforward reduction 
ratio I.sub.f. At a step 410, the data N.sub.a is subtracted from the data 
N.sub.t to store the result as a deviation e. At a step 412, the integral 
of the deviation e is calculated to store the result as an intgral S. At a 
step 414, a feedback reduction ratio is calculated using the integral S as 
a variable to store the result as the feedback reduction ratio I.sub.b. At 
a step 416, a constant is stored as a reference value I.sub.k. At a 
comparison step 418, it is determined whether the feed forward reduction 
ratio I.sub.f is less than or equal to the reference value I.sub.k. If the 
answer to the inquiry at the step 418 is YES, i.e., I.sub.f is less than 
or equal to I.sub.k, the sum of I.sub.f and I.sub.b is calculated at a 
step 420. At a step 422, a number of steps along which a stepper motor 
drive signal should be moved is determined and the result is stored as a 
step number M. Then, at a step 424, the stepper motor drive signal is 
moved along the number of steps M. In this example, the reference value 
I.sub.k is set equal to a reduction ratio value less than and close to the 
maximum reduction ratio provided by the continuously variable transmission 
mechanism. 
If the gearing mechanism still takes over a drive, or the transition in 
drive from the gearing mechanism to the continuously variable transmission 
mechanism has not been completed, the feedforward reduction ratio I.sub.f 
is still greater than the reference value I.sub.k and, thus, the answer to 
the inquiry at the step 418 is NO. In this case, the program proceeds from 
this step 418 to a step 426, wherein the feedback reduction ratio I.sub.b 
is set equal to zero and then to the step 420, 422 and 424. In this case, 
the feedback control is prohibited. 
From the above explanation along with this flowchart, it will be 
appreciated that feedback control is prohibited if the gearing mechanism 
still takes over a drive or the transition in drive from the gearing 
mechanism to the continuously variable transmission mechanism has not been 
completed yet. 
A time period required from the instant when a command is issued to the 
instant when a transition in drive from the gearing mechanism to the 
continuously variable transmission mechanism is completed is not 
invariable but variable as a function of throttle opening degree TH. Thus, 
it is preferrable to set the reference value I.sub.k equal to a function 
I.sub.k (TH) as shown in FIG. 5A. Referring to FIG. 5A, a shadow region 
illustrates the region where feedback control is inhibited. It will now be 
appreciated that the region where the feedback control be prohibited 
extends toward the minimum reduction ratio as the throttle opening degree 
TH increases. 
Referring to FIG. 6, a control concept according to the present invention 
is described. For this control, the outputs of the throttle sensor 303 and 
vehicle speed sensor 302 are supplied to a functional block 500 where the 
target driver pulley revolution speed N.sub.t for a throttle opening 
degree TH indicated by the output of the throttle sensor 303 and a vehicle 
speed V indicated by the output of the vehicle speed sensor 302 is 
determined. The output N.sub.t of this block 500 is supplied to a block 
502 where a feedforward reduction ratio I.sub.f is determined by 
calculating an equation I.sub.f =N.sub.t /V. The output I.sub.f is 
supplied to a block 504 where the feedforward reduction ratio I.sub.f is 
added to a feedback reduction ratio I.sub.b. The sum of I.sub.f and 
I.sub.b is supplied to a block 506 where a stepper motor pulse number M is 
determined. The output of this block 506 is supplied to the shift motor 
110. 
