Control system for a continuously variable transmission

A control system for a toroidal type continuously variable transmission which incorporates a toroidal continuously variable gear mechanism comprising an input and an output toroidal disk and a roller disposed between the input and output toroidal disks and continuously varies a gear ratio according to inclinations of the roller actuates a stepping motor, when a specific driving condition which causes the continuously variable transmission to switch between a first and a second control mode and directional reversal of torque transmission through the toroidal continuously variable gear mechanism occurs with a result of causing a shift of the roller from one side to another side of the neutral position, to operate a double-slider shift valve to cause the shift of the roller in advance.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The invention relates to a control system for a continuously variable 
transmission installed to a vehicle, and more particularly, to a control 
system for a continuously variable transmission of a type having a 
toroidal continuously variable mechanism. 
2. Description of the Related Art 
There have been known various continuously variable transmissions such as 
incorporating a toroidal continuously variable mechanism installed to an 
automobile vehicle. Such a toroidal type continuously variable 
transmission, which comprises an input and output toroidal disk and a 
roller disposed between the input and output toroidal disks, continuously 
varies a gear ratio according to inclinations of the roller relative to 
the toroidal disks. The roller is supported for rotation by a support 
member which is called a trunnion. In order to incline the roller the 
trunnion is shifted in a direction tangent to the toroidal surface of the 
toroidal disk by a hydraulically operated actuator. A shift of the 
trunnion, and hence an inclination of the roller, is controlled by 
hydraulic working pressure supplied to the hydraulically operated actuator 
through a shift valve of a double-slider type which has a hollow valve 
sleeve axially slidable in a valve body and a valve spool axially slidable 
in the valve sleeve. 
The valve sleeve is shifted and changed in position relative to the valve 
spool by a stepping motor to regulate hydraulic working pressure supplied 
to the actuator so as to shift the roller through the trunnion in a 
direction tangent to the toroidal surface from a neutral position. As a 
result of the shift, the roller inclines relative to the toroidal disk to 
change a gear ratio between an input speed to the toroidal continuously 
variable mechanism and output speed from the toroidal continuously 
variable mechanism. The inclination of the roller is mechanically fed back 
to the shift valve to shift the valve spool in the same direction in which 
the valve sleeve has been shifted, as a result of which the shift valve 
returns to a balanced or neutral position. During the return of the shift 
valve to the neutral position, while the roller is left inclined, it is 
shift back to the neutral position and held in the neutral position. 
As described in, for example, Japanese Unexamined Patent Publications Nos. 
3-223555 and 6-101754, it has been known that what is called a geared 
neutral starting system is employed in the continuously variable 
transmission. In the toroidal type continuously variable transmission 
equipped with the geared neutral starting system, the toroidal 
continuously variable mechanism is mounted on a transmission input shaft 
connected to the engine, and a planetary gear mechanism is mounted on a 
secondary shaft in parallel to the input shaft. The planetary gear 
mechanism is comprised of three rotary elements, namely a sun gear, an 
internal gear and a pinion carrier supporting a pinion carrier meshed with 
the sun gear and the internal gear. One of these rotary elements, i.e. the 
internal gear is used as an transmission output gear. Rotation of an 
engine is imparted to the planetary gear mechanism partly directly through 
the pinion carrier and partly through the sun gear via the toroidal 
continuously variable mechanism. 
The ratio of rotation between the pinion carrier and the sun gear is varied 
by controlling the gear ratio of the toroidal continuously variable 
mechanism so as to hold the transmission output element, i e. the sin 
gear, remain stand still, providing a neutral condition. By increasingly 
or decreasingly varying the gear ratio of the toroidal continuously 
variable mechanism causes the internal gear as the transmission output 
element to rotate in a forward direction or in a reverse direction. 
Further, some toroidal type continuously variable transmissions equipped 
with the geared neutral starting system are switchable between what is 
called a high gear ratio control mode in which output rotation of the 
toroidal continuously variable mechanism is transmitted directly to the 
secondary shaft without the aid of the planetary gear mechanism and what 
is called a low gear ratio control mode in which rotation of an engine is 
transmitted directly to the pinion carrier and transmitted to the sun gear 
through the toroidal continuously variable mechanism. In such a toroidal 
type continuously variable transmission, in order to cause a switch 
between the high and low modes, the toroidal type continuously variable 
transmission is provided with a low mode clutch to connect and disconnect 
the input shaft and the pinion carrier of the planetary gear mechanism on 
the secondary shaft and a high mode clutch to connect and disconnect the 
toroidal continuously variable mechanism on the input shaft and the 
secondary shaft. These mode clutches are alternately locked and unlocked 
in a state where a gear ratio of the continuously variable transmission 
remains identical in both high and low modes. 
As was previously described, after a shift in a direction tangent to the 
toroidal surface of the toroidal disks and inclination, the roller of the 
toroidal continuously variable mechanism is returned to its neutral 
position, i.e. a position of a plane perpendicular to the tangential 
direction when the plane passes a center axis of rotation of the toroidal 
disks, and is balanced in that position. However, the roller is 
practically balanced in a position slightly shifted on a specific side of 
the neutral position in the tangential direction. 
Specifically, as shown in FIG. 24, when torque is transmitted, for example, 
from an input toroidal disk A to an output toroidal disk B through a 
roller D supported by a trunnion C (this direction of torque transfer is 
referred to as a normal direction) while the input and output toroidal 
disks A and B and the roller D are rotating in directions indicated by 
arrows A', B' and C', respectively, drag force from the input or drive 
toroidal disk A or reaction force of the output or driven toroidal disk B 
acts on the trunnion C as traction force T directed downward as viewed in 
the figure. Hydraulic working pressure is supplied to a roller actuator E 
to hold the trunnion C or the roller D in the neutral position against the 
traction force T. In this instance, the roller D is practically balanced 
in a position shifted according to input torque in a direction opposite to 
the traction force T (upward as viewed in the figure), i.e. in a direction 
in which the toroidal continuously variable mechanism changes its gear 
ratio toward the higher speed side (which is referred to as a plus side), 
from the neutral position. The shift direction depends upon the transfer 
direction of torque through the toroidal continuously variable mechanism. 
For example, when torque is transmitted in a direction to the input 
toroidal disk A from the output toroidal disk B (which is referred to as a 
reverse direction) while the input and output toroidal disks A and B and 
the roller D are rotating in directions indicated by arrows A', B' and C', 
respectively, the roller D is balanced in a position shifted downward as 
viewed in the figure, i.e. in a position shifted in a direction in which 
the toroidal continuously variable mechanism changes its gear ratio toward 
the lower speed side (which is referred to as a minus direction), from the 
neutral position. Accordingly, the roller D shifts instantaneously from 
one side to another side of the neutral position. The directional reversal 
of torque transfer occurs in response to a switch from a normal drive 
state in which the vehicle is driven by the engine to a reverse drive 
state in which the engine is driven by the vehicle running with inertia or 
vice versa. 
In the toroidal type continuously variable transmissions switchable between 
the high and low gear ratio control modes, torque is transmitted from the 
output toroidal disk to the input toroidal disk in the low mode while the 
vehicle is in the normal drive state and from the input toroidal disk to 
the output toroidal disk in the high mode while the vehicle is in the 
normal drive state, and, on the other hand, from the input toroidal disk 
to the output toroidal disk in the low mode while the vehicle is in the 
reverse drive state and from the output toroidal disk to the input 
toroidal disk in the high mode while the vehicle is in the reverse drive 
state. The directional reversal of torque transfer occurs also in response 
to a switch between the low and high gear ratio control modes. An 
instantaneous shift, and hence an instantaneous inclination, of the roller 
is accompanied by an occurrence of shocks. 
SUMMARY OF THE INVENTION 
It is an objective of the invention to provide a control system for a 
toroidal type of continuously variable transmission in which a switch of 
drive state of the vehicle and a switch of control mode of the 
transmission are achieved without being accompanied by shocks. 
The foregoing object of the present invention is achieved by providing a 
control system for a toroidal type continuously variable transmission 
which incorporates a toroidal continuously variable gear mechanism which 
comprises an input toroidal disk, an output toroidal disk, a roller 
disposed between the toroidal input and output disks and a support member 
such as a trunnion for supporting the roller for rotation, and further 
incorporates a hydraulically operated roller actuator for shifting and 
inclining the roller relative to the input and output toroidal disks from 
a neutral position through the support member according to hydraulic 
pressure to continuously vary a gear ratio between an input speed to the 
input toroidal disk from a driving power source, such as a vehicle engine, 
and an output speed to a driven element, such as drive wheels of the 
vehicle, from the output toroidal disk and a shift valve for controlling 
hydraulic pressure supplied to the roller actuator according to vehicle 
running conditions to adjust an inclination of the roller and stopping 
supply of the hydraulic pressure to the roller actuator by means of a 
mechanical feedback of the inclination of the roller thereto to return the 
roller back to the neutral position. The continuously variable 
transmission control system controls the shift valve to control the 
hydraulic pressure such that, when a predetermined driving condition of 
the driving power source, such as a predetermined engine speed, which 
causes directional reversal of torque transmission through the toroidal 
continuously variable gear mechanism with a result of a shift of the 
roller from one side to another side of the neutral position, the roller 
actuator causes the shift of the roller in advance. The shift valve may be 
of a double-slider type which comprises a valve body, a hollow valve 
sleeve received to slide axially in the valve body and a valve spool 
received to slide axially in the hollow valve sleeve. The hollow valve 
sleeve and the valve spool are relatively shifted to control and supply 
hydraulic pressure to the roller actuator to shift and incline the roller 
according to a shifted relative position of them and the valve spool is 
shifted by means of a mechanical feedback of the shift and inclination of 
the roller thereto to stop supply of the hydraulic pressure to the roller 
actuator. 
The toroidal type of continuously variable transmission may further 
incorporate a planetary gear mechanism which comprises three rotary 
elements, namely a pinion carrier connected to an input member through 
which driving power is transmitted to the toroidal continuously variable 
gear mechanism from the driving power source, a sun gear connected to the 
output toroidal disk and an internal gear connected to the driven element. 
The continuously variable transmission control system switches the 
continuously variable transmission between a first gear ratio control mode 
in which driving power is transmitted through a first power transmission 
path including both toroidal continuously variable gear mechanism and 
planetary gear mechanism and a second gear ratio control mode in which 
driving power is transmitted through a second power transmission path 
including the toroidal continuously variable gear mechanism but omitting 
the planetary gear mechanism according to driving conditions and controls 
the shift valve to control hydraulic pressure such that, when the 
predetermined driving condition with which directional reversal of torque 
transmission through the toroidal continuously variable gear mechanism is 
caused with a result of a shift of the roller from one side to another 
side of the neutral position is detected, the roller actuator causes the 
shift of the roller in advance. 
The continuously variable transmission control system may force the shift 
valve to control hydraulic pressure so as to cause a shift of the roller 
in a direction in which the toroidal continuously variable gear mechanism 
causes a change in gear ratio toward a higher speed side (a lower gear 
ratio side) when there occurs directional reversal of torque transmission 
through the toroidal continuously variable gear mechanism due to a switch 
of the continuously variable transmission from the first gear ratio 
control mode to the second gear ratio control mode while the vehicle is in 
a normal drive state in which the engine drives the vehicle or when there 
occurs directional reversal of torque transmission through the toroidal 
continuously variable gear mechanism due to a switch of the continuously 
variable transmission from the second gear ratio control mode to the first 
gear ratio control mode while the vehicle is in a reverse drive state in 
which the engine is driven by inertial running of the vehicle, and to 
control hydraulic pressure so as to cause a shift of the roller in a 
direction in which the toroidal continuously variable gear mechanism 
causes a change in gear ratio toward a lower speed side (higher gear ratio 
side) when there occurs directional reversal of torque transmission 
through the toroidal continuously variable gear mechanism due to a switch 
of the continuously variable transmission from the second gear ratio 
control mode to the first gear ratio control mode while the vehicle is in 
the normal drive state or when there occurs directional reversal of torque 
transmission through the toroidal continuously variable gear mechanism due 
to a switch of the continuously variable transmission from the first gear 
ratio control mode to the second gear ratio control mode while the vehicle 
is in the reverse drive state. The hydraulic pressure the shift valve 
provides is determined based on input torque to the toroidal continuously 
variable gear mechanism at a point of time that the directional reversal 
of torque transmission occurs. Specifically, the hydraulic pressure is 
determined to become greater as the input torque to the toroidal 
continuously variable gear mechanism rises. 
The continuously variable transmission control system may further force the 
shift valve to control hydraulic pressure so as to cause a shift of the 
roller in a direction in which the toroidal continuously variable gear 
mechanism causes a change in gear ratio toward a higher speed side when 
there occurs a switch from the normal drive state to the reverse drive 
mode while the continuously variable transmission is in the first control 
mode or when there occurs a switch from the reverse drive state to the 
normal drive state while the continuously variable transmission is in the 
second control mode, and controls the shift valve to control hydraulic 
pressure so as to cause a shift of the roller in a direction in which the 
toroidal continuously variable gear mechanism causes a change in gear 
ratio toward a lower speed side when there occurs a switch from the 
reverse drive state to the normal drive mode while the continuously 
variable transmission is in the first control mode or when there occurs a 
switch from the normal drive state to the reverse drive state while the 
continuously variable transmission is in the second control mode. 
Switching of the continuously variable transmission between the first and 
second gear ratio control modes may be performed by means of first and 
second friction coupling means such as clutches. Specifically, the first 
friction coupling means or clutch is locked to bring the continuously 
variable transmission into the first gear ratio control mode by 
incorporating both toroidal continuously variable gear mechanism and 
planetary gear mechanism in the first power transmission path, and the 
second friction coupling means or clutch is locked to bring the 
continuously variable transmission into the second gear ratio control mode 
by incorporating the toroidal continuously variable gear mechanism only in 
the second power transmission path. The continuously variable transmission 
control system causes the first and second friction coupling means or 
clutches to switch the continuously variable transmission between the 
first and second control modes according to the driving conditions and 
controls the shift valve to control the hydraulic pressure such that, when 
a predetermined driving condition with which a switch of the continuously 
variable transmission between the first and second gear ratio control 
modes occurs is caused and directional reversal of torque transmission 
through the toroidal continuously variable gear mechanism occurs with a 
result of a shift of the roller from one side to another side of the 
neutral position occurs is detected, the roller actuator causes the shift 
of the roller in advance while the first and second friction coupling 
means or clutches remain locked. 
According to the continuously variable transmission control system of the 
invention, when directional reversal of torque transmission through the 
toroidal continuously variable gear mechanism occurs with a result of a 
shift of the roller from one side to another side of the neutral position, 
the shift valve is controlled to provide hydraulic pressure to force the 
roller actuator to cause the shift of the roller in advance. As a result, 
the roller is prevented from being abruptly shifted and inclined, so that 
the continuously variable transmission prevents an occurrence of shift 
shock. The same is true for the continuously variable transmission which 
is switchable between two gear ratio control modes and between two drive 
states, namely the normal drive state and the revere drive state. either 
or both. 
An advance shift distance of the roller is determined based on input torque 
to the continuously variable transmission at the point of time the 
directional reversal of torque transmission occurs, which cause a precise 
shift of the roller timely with an effect of preventing shift shock more 
effectively. When continuously variable transmission is switched between 
the high and low gear ratio control modes by alternately locking and 
unlocking two friction coupling means, namely two clutches, a shift of the 
roller on directional reversal of torque transmission through the toroidal 
continuously variable gear mechanism is caused while both friction 
coupling means remain locked, so that the continuously variable 
transmission holds a gear ratio unchanged during the shift of the roller. 
This is advantageous to prevent shift shock.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring to the drawings in detail and, in particular, to FIG. 1 showing a 
toroidal type continuously variable transmission (CVT) 10 with a control 
system in accordance with an embodiment of the invention, the continuously 
variable transmission 10 has three shafts, namely an input shaft 11 
connected to an output shaft 2 of an engine 1 through a tortional damper 
3, a hollow primary shaft 12 in which the output shaft 11 is coaxially 
disposed and a secondary shaft 13 disposed in parallel to the primary 
shaft 12. These transmission shafts 11-13 extend transversely in the 
vehicle body. The primary shaft 12 mounts thereon a toroidal type first or 
rear and second or front continuously variable gear mechanisms 20 and 30 
in order from a side remote from the engine 1 and a loading cam 40. The 
secondary shaft 13 mounts thereon a planetary gear mechanism 50 and two 
clutches, namely a low mode clutch 60 and a high mode clutch 70. There are 
low and high mode gear trains 80 and 90 between input shaft 11 and the 
secondary shaft 13. The toroidal type first and second toroidal 
continuously variable gear mechanisms 20 and 30 have the same mechanism. 
Each toroidal continuously variable gear mechanism 20, 30 comprises an 
input disk 21, 31, an output disk 22, 32 and a pair of first and second 
roller 23, 33 between the input and output surfaces to transmit driving 
power from one to another. The input disks 21 and 31 are mounted on the 
primary shaft 12 by means of spline-engagement. The output disks 22 and 32 
are formed integrally as one whole, which is hereafter referred to as an 
integrated output disk 34. 
First toroidal continuously variable gear mechanism 20 is placed to direct 
the face of the input disk 21 toward the engine 1, and, however, the 
second toroidal continuously variable gear mechanism 30 is placed to 
direct the face of the output disk 32 toward the engine 1. The input disks 
21 and 31 of the first and second toroidal continuously variable gear 
mechanisms 20 and 30 are fixedly mounted to opposite ends of the primary 
shaft 12, respectively, and the output disks 22 and 32 of the first and 
second toroidal continuously variable gear mechanisms 20 and 30 are 
mounted for rotation on the primary shaft 12. 