The output N.sub.t of the block 500 and the output of the turbine 
revolution speed sensor 305 are supplied to a block 508. The output of the 
turbine revolution speed sensor 305 is indicative of a driver pulley 
revolution speed N.sub.a. At the block 508, a deviation e between N.sub.t 
and N.sub.a is calculated. The output e of the block 508 is supplied to a 
block 510 where the integral S, with respect to time, of the deviation e 
is calculated. The output S of the block 510 is supplied to a block 512 
where a feedback reduction ratio I.sub.b is calculated as a function of 
the integral S. The output I.sub.b of the block 512 is supplied via a gate 
514 to the before mentioned block 504. The feedforward reduction ratio 
I.sub.f is supplied from the block 502 to a block, where I.sub.f is 
compared with a reference value I.sub.k supplied from a block 518. At the 
block 518, I.sub.k is set equal to a constant or a function I.sub.k (TH) 
(see FIG. 5A). At the block 516, it is determined whether I.sub.f is less 
than or equal to the reference value I.sub.k. If the feedforward reduction 
ratio I.sub.f is still larger than the reference value I.sub.k and thus 
the transition from the drive owing to the gearing mechanism to the drive 
owing to the continuously variable transmission mechanism has not been 
completed yet, the gate 514 shifts in response to the output of the block 
516 to a state where the feedback reduction ratio I.sub.b is set equal to 
zero. Thus, the feedback control is inhibited. Of course, if the 
feedforward reduction ratio I.sub.f becomes equal to or less than the 
reference value I.sub.k, the gate 514 allows the passage of feedback 
reduction ratio I.sub.b to the block 504, thus allowing the feedback 
control to take effect. 
Referring to the flowchart shown in FIG. 7, another embodiment is 
described. This embodiment is substantially the same as the previously 
described embodiment except the fact the region where the feedback control 
is inhibited is defined by from comparing an actual reduction ratio 
I.sub.a with a reference value I.sub.c that corresponds to the maximum 
reduction ratio provided by the continuously variable transmission 
mechanism, instead of comparing the feedback reduction ratio I.sub.f with 
the reference value I.sub.k. 
Specifically, in FIG. 7, the program proceeds from a step 414 to a step 
416A where an actual reduction ratio I.sub.a is calcualted using an 
equation I.sub.a =N.sub.a /V. Then, at a step 418A, it is determined 
whether the actual reduction ratio I.sub.a is less than or equal to a 
reference value I.sub.c corresponding to the maximum reduction ratio 
provided by the continuously variable transmission. If an answer to the 
inquiry is NO, the program proceeds to a step 426, while if the answer is 
YES, the program proceeds directly to a step 420. 
FIG. 8 is a block diagram similar to FIG. 6 and illustrate a control 
concept of this embodiment. This block diagram is substantially the same 
as FIG. 6, except for the addition of block 520 and addition of blocks 
518A and 516A in the place of the blocks 518 and 516. 
According to this embodiment, the feedback control is inhibited until the 
transition in drive from the gearing mechanism to the continuously 
variable transmission mechanism is completed. 
The present invention is applicable to a hybrid transmission including a 
gearing mechanism which is combined with a continuously variable 
transmission mechanism, such that the gearing mechanism takes over a drive 
to provide a reduction ratio smaller than the minimum reduction ratio 
provided by the continuously variable transmission mechanism. One example 
of such a hybrid transmission is illustrated in FIGS. 10 through 13. 
The hybrid transmission shown in FIG. 10 is substantially the same as the 
previously described hybrid transmission shown in FIG. 1, except fact fact 
that the sizes of gears 58A, 50A, 42A, 48A, 38A, and 54A are different 
from their counterparts 58, 50, 42, 48, 38, and 54. 
FIG. 11 illustrates in the fully drawn line that the pulley unit 16, 24, 
and 26 and the gears 58A, 50A, 62, and 64 play a role when the 
continuously variable transmission takes over a forward drive in the 
hybrid transmission shown in FIG. 10. 
FIG. 12 illustrates in the fully drawn line that the gears 42A, 48A, 62, 
and 64 play a role when the gearing mechanism takes over a forward drive 
in the hybrid transmission shown in FIG. 10. When it takes over the drive, 
the gearing mechanism provides a reduction ratio smaller than the minimum 
or smallest reduction ratio provided by the continuously variable 
transmission. 
FIG. 13 illustrates in the fully drawn line that the gears 38A, 54A, 50A, 
62, and 64 play a role when the gearing mechanism takes over a reverse 
drive in the hybrid transmission shown in FIG. 10. 
The present invention is embodied in this hybrid transmission such that the 
feedback control of the continuously variable transmission mechanism is 
prevented until a transition from the forward drive owing to the gearing 
mechanism to the drive owing to the continuously variable transmission 
mechanism is completed.