The input shaft 11 is provided with a low mode first gear 81 forming a part 
of the low mode gear train 80 secured to one end thereof remote from the 
engine 1 and the loading cam 40 between the low mode first gear 81 and the 
first toroidal continuously variable gear mechanism 20. The integrated 
output disk 34 is provided with an peripheral first high mode gear 91 
forming a part of the high mode gear train 90. The secondary shaft 13 
mounts thereon for rotation a low mode second gear 82 forming a part of 
the low mode gear train 80 and a high mode second gear 92 for forming a 
part of the high mode gear train 90. These first and second low mode gears 
81 and 82 of the low mode gear train 80 are connected through a high mode 
idle gear 83, and similarly, these high mode first and second gears 91 and 
92 of the high mode gear train 90 are connected through a high mode idle 
gear 93. The secondary shaft 13 mounts thereon the planetary gear 
mechanism 50 having a sun gear 52 meshed with the high mode second gear 92 
and an internal gear 53 secured to the secondary shaft 13. The high mode 
clutch 70 is placed adjacent to the high mode second gear 92 to connect 
the high mode second gear 92 to the secondary shaft 13 and disconnect the 
high mode second gear 92 from the secondary shaft 13. The secondary shaft 
13 at one of its end is connected to a differential 5 by means of an 
output gear train 4 comprising first and second gears 4a and 4b. Driving 
power is transmitted to right and left wheel drive axles 6a and 6b through 
the differential 5. 
As was previously described, because the toroidal continuously variable 
gear mechanisms 20 and 30 have the same mechanism, the following 
description is directed to the toroidal continuously variable gear 
mechanism 20 only. 
Referring to FIGS. 2 and 3 in detail, each roller 22, 23 is supported by a 
trunnion 25 through a shaft 24 radially extending. The roller 22, 23 at 
its both ends is in contact with the opposite toroidal surfaces of the 
input and output disks 21 and 22. The first and second rollers 22 and 23 
are placed in the same plane passing the center axis of rotation of the 
input shaft 12 and on opposite sides of the center axis of rotation of the 
input shaft 12. The trunnion 25 is held at axially opposite sides by 
supports 26 so as to rotate about a horizontal axis X tangential to the 
toroidal surface and perpendicular to the shafts 24 and to move back and 
forth along the axis X. The trunnion 25 is provided with a trunnion rod 27 
extending in a direction of the axis X. The roller 23 is rolled by means 
of a speed change control unit 110 secured to the transmission housing 100 
through the trunnion 25 and trunnion rod 27. 
The speed change control unit 110 includes a hydraulic control section 111 
and a trunnion drive section 112. The trunnion drive section 112 includes 
a piston 1131 and a piston 1141 both of which are attached to the rod 27 
for the first roller 231 and a piston 1132 and a piston 1142 both of which 
are attached to the rod 27 for the s second roller 232. Hydraulic pressure 
chambers 1151 and 1161 are formed facing to the first piston 1131 and the 
first piston 1141, respectively, and similarly, hydraulic pressure 
chambers 1152 and 1162 are formed facing to the second piston 1132 and the 
second piston 1142, respectively. The hydraulic pressure chamber 1151 for 
the first piston 1131 is placed close to the first roller 231 and the 
hydraulic pressure chamber 1161 for the first piston 1131 is placed far 
from the first roller 231. On the other hand, the hydraulic pressure 
chamber 1162 for the second piston 1132 is placed close to the second 
roller 232 and the hydraulic pressure chamber 1152 for the second piston 
1132 is placed far from the first roller 231. Hydraulic pressure provided 
by the hydraulic control section 111 is delivered into the hydraulic 
pressure chambers 1151 and 1152 through oil paths 117 and 118, 
respectively, for and into the hydraulic pressure chambers 1151 and 1162 
through oil paths (not shown). 
In hydraulic pressure delivery control for the toroidal continuously 
variable gear mechanism 20 by way of example, when the hydraulic pressure 
PH in the first and second hydraulic pressure chambers 1151 and 1152 
becomes higher than a neutral level and relatively to the hydraulic 
pressure PL in the first and second reduction hydraulic pressure chambers 
1161 and 1162, the first trunnion 251 is forced to move horizontally 
toward the right as viewed in FIG. 3 and the second trunnion 252 is forced 
toward the left. If the output disk 22 is rotating in a clockwise 
direction as viewed in FIG. 3, the first roller 231 receives downward 
force from the output disk 22 and upward force from the input disk 21 
rotating in a counterclockwise direction during moving toward the right, 
and conversely the second roller 232 receives upward force from the output 
disk 22 and downward force from the input disk 21 during moving toward the 
left. As a result, both rollers 231 and 232 incline so as to shift their 
contact points with toroidal surfaces of the input disk 21 radially 
outward and their contact points with toroidal surfaces of the output disk 
22 radially inward, thereby lowering a gear ratio of the toroidal 
continuously variable gear mechanism 20. On the other hand, when the 
hydraulic pressure PL in the first and second reduction hydraulic pressure 
chambers 1161 and 1162 becomes higher than a neutral level and relatively 
to the hydraulic pressure PH in the first and second speed increase 
pressure chambers 1151 and 1152, the first trunnion 251 is forced to move 
horizontally toward the left as viewed in FIG. 3 and the second trunnion 
252 is forced toward the right. At this time, when the output disk 22 
rotates in the direction X, the first roller 231 receives upward force 
from the output disk 22 and downward force from the input disk 21 during 
moving toward the left, and the second roller 232 receives downward force 
from the output disk 22 and upward force from the input disk 21 during 
moving toward the left. As a result, both rollers 231 and 232 incline so 
as to shift their contact points with toroidal surfaces of the input disk 
21 radially inward and their contact points with toroidal surfaces of the 
output disk 22 radially outward, thereby changing the toroidal 
continuously variable gear mechanism 20 to a higher gear ratio. 
Controlling of the hydraulic pressure PH or PL will be described in detail 
later with reference to a hydraulic control circuit 200 shown in FIG. 8. 
Operation of the toroidal continuously variable gear mechanism 30 occurs in 
the same way as described above regarding the toroidal continuously 
variable gear mechanism 20. 
Because the splined input disks 21 and 22 are placed on opposite splined 
ends of the primary shaft 12, respectively, and the output disks 22 and 32 
are integrated, the first and second toroidal continuously variable gear 
mechanisms 20 and 30 always have same input speed and output speed, and 
provide always the same gear ratio. 
FIG. 4 shows the loading cam 40 in detail. The loading cam 40 has a cam 
disk 41 disposed between the low mode first gear 81 of the low mode gear 
train 80 and the input disk 21 of the first toroidal continuously variable 
gear mechanism 20. The cam disk 41 has a face cam with alternate 
arrangement of crests and roots 21a formed thereon. The input disk 21 at 
its back has a face cam formed thereon correspondingly to the face cam of 
the cam disk 41. A plurality of rollers 43 held by a retainer disk 42 are 
interposed between the face cams. The cam disk 41 is mechanically coupled 
to the low mode first gear 81 by means of a plurality of coupling pins 44. 
As shown in FIG. 5, there are arranged conical disk springs 45, a needle 
bearing 46 and a bearing race 47 between the cam disk 41 and primary shaft 
flange 12a. The cam disk 41 is forced against the input disk 21 by the 
conical disk springs 45. By means of the structure, the rollers 43 are 
interposed between the roots 21a and 41a of the face cams to transmit 
drive torque input to the cam disk 41 through the input shaft 11 via the 
low mode first gear 81 to the input disk 21 of the first toroidal 
continuously variable gear mechanism 20 and further to the input disk 31 
of the second toroidal continuously variable gear mechanism 30 through the 
primary shaft 12. 
As shown in FIG. 5, the transmission rear end cover 101 is provided with an 
oil pump 102 which is driven by means of the low mode first gear 81. 
Referring to FIG. 6 showing the planetary gear mechanism 50 and low and 
high mode clutches 60 and 70, the secondary shaft 13 at its opposite ends 
is supported for rotation by transmission end covers 101 and 103 through 
bearings 141 and 142, respectively. The secondary shaft 13 at its middle 
portion mounts thereon the high mode second gear 92 and the planetary gear 
mechanism 50 adjacent to the high mode second gear 92 on one side remote 
from the engine 1. The sun gear 52 of the planetary gear mechanism 50 is 
meshed with the high mode second gear 92. A splined sleeve flange 54 with 
an external gear is placed on the splined end of the secondary shaft 13 
behind the planetary gear mechanism 50 and is meshed with the internal 
gear 53 of the planetary gear mechanism 50. Further, the low mode clutch 
60 is mounted for rotation on the secondary shaft 13 behind the sleeve 
flange 54. This low mode clutch 60 comprises an internally splined clutch 
drum 61 to which the low mode second gear 82 is secured, an externally 
splined clutch hub 62 disposed radially inside the clutch drum 61 and 
connected to a flange 55 by means of an externally splined pinion carrier 
51, a plurality of splined clutch plates 63 alternately coupled to both 
clutch drum 61 and clutch hub 62, and a spring loaded piston 64 installed 
within the clutch drum 61. The clutch drum 61 defines a hydraulic chamber 
65 therein behind the piston 64. The piston 64 is forced axially toward 
the planetary gear mechanism 50 against a spring 66 by coupling hydraulic 
pressure supplied into the hydraulic pressure chamber 65 by a clutch 
control unit 120 (see FIG. 3), so as to frictionally couple the clutch 
plates 63 all together, thereby bringing the low mode second gear 82 and 
the pinion carrier 51 into engagement with each other. The piston 64 is 
provided with a balancing piston 67 secured to its front wall to provide a 
balancing hydraulic chamber 68 therebetween. Lubrication oil in the 
balancing hydraulic pressure chamber 68 cancels thrust force unevenly 
acting on the piston 64 by means of centrifugal force acting on the oil in 
the hydraulic pressure chamber 65. Adjacent to the high mode second gear 
92 there is a high mode clutch 70. The high mode clutch 70 comprises an 
internally splined clutch drum 71 which is coupled to a first gear 4a of 
the output gear train 4 placed on the splined secondary shaft 13 through a 
parking gear 4d, an externally splined clutch hub 72 disposed radially 
inside the clutch drum 71 and connected to the high mode second gear 92, a 
plurality of splined clutch plates 73 alternately coupled to both clutch 
drum 71 and clutch hub 72, and a spring loaded piston 72 installed within 
the clutch drum 71. The clutch drum 71 defines a hydraulic pressure 
chamber 75 therein behind the piston 74. The piston 74 is forced axially 
toward the planetary gear mechanism 50 against a spring 76 by coupling 
hydraulic pressure applied into the hydraulic pressure chamber 75 by the 
clutch control unit 120, so as to frictionally couple the clutch plates 73 
all together, thereby bringing the high mode second gear 92 and the first 
gear 4a of the output gear train 4 placed on the splined secondary shaft 
13. The piston 74 is provided with a balancing piston 77 secured to its 
back wall to provide a balancing hydraulic pressure chamber 78 
therebetween. Lubrication oil in the balancing hydraulic pressure chamber 
78 cancels thrust force unevenly acting on the piston 74 by means of 
centrifugal force acting on the oil in the hydraulic pressure chamber 75. 
The transmission end cover 101 is provided with axial oil paths 131 and 
133. Hydraulic oil from the clutch control unit 120 is supplied to the 
hydraulic pressure chamber 65 of the low mode clutch 60 through the axial 
oil path 131 and to the hydraulic pressure chamber 75 of the high mode 
clutch 70 through the axial oil path 133 via an axial oil path 132 formed 
in the secondary shaft 13. 
[0078] In the mechanical operation of the continuously variable 
transmission 10, while the vehicle is stopping, in the low mode control 
where the low mode clutch 60 is locked and the high mode clutch 70 is 
released, rotation of the engine 1 is transmitted to the secondary shaft 
12 from the input shaft 11 through the low mode gear train 80 comprising 
the first gear 81, the idle gear 83 and the second gear 82 and 
simultaneously transmitted to the planetary gear mechanism 50 through the 
pinion carrier 51 via the low mode clutch 60. The rotation imparted to the 
input shaft 11 is transmitted to the input disk 21 of the first toroidal 
continuously variable gear mechanism 20 from the low mode first gear 81 
through the loading cam 40 and further transmitted to the integrated 
output disk 34 through the rollers 23. Simultaneously, the rotation is 
imparted to the input disk 31 of the second toroidal continuously variable 
gear mechanism 30 from the input disk 21 of the first toroidal 
continuously variable gear mechanism 20 through the primary shaft 12 and 
further transmitted to the integrated output disk 34 through the rollers 
23. At this time, the speed change control unit 110 controls the hydraulic 
pressure PH for speed increase or the hydraulic pressure PL for speed 
reduction to hold the rollers 23 of the first and second toroidal 
continuously variable gear mechanisms 20 and 30 at an inclination angle 
for a specified gear ratio. The rotation imparted to the integrated output 
disk 34 is further transmitted to the sun gear 52 of the planetary gear 
mechanism 50 through the high mode gear train 90 comprising the first and 
second high mode gears 91 and 92. At this time, the speed change control 
unit 110 controls the hydraulic pressure PH for speed increase or the 
hydraulic pressure PL for speed reduction to hold the rollers 23 of the 
first and second toroidal continuously variable gear mechanisms 20 and 30 
at an inclination angle for a given gear ratio. In this way, the planetary 
gear mechanism 50 receives the rotation through both pinion carrier 51 and 
sun gear 52. The carrier 51 and the sun gear 52 rotate at a same speed due 
to the control of gear ratio of the first and second toroidal continuously 
variable gear mechanisms 20 and 30, not causing rotation of the internal 
gear 53 of the planetary gear mechanism 50, i e. rotation transmitted to 
the differential 5 from the secondary shaft 12 through the output gear 
ratio 4, at all. As a result, the continuously variable transmission 10 
remains put in a geared neutral state. When varying the gear ratios of the 
first and second toroidal continuously variable gear mechanisms 20 and 30 
to cause a change in speed ratio between rotation imparted to the pinion 
carrier 51 and the sun gear 52, respectively, the internal gear 13 rotates 
in either direction to cause the vehicle to start forward or backward in 
the low mode control in which the continuously variable transmission 10 is 
at a high resultant gear ratio. 
In the low mode gear ratio control, as schematically shown in FIG. 7, while 
the engine torque is transmitted to the secondary shaft 13 through the low 
mode gear train 80 via one end of the input shaft 11 remote from the 
engine 1 as indicated by an arrow a, torque is circulated as reaction 
force caused in the planetary gear mechanism 50 to the integrated output 
disk 34 of the first and second toroidal continuously variable gear 
mechanisms 20 and 30 via the high mode gear train 90 as indicated by an 
arrow b. This recirculating torque is returned again to the low mode gear 
train 80 via the input toroidal disks 21 and 31, the primary shaft 12 and 
the loading cam 40 as indicated by an arrow b. Accordingly, in the low 
mode gear ratio control, torque is transmitted in reverse direction to the 
input disks 21 and 23 of the first and second toroidal continuously 
variable gear mechanisms 20 and 30 from the integrated output disk 34. On 
the other hand, when releasing the low mode clutch 60 and locking the high 
mode clutch 70 simultaneously at a specified timing after the vehicle 
starts to move forward, rotation imparted to the input shaft 11 from the 
engine 1 is admitted to the input disks 21 and 31 of the first and second 
toroidal continuously variable gear mechanisms 20 and 30 through the 
loading cam 40 and further to the integrated output disk 34 through the 
rollers 23 and 33, and then further imparted to the secondary shaft 13 
through the high mode clutch 70 via the high mode gear train 90. At this 
time, the planetary gear mechanism 50 races, the continuously variable 
transmission 10 provides a resultant gear ratio depending only upon the 
gear ratios of the first and second toroidal continuously variable gear 
mechanisms 20 and 30. In other words, the continuously variable 
transmission 10 is controlled to continuously vary its resultant gear 
ratio in the high mode gear ratio control. 
FIG. 8 shows the hydraulic control circuit 200 comprised of the speed 
change control unit 110 and the clutch control unit 120 by which the 
continuously variable transmission 10 is controlled in operation. The 
hydraulic control circuit 200 includes various spool valves, namely a 
regulator valve 201 for regulating the pressure of a working oil 
discharged from the oil pump 102 to a specified level of pressure and 
delivering it into a main pressure line 201, a relief valve 204 for 
regulating the primary pressure in the main pressure line 201 to a 
specified level of relief pressure and delivering it into a relief 
pressure line 203, and a manual shift valve 208 operated by a manual range 
shift stick (not shown) to bring the main pressure line 201 into 
communication with first and second primary pressure lines 205 and 206 in 
a drive (D) range or with first and third primary pressure line 205 and 
207 in a reverse (R) range or to disconnect communication of the main 
pressure line with all of the first to third primary pressure lines 
205-207 in a neutral (N) range or a park (P) range. The regulator valve 
202 and the relief valve 204 are accompanied with linear solenoid valves 
209 and 210, respectively. Each linear solenoid valve 209, 210 generates a 
control pressure based on a pressure regulated to a specified level by a 
reducing valve 211. The regulator valve 202 receives the control pressure 
at its control pressure port 202a to regulate the specified level of line 
pressure. Similarly, the relief valve 204 receives the control pressure at 
its control pressure port 204a to regulate the specified level of relief 
pressure. 
The hydraulic control circuit 200 further includes three spool valves, 
namely a double-slider forward shift valve (which will be refereed to as a 
forward shift valve for simplicity) 220 for developing a speed increase 
hydraulic pressure PH according a line pressure and a relief pressure in 
the drive (D) range, a double-slider reverse shift valve (which will be 
refereed to as a reverse shift valve for simplicity) 230 for developing a 
speed reduction hydraulic pressure PL in the reverse (R) range, and a 
shift valve 241 for actuating selectively the shift valves 220 and 230. 
The shift valve 241 shifts its spool between two positions according to 
whether a pressure is present at the control pressure port 241a. 
Specifically, the shift valve 241 shifts the spool to the right end 
position as seen in FIG. 10 to bring the main pressure line 201 into 
communication with a line pressure line 242 leading to the forward shift 
valve 220 when receiving no line pressure at the control pressure port 
241a, or to the left end position to bring the main pressure line 201 into 
communication with a pressure line 243 leading to the reverse shift valve 
230 when receiving the line pressure at the control pressure port 241a. 
The shift valves 220 and 230 are of the same structure. The forward shift 
valve 220 has an outer sleeve 222 fitted for axial slide movement into an 
axial bore 221 (see FIG. 11) formed in a valve body 111a of the hydraulic 
control section 111 of the shift control unit 110 and an inner sleeve 223 
fitted for axial slide movement into the outer sleeve 222, and the reverse 
shift valve 230 has a sleeve 232 fitted for axial slide movement into an 
axial bore 231 (see FIG. 9) formed in a valve body 111a of the hydraulic 
control section 111 of the shift control unit 110 and an inner sleeve 233 
fitted for axial slide movement into the outer sleeve 232. The forward 
shift valve 220 has five ports, namely a line port 224 disposed at the 
middle in an axial direction and connected to the line pressure line 242, 
first and second relief ports 225 and 226 disposed at opposite ends and 
connected to the relief pressure line 203, a speed increase pressure port 
227 disposed between the line pressure port 224 and the first relief port 
225 and a speed reduction pressure port 228 disposed between the line 
pressure port 224 and the second relief port 226. Similarly, the reverse 
shift valve 230 has five ports, namely a line pressure port 234 disposed 
at the middle in an axial direction and connected to the line pressure 
line 242, first and second relief ports 235 and 236 disposed at opposite 
ends and connected to the relief pressure line 203, a speed increase 
pressure port 237 disposed between the line pressure port 234 and the 
first relief port 235 and a speed reduction pressure port 238 disposed 
between the line pressure port 234 and the second relief port 236. 
The pressure lines 244 and 245 respectively extending from the speed 
increase pressure ports 227 and 237 of the forward and reverse shift valve 
220 and 230 and the pressure lines 246 and 247 respectively extending from 
the speed reduction pressure ports 228 and 238 of the forward and reverse 
shift valve 220 and 230 are connected to the shift valve 241. Shift valve 
241 places its valve spool to the right end position to bring the pressure 
lines 244 and 246 extending from the speed increase and reduction pressure 
ports 227 and 228 of the forward shift valve 220, respectively, into 
communication with a pressure line 248 leading to the speed increase 
pressure chambers 1151 and 1152 of the trunnion driving section 112 and a 
pressure line 249 leading to the speed reduction pressure chambers 1161 
and 1162 of the trunnion driving section 112, respectively. On the other 
hand, the shift valve 241 places its valve spool to the left end position 
to bring the pressure lines 245 and 247 extending from the speed increase 
and reduction pressure ports 237 and 238 of the reverse shift valve 230, 
respectively, into communication with the pressure lines 248 and 249, 
respectively. 
When the forward shift valve 220 shifts the valve sleeve 222 in position 
toward the right relative to the valve spool 223 from a neutral position 
shown in FIG. 8, the valve sleeve 222 increases an inter-communication 
opening between the line pressure port 224 and the speed increase pressure 
port 227, and an inter-communication opening between the second relief 
port 226 and the speed reduction pressure port 228. As a result, the speed 
increase hydraulic pressure PH rises and the speed decrease hydraulic 
pressure PL drops. Conversely, when the forward shift valve 220 changes 
the valve sleeve 222 in position toward the left relative to the valve 
spool 223 from the neutral position, the valve sleeve 222 increases an 
inter-communication opening between the line pressure port 224 and the 
speed reduction pressure port 228, and an inter-communication opening 
between the first relief port 225 and the speed increase pressure port 
227, so that the speed increase hydraulic pressure PH rises and the speed 
decrease hydraulic pressure PL drops. 
FIGS. 9 and 10 show a cam mechanism 260 which moves axially each spool 223, 
233 of the forward and reverse shift valves 220 and 230 against a return 
spring 229, 239 according to axial movement of the valve sleeve 222, 232 
caused by a stepping motor 251, 252. The stepping motors 251 and 252 are 
connected to the valve sleeves 222 and 232 through connecting members 253 
and 254, respectively. The cam mechanism 260 includes a cam 261, a shaft 
262, a cam follower lever or slider 263 and drive levers 264 and 265. The 
cam 261 having a cam face 261a is mounted to a trunnion rod 27 of the 
trunnion 25 of the second toroidal continuously variable gear mechanism 
30. The shaft 262 is disposed adjacent and perpendicularly to the valve 
spools 223 and 233 and supported for rotation by the valve body 111a of 
the hydraulic control section 111. The cam follower lever 263 is attached 
at one of its ends to the shaft and is forced at another end to contact 
with the cam face 261a of the cam 261. The drive lever 264 for forward 
shift is attached at one of its ends to the shaft 262 and is engaged at 
another end with an end key slot 223a of the valve spool 223 of the 
forward shift valve 220. Similarly, the drive lever 265 for reverse shift 
is attached at one of its ends to the shaft 262 and is engaged at another 
end with an end key slot 233a of the valve spool 233 of the reverse shift 
valve 230. 
When the first roller 331 of the second toroidal continuously variable gear 
mechanism 30 inclines to turn the trunnion 251 and trunnion rod 27 
together about the axis X1, the cam 261 turns to force the cam follower 
lever 263, as a result of which the drive levers 264 and 265 are turned 
through a same angle by means of the shaft 262. In this way, the valve 
spools 223 and 233 of the forward and reverse shift valve 220 and 230 are 
axially shifted according to the angle of inclination of the roller first 
roller 331. Accordingly, the axial spool position depends upon the angle 
of inclination of the rollers 33 of the second toroidal continuously 
variable gear mechanism 30, and also of the rollers 23 of the first 
toroidal continuously variable gear mechanism 20, i.e. upon the resultant 
gear ratio of the toroidal continuously variable gear mechanisms 20 and 
30. 
Referring back to FIG. 8, the hydraulic pressure control circuit 200 is 
provided with first and second solenoid valves 271 and 272 for clutch 
control. The first solenoid valve 271 is communicated with the manual 
shift valve 208 through the first primary pressure line 205. Similarly, 
the second solenoid valve 272 is communicated with the manual shift valve 
208 through the second primary pressure line 206. When the first solenoid 
valve 271 opens, a clutch locking pressure produced by regulating the line 
pressure from the first primary pressure line 205 is supplied into the 
hydraulic pressure chamber 65 of the low mode clutch 60 through a clutch 
pressure line 274 via a fail-safe valve 273 to lock up the low mode clutch 
60. Similarly, when the second solenoid valve 272 opens, a clutch locking 
pressure produced by regulating the line pressure from the second primary 
pressure line 206 is supplied into the hydraulic pressure chamber 75 of 
the high mode clutch 70 through a clutch pressure line 275 to lock up the 
high mode clutch 70. The clutch pressure lines 274 and 275 are accompanied 
with accumulators 276 and 277, respectively, to provide gradual 
development of the clutch locking pressure in the hydraulic pressure 
chambers 65 and 75 so as thereby to prevent an occurrence of shift shocks. 
The third primary pressure line 207 extending from the manual shift valve 
208 is connected to the control pressure port 241a of the shift valve 241 
via the fail-safe valve 273. The shift valve 241 receives a line pressure 
at its control pressure port 241a to place the valve spool to the left end 
position (reverse position) when the manual shift valve 208 is in the 
reverse (R) range position. The fail-safe valve 273 is accompanied with a 
solenoid valve 278. The solenoid valve 278 provides a control pressure to 
force the valve spool of the fail-safe valve 273 to the right end position 
so as thereby to bring the first primary pressure line 205 in 
communication with the low mode clutch pressure line 274. The solenoid 
valves 271, 272 and 278 are of a three-way type valve which drains a 
downstream side when both upstream and downstream sides are shut off. 
The hydraulic control circuit 200 is further provided with a lubrication 
oil line 281 which extends from a drain port of the regulator valve 202 
and branches off to a lubrication oil line 282 leading to the first and 
second toroidal continuously variable gear mechanisms 20 and 30 and a 
lubrication oil line 283 leading to continuously variable transmission 
parts other than the toroidal continuously variable gear mechanisms 20 and 
30. The lubrication oil line 281 is provided with a relief valve 284 to 
adjust the lubrication oil at a specified level of pressure. An upstream 
part of the lubrication oil line 282 branches off into an oil line 286 
provided with a cooler 285 for cooling the lubrication oil and an oil line 
287 bypassing the cooler 285. The oil line 286 upstream from the cooler 
285 is provided with an orifice 288 and a first switch valve 289 disposed 
in parallel. The bypass oil line 287 is provided with a second switch 
valve 290. Delivery of the lubrication oil to the first and second 
continuously variable transmission mechanisms 20 and 30 is controlled by 
means of the first and second switch valves 289 and 290. According to 
incoming signals from a control unit 300 (which will e be described in 
detail later with reference to FIG. 13) comprised mainly of a 
microprocessor, the second switch valve 290 opens to permit the working 
oil or lubrication oil to flow to the toroidal continuously variable gear 
mechanisms 20 and 30 bypassing the cooler 285 when the lubrication oil is 
at a temperature lower than a specified temperature and at a pressure 
higher than a specified pressure for the purpose of preventing aggravation 
of flowability of the lubrication oil due to mechanical resistance of the 
cooler 285 and preventing the cooler 285 from encountering damages and a 
decrease in durability due to the high pressure lubrication oil. In all 
other cases, the second switch valve 290 closes to make the lubrication 
oil pass through the cooler before reaching the toroidal continuously 
variable gear mechanisms 20 and 30. By this way, oil films on the toroidal 
surfaces of the input and output disks 21, 22, 31 and 32 are maintained in 
good conditions, so as to protect contact surfaces of the toroidal disks 
with the rollers 23 and 33. The first switch valve 289 opens and closes 
according to incoming signals from the control unit 300. Specifically, the 
first switch valve 289 closes when, while the second switch valve 290 
remains closed, the engine 1 operates at a speed of rotation lower than a 
specified speed of rotation and the vehicle runs at a velocity lower than 
a specified velocity. This is because, in the case of lower engine speeds 
of rotation and/or lower vehicle velocities, while the toroidal 
continuously variable gear mechanisms 20 and 30 has a demand for a small 
amount of lubrication oil, the low and high mode clutches 60 and 70 needs 
a specified amount of lubrication oil. The lubrication oil supplied to the 
toroidal continuously variable gear mechanisms 20 and 30 through the 
lubrication oil line 282 is also supplied to bearings supporting the 
rollers 23 and 33 through an oil line 282a (see FIG. 3) and sprayed on the 
toroidal surfaces through a nozzle 282b (see FIG. 3). 
Gear ratio control of the continuously variable transmission 10 depicted in 
FIG. 1 is executed through the control unit 300 shown in block diagram in 
FIG. 16. 
Referring to FIG. 11, the control unit 300 receives various signals from 
sensors and switches including at least an engine speed sensor 302, a 
throttle position or opening sensor 303, a transmission position sensor 
304, an oil temperature sensor 305, speed sensors 306 and 307 and an idle 
switch 308. The speed sensor 306 attached to the low mode clutch drum 61 
detects a speed of input disk 21 of the first toroidal continuously 
variable gear mechanism 20, and the speed sensor 307 attached to the 
second gear 92 of the high mode gear train 90 detects a speed of input 
disk 31 of the first toroidal continuously variable gear mechanism 30. The 
idle switch 308 detects release of the accelerator pedal. These sensors 
and switches are well known in various types in the art and may take any 
known type. The control unit 300 provides control signals for various 
solenoid valves 209, 210, 271, 272, 278, 289 and 290, stepping motors 251 
and 252, and other electrically controlled elements in the hydraulic 
control circuit 200 according to driving conditions represented by signals 
from the switches and sensors 301-308. 
The following description will be directed to basic speed change operation 
of the continuously variable transmission 10. As was previously described 
above, the hydraulic control circuit 200 shown in FIG. 10 is in the drive 
(D) range in which the manual shift valve 208 takes the drive (D) position 
to force the shift valve 241 to maintain the valve spool in the right end 
position (the forward position). Because the toroidal continuously 
variable gear mechanisms 20 and 30 shown in FIG. 3 operate in the same 
way, the explanation will be given relating to the roller 231 and trunnion 
251 of the first toroidal continuously variable gear mechanism 20 by way 
of example, and the same is true for other rollers and trunnions. 
When the hydraulic control circuit 200 is actuated in response to a signal 
from the control unit 300, the solenoid valves 209 and 210 are actuated to 
generate a specified level of pressure as a line pressure at the control 
pressure port 202a of the regulator valve 202 and a specified level of 
pressure as a relief pressure at the control pressure port 204a of the 
relief valve 204. The line pressure is supplied to the inlet port 224 of 
the forward shift valve 220 through the main pressure line 201 and 
pressure line 242 via the shift valve 241, and the relief pressure is 
supplied to the first and second relief ports 225 and 226 of the forward 
shift valve 220 through the pressure line 203. Based on these line 
pressure and relief pressure, the forward shift valve 220 controls the 
hydraulic pressure difference (.DELTA.P=PH-PL) between a hydraulic 
pressure PH for speed increase and a hydraulic pressure PL for speed 
reduction. This hydraulic pressure difference control is performed to hold 
the trunnion 35 and roller 33 in the neutral positions against traction 
force T exerted on the trunnion 35 or to force them in the axial direction 
X from the neutral position so as to vary the inclination of the roller 
for varying the gear ratio of the toroidal continuously variable gear 
mechanism 30. As shown in FIG. 10, when the roller 33 is driven by the 
input disk 21 rotating in a direction indicated by an arrow d, these 
trunnion 35 and roller 33 are applied with force by which they are drawn 
in the same direction d. On the other hand, when the output disk 32 is 
rotated in a direction indicated by an arrow f by the roller 33 rotating 
in a direction indicated by an arrow f, reaction force is exerted as 
traction force T in a direction opposite to the rotational direction e of 
the output disk 32 on the roller 33 and trunnion 35. In order to hold the 
roller 33 in the neutral position against the traction force T, the speed 
increase and reduction pressure chambers 115 and 116 are supplied with 
speed increase and speed decrease hydraulic pressures PH and PL, 
respectively which are controlled to provide a hydraulic pressure 
difference (.DELTA.P=PH-PL) balanced with the traction force T. When 
increasing the gear ratio of the toroidal continuously variable gear 
mechanism 30 for forward drive of the vehicle, the forward shift valve 220 
forces the valve sleeve 222 to shift toward in a direction indicated by an 
arrow g in FIG. 10 to decrease the inter-communication openings between 
the inlet pressure port 224 and the speed increase pressure port 227 and 
between the second relief port 226 and the speed reduction pressure port 
228. As a result, the hydraulic pressure PH introduced into the speed 
increase pressure chamber 115 (see FIG. 8) rises due to the relief 
pressure which is relatively high, and the hydraulic pressure PL 
introduced into the speed reduction pressure chamber 116 drops due to the 
line pressure which is relatively low, as a result of which the traction 
force T becomes higher than the hydraulic pressure difference 
(.DELTA.P=PH-PL) to force the trunnion 25 and roller 23 in a direction 
indicated by an arrow h as shown in FIG. 10. Following the movement, the 
roller 33 inclines in a direction in which it shifts its contact point 
with the input disk 31 radially outward and its contact point with the 
output disk 32 radially inward to reduce the gear ratio. The same 
inclination of the roller 33 is caused in the second toroidal continuously 
variable gear mechanism 30. Due to the traction force T exceeding the 
hydraulic pressure difference (.DELTA.P=PH-PL), the roller 33 inclines in 
a direction in which it shifts its contact point with the input disk 31 
radially outward and its contact point with the output disk 32 radially 
inward following movement of the trunnion 35 in a direction indicated by 
an arrow h in FIG. 10. At this time, however, the cam 261 of the cam 
mechanism 260 turns through the same angle as the roller 33 in the same 
direction as indicated by an arrow i in FIG. 9, as a result of which the 
cam follower lever 263, and hence the shaft 262 and the drive lever 264, 
turns in a direction indicated by an arrow j shown in FIG. 10. 
Consequently, the forward shift valve 220 shifts the valve spool 223 in a 
direction indicated by an arrow k in FIG. 10, under influence of the 
return spring 229. Because this direction k is coincident with the 
direction in which the valve sleeve 222 is shifted by the stepping motor 
251, the inter-communication openings between the inlet port 224 and the 
speed increase pressure port 227 and between the second relief port 226 
and the speed reduction pressure port 228 regain their initial balanced 
positions, so as to balance the hydraulic pressure difference 
(.DELTA.P=PH-PL) with the traction force T, thereby achieving the gear 
ratio change of the toroidal continuously variable gear mechanism 30. The 
gear ratio change of the toroidal continuously variable gear mechanism 30 
is fixed at the gear ratio. In this instance, the speed change is 
completed at a point of time that the valve spool 223 reaches the balanced 
position relative to the valve sleeve 222. Since the neutral position is 
the position to which the valve sleeve 222 has been shifted by the 
stepping motor 251 and corresponds to the inclined angle of the roller 33 
caused by the cam mechanism 260, the position of the valve sleeve 222 
corresponds to the inclined angle of the roller 33, and hence the trunnion 
35. This means that the controlled variable of the stepping motor 251 
determines the gear ratio of the toroidal continuously variable gear 
mechanism 30. Accordingly, the gear ratio of the toroidal continuously 
variable gear mechanism 30 is varied according to the number of pulses 
applied to the stepping motor 251. The gear ratio control of the toroidal 
continuously variable gear mechanism 20 is achieved in the same way when 
the valve sleeve 222 of the forward shift valve 220 has been shifted in 
the opposite direction. In this instance, the toroidal continuously 
variable gear mechanism 30 increases its gear ratio. 
FIG. 12 shows the relationship between the number of pulses for the 
stepping motor 251, 252 and gear ratio of the toroidal continuously 
variable gear mechanism 20, 30. As apparent, the gear ratio of the 
toroidal continuously variable gear mechanism 20, 30 decreases with an 
increase in the number of pulses. 
FIG. 13 shows the relationship between the number of pulses for the 
stepping motor 251, 252 and resultant gear ratio N of the continuously 
variable transmission 10 by the control of gear ratios of the toroidal 
continuously variable gear mechanisms 20 and 30. Although the following 
description is directed to the toroidal continuously variable gear 
mechanism 30 only, the same description is true for the toroidal 
continuously variable gear mechanism 20. As was previously described, 
while the toroidal continuously variable gear mechanisms 30 changes its 
gear ratio according to the number of pulses applied to the stepping 
motors 251 and 261, the continuously variable transmission 10 provides a 
resultant gear ratio N which differs according to modes of gear ratio 
control, i.e. according to which mode clutch 60 or 70 has been locked. 
When the continuously variable transmission 10 is put in the high mode 
gear ratio control, rotation of the integrated output disk 34 is imparted 
directly to the secondary shaft 13 through the high mode gear train 90 and 
the high mode clutch 70 locked up. As shown in FIG. 13, the characteristic 
curve of resultant gear ratio N of the continuously variable transmission 
10 relative to the number of pulses agrees with the characteristic curve 
of gear ratio of the toroidal continuously variable gear mechanism 30 
shown in FIG. 12. It is of course that these gear ratios of the toroidal 
continuously variable gear mechanism 30 are different from each other 
according to the difference in diameter of the first and second gears 91 
and 92 of the high mode gear train 90 or the number of teeth between the 
first and second gears 91 and 92 of the high mode gear train 90. On the 
other hand, when the continuously variable transmission 10 is put in the 
low mode gear ratio control, while rotation of the engine 1 is imparted to 
the pinion carrier 51 of the planetary gear mechanism 50 from the input 
shaft 11 through the low mode gear train 80 and the low mode clutch 60 
locked up, rotation of the integrated output disk 34 is imparted to the 
sun gear 52 of the planetary gear mechanism 50 through the high mode gear 
train 90. In this case, when the planetary gear mechanism 50 holds the 
internal gear 53, which is a transmission output gear, at a rotational 
speed of 0 (zero) by controlling the toroidal continuously variable gear 
mechanism 30 to provide a specified ratio of input rotational speeds 
between the pinion carrier 51 and sun gear 52, the continuously variable 
transmission 10 is put in the geared neutral state in which the internal 
gear 53 stands still. At this time, while the resultant gear ratio N 
increases infinitely as indicated by arrows P and Q in FIG. 13, when the 
numbers of pulses admitted to the stepping motors 251 and 252 are 
subsequently decreased so that the toroidal continuously variable gear 
mechanism 30 is forced to increase their gear ratios and the speed of 
rotation imparted to the sun gear 52 of the planetary gear mechanism 50 
drops consequently, the planetary gear mechanism 50 causes the internal 
gear 53 to start rotation in the forward direction. In this way, the 
continuously variable transmission 10 reduces its resultant gear ratio N 
following reduction in the number of pulses admitted to the stepping motor 
251, 252, establishing the low mode gear ratio control in the drive (D) 
range where the resultant gear ratio N varies along the characteristic 
curve L. These high and low gear ratio control characteristic curves H and 
L in the drive (D) range intersect at a gear ratio (switchover gear ratio) 
Gc of, for example, approximately 1.8 indicated by an arrow R which is 
provided correspondingly to 500 pulses. Accordingly, the gear ratio 
control is changed over during varying continuously the resultant gear 
ratio N of the continuously variable transmission 10 by switching lock-up 
from one to another between the low and high mode clutches 60 and 70 at 
the toroidal gear ratio. On the other hand, when the number of pulses to 
the stepping motor 251, 252 is increased subsequently to achievement of 
the geared neutral state so that the toroidal continuously variable gear 
mechanism 30 is forced to reduces their gear ratios and the speed of 
rotation imparted to the sun gear 52 of the planetary gear mechanism 50 
rises consequently, the planetary gear mechanism 50 causes the internal 
gear 53 to start rotation in the reverse direction. In this way, the 
continuously variable transmission 10 provides the reverse mode of gear 
ratio control in the reverse (R) range where the resultant gear ratio N 
increases along the characteristic curve R as the number of pulses 
increases. 
The control unit 300 controls the resultant gear ratio N of the 
continuously variable transmission 10 based on the characteristic curves 
according to driving conditions. Specifically, the control unit 300 
detects current vehicle speed V and throttle opening .theta. based on 
incoming signals from the speed sensor 301 and throttle position sensor 
303 to determine a target engine speed of rotation Neo with reference to a 
gear ratio control map shown in FIG. 14. Pulsing control of the first and 
second stepping motor 251 and 252 and locking and unlocking control of the 
mode clutches 60 and 70 are performed so that the continuously variable 
transmission 10 provides a resultant gear ratio N (which is represented by 
an angle .alpha. in FIG. 14) corresponding to the target engine speed of 
rotation Neo on the basis of the resultant gear ratio control curve L, H 
or R. 
Incidentally, as was previously described in connection with FIG. 24, the 
roller 33 is forced off its neutral position in a direction opposite to 
the direction of traction force T. Accordingly, when input torque 
transmitting through the toroidal continuously variable gear mechanism 30 
is reversed, the roller 33 switches its shift from one side to another 
side of the neutral position following reversal of the traction force T in 
direction. 
Referring to FIG. 15 showing the relationship between transfer torque 
through the toroidal continuously variable gear mechanism 30 and shift 
distance of the roller 33 based on actual measurements. In FIG. 15, the 
transfer torque passing from the input toroidal disk 31 to the output 
toroidal disk 32 takes a plus value, and the transfer torque passing from 
the output toroidal disk 32 to the input toroidal disk 31 takes a minus 
value. The shift distance D takes a plus value when the roller 33 shifts 
toward the higher speed side while the transfer torque is plus and takes a 
minus value when the roller 33 shifts toward the lower speed side while 
the transfer torque is minus. Accordingly the transfer torque and shift 
distance D take plus values, respectively, while the toroidal continuously 
variable gear mechanism 30 is in the normal drive state in the high mode 
or in the reverse drive state in the low mode, and they take minus values, 
respectively, while the toroidal continuously variable gear mechanism 30 
is in the reverse drive state in the high mode or in the normal drive 
state in the low mode. When the toroidal continuously variable gear 
mechanism 30 switches to the low mode from the high mode in the normal 
drive state or switches to the reverse drive mode from the normal drive 
state in the low mode, the transfer torque reverses in transfer direction, 
so that the shift distance D of the roller 33 turns from a minus value to 
a plus value as indicated by an arrow Da. The roller 33 inclines to change 
a gear ratio due to a shift thereof following directional reversal of 
transfer torque, which is always accompanied by a shift shock. The control 
unit 300 is designed and adapted to perform gear ratio control in which an 
occurrence of a shift shock is prevented or significantly restrained. 
FIG. 16 shows a flow chart illustrating a main sequence routine of the 
shift shock restrained gear ratio control for the microprocessor of the 
control unit 300. 
When the flow chart logic commences and control proceeds to a function 
block at step S101 where a current gear ratio control mode and driving 
conditions are detected. Subsequently, at step S102, a judgement is made 
as to whether in which mode the toroidal continuously variable gear 
mechanism 30 is put. when in the low mode, a resultant gear ratio Gf is 
compared with the switchover gear ratio Gc at step S103. When the 
resultant gear ratio Gf has dropped below the switchover gear ratio Gc, 
mode switching control for the switch to the high mode from the low mode 
(which is hereafter referred to as low-to-high mode switching control) is 
performed at step S104. On the other hand, when in the high mode, a 
resultant gear ratio Gf is compared with the switchover gear ratio Gc at 
step S105. When the resultant gear ratio Gf has risen above the switchover 
gear ratio Gc, mode switching control for the switch to the low mode from 
the high mode (which is hereafter referred to as high-to-low mode 
switching control) is performed at step S106. However, when the resultant 
gear ratio Gf has not yet dropped below the switchover gear ratio Gc even 
while the toroidal continuously variable gear mechanism 30 is put in the 
low mode or when the resultant gear ratio Gf has not yet risen above the 
switchover gear ratio Gc even while the toroidal continuously variable 
gear mechanism 30 is put in the high mode, a judgement is made at step 
S107 as to whether the drive state has been switched. When the drive state 
has been switched, a further judgement is made at step S308 as to whether 
the drive state has been switched to the normal state from the reverse 
state or vice versa. According to switches of the drive state, shift shock 
restrained gear ratio control separately provided for the switch from the 
normal state to the reverse state or for the switch from the reverse state 
to the normal state are executed at step S109 or S110. 
FIG. 17 shows a flow chart illustrating a sequence subroutine of the 
low-to-high mode switchover control for the microprocessor of the control 
unit 300. 
When the low-to-high mode switchover control subroutine commences and 
control proceeds to a function block at step S201 where the second 
solenoid valve 272 is actuated to lock the high mode clutch 70. 
Subsequently, a judgement is made at step S202 as to whether the toroidal 
continuously variable gear mechanism 30 is in the normal drive state or in 
the reverse drive state. When the toroidal continuously variable gear 
mechanism 30 is in the nor mal drive state, a target shift distance of the 
roller 33 of the toroidal continuously variable gear mechanism 30 toward 
the higher speed side is determined at step S203. That is, upon an 
occurrence of switch to the high mode from the low mode in the normal 
drive state, the toroidal continuously variable gear mechanism 30 turns 
the transfer of torque in direction to the normal from the reverse 
direction with a result of turning the shift distance D of the roller 33 
to the plus side or the higher speed side from tile minus side or the 
lower speed side, as indicated by the arrow Da in FIG. 15. In this 
instance, the shift distance D of the roller 33, which corresponds to 
input torque to the toroidal continuously variable gear mechanism 30, is 
determined by the use of a characteristic curve specified in the map shown 
in FIG. 15. The input torque to the toroidal continuously variable gear 
mechanism 30 is proportional to the hydraulic pressure difference 
(.DELTA.P=PH-PL) and consequently determined based on the hydraulic 
pressure difference (.DELTA.P=PH-PL). 
Thereafter, the number of drive pulses for the stepping motor 251 necessary 
to move the valve sleeve 222 of the forward shift valve 220 by a distance 
corresponding to the target shift distance D of the roller 33 is 
determined at step S204, and the stepping motor 251 is driven with the 
pulses to shift the roller 33 to the higher speed side at step S205. 
On the other hand, when the toroidal continuously variable gear mechanism 
30 is in the reverse drive state, a target shift distance of the roller 33 
of the toroidal continuously variable gear mechanism 30 toward the lower 
speed side is determined at step S207. At this time, the toroidal 
continuously variable gear mechanism 30 turns the transfer of torque in 
direction to the reverse from the normal direction with a result of 
turning the shift distance D of the roller 33 to the minus side or the 
lower speed side from the plus side or the higher speed side, as indicated 
by the arrow Db in FIG. 15. Thereafter, the number of pulses for the 
stepping motor 251 necessary to cause the target shift distance D of the 
roller 33 is determined at step S208, and the stepping motor 251 is driven 
with the pulses to shift the roller 33 to the lower speed side at step 
S209. 
Subsequently to shifting the roller 33 at step S205 or S209, the first 
solenoid valve 271 is actuated to unlock the low mode clutch 60 at step 
S210. 
Operation of a switch to the low mode from the high mode is hereafter 
described for the normal drive state by way of example with reference to 
FIG. 18. If the mode clutches 60 and 70 are alternately locked and 
unlocked without controlling the forward shift valve 220 to regulate 
working oil pressure admitted to the roller 33, the toroidal continuously 
variable gear mechanism 30 turns the transfer of torque in direction to 
the normal direction from the reverse direction immediately after 
unlocking the low mode clutch 60 as shown by an arrow A, which is 
accompanied by a shift of the roller 33 to the plus side as shown by an 
arrow B. As a result, the roller 33 inclines toward the higher speed side. 
The inclination of the roller 33 is accompanied by a shift of the valve 
spool 223 of the shift valve 220 through the trunnion 35 as shown by an 
arrow C. A sharp change in gear ratio resulting from the inclination of 
the roller 33 causes a shock as shown by an arrow D. 
As contrasted with the above, a shift direction and a shift distance of the 
roller 33 following directional reversal of the transfer of torque are 
estimated in a period in which both low and high mode clutches 60 and 70 
remain locked and the stepping motor 251 is driven to shift the valve 
sleeve 222 by a distance corresponding to the estimated shift distance of 
the roller 33 in the shift direction in advance as shown by an arrow E. An 
inclination of the roller 33 following the shift causes gradual 
directional reversal of transfer torque as shown by an arrow F. Since the 
roller 33 has been shifted by the estimated shift distance at a point of 
time that the low mode clutch 60 is unlocked, there is no shock due to a 
sharp change in gear ratio resulting from the inclination of the roller 33 
as shown by an arrow G. 
FIG. 19 shows a flow chart illustrating a sequence subroutine of the 
high-to-low mode switchover control for the microprocessor of the control 
unit 300. 
When the low-to-high mode switchover control subroutine commences and 
control proceeds to a function block at step S301 where the first solenoid 
valve 271 is actuated to lock the low mode clutch 60. Subsequently, a 
judgement is made at step S302 as to whether the toroidal continuously 
variable gear mechanism 30 is in the normal drive state or in the reverse 
drive state. When the toroidal continuously variable gear mechanism 30 is 
in the normal drive state, a target shift distance of the roller 33 of the 
toroidal continuously variable gear mechanism 30 toward the lower speed 
side is determined at step S303. That is, upon an occurrence of switch to 
the low mode from the high mode in the normal drive state, the toroidal 
continuously variable gear mechanism 30 turns the transfer of torque in 
direction to the reverse from the normal direction with a result of 
turning the shift distance D of the roller 33 to the minus side or the 
lower speed side from the plus side or the higher speed side, as indicated 
by the arrow Db in FIG. 15. Subsequently, the number of drive pulses for 
the stepping motor 251 necessary to move the valve sleeve 222 of the 
forward shift valve 220 by a distance corresponding to the target shift 
distance D of the roller 33 is determined at step S304, and the stepping 
motor 251 is driven with the pulses to shift the roller 33 to the lower 
speed side at step S305. 
On the other hand, when the toroidal continuously variable gear mechanism 
30 is in the reverse drive state, a target shift distance of the roller 33 
of the toroidal continuously variable gear mechanism 30 toward the higher 
speed side is determined at step S307. At this time, the toroidal 
continuously variable gear mechanism 30 turns the transfer of torque in 
direction to the normal from the reverse direction with a result of 
turning the shift distance D of the roller 33 to the plus side or the 
higher speed side from the minus side or the lower speed side, as 
indicated by the arrow Da in FIG. 15. Thereafter, the number of pulses for 
the stepping motor 251 necessary to cause the target shift distance D of 
the roller 33 is determined at step S308, and the stepping motor 251 is 
driven with the pulses to shift the roller 33 to the higher speed side at 
step S309. 
Subsequently to shifting the roller 33 at step S305 or S309, the first 
solenoid valve 271 is actuated to unlock the high mode clutch 70 at step 
S310. 
Operation of a switch to the low mode from the high mode is hereafter 
described for the reverse drive state by way of example with reference to 
FIG. 20. If the mode clutches 60 and 70 are alternately locked and 
unlocked without controlling the shift valve 220 to regulate working oil 
pressure admitted to the roller 33, the toroidal continuously variable 
gear mechanism 30 turns the transfer of torque in direction to the normal 
direction from the reverse direction immediately after unlocking the high 
mode clutch 70 as shown by an arrow H, which is accompanied by a shift of 
the roller 33 to the plus side as shown by an arrow J. As a result, the 
roller 33 inclines toward the higher speed side. The inclination of the 
roller 33 is accompanied by a shift of the valve spool 223 of the shift 
valve 220 as shown by an arrow K. A sharp change in gear ratio resulting 
from the inclination of the roller 33 causes a shock as shown by an arrow 
L. 
As contrasted with the above, a shift direction and a shift distance of the 
roller 33 following directional reversal of the transfer of torque are 
estimated in a period in which both low and high mode clutches 60 and 70 
remain locked and the stepping motor 251 is driven to shift the valve 
sleeve 222 by a distance corresponding to the estimated shift distance of 
the roller 33 in the shift direction in advance as shown by an arrow M. An 
inclination of the roller 33 following the shift causes gradual 
directional reversal of transfer torque as shown by an arrow N. Since the 
roller 33 has been shifted by the estimated shift distance at a point of 
time that the low mode clutch 60 is unlocked, there is no shock due to a 
sharp change in gear ratio resulting from the inclination of the roller 33 
as shown by an arrow P. While there is provided a period of time for which 
both low and high mode clutches 60 and 70 remain locked during a switch of 
modes, the gear ratio is fixed at the switchover gear ratio Gc for the 
period of time, so as to prevent the engine from suddenly increasing its 
speed with an effect of causing a steady shift of the roller 30 under this 
stable engine operating condition. 
FIG. 21 shows a flow chart illustrating a sequence subroutine of the shift 
shock restrained gear ratio control during normal-to-reverse state 
switching for the microprocessor of the control unit 300. 
When the low-to-high mode switchover control subroutine commences and 
control proceeds to a function block at step S401 where a judgement is 
made as to whether the toroidal continuously variable gear mechanism 30 is 
in the high mode or in the low mode. When the toroidal continuously 
variable gear mechanism 30 is in the low mode, a target shift distance of 
the roller 33 of the toroidal continuously variable gear mechanism 30 
toward the higher speed side is determined based on the input torque and 
gear ratio at step S402. Subsequently, the number of drive pulses for the 
stepping motor 251 necessary to move the valve sleeve 222 of the forward 
shift valve 220 by a distance corresponding to the target shift distance D 
of the roller 33 is determined at step S403, and the stepping motor 251 is 
driven with the pulses to shift the roller 33 to the higher speed side at 
step S404. 
On the other hand, when the toroidal continuously variable gear mechanism 
30 is in the high mode, a target shift distance of the roller 33 of the 
toroidal continuously variable gear mechanism 30 toward the lower speed 
side is determined based on the input torque and gear ratio at step S405. 
Subsequently, the number of pulses for the stepping motor 251 necessary to 
cause the target shift distance D of the roller 33 is determined at step 
S406, and the stepping motor 251 is driven with the pulses to shift the 
roller 33 to the lower speed side at step S407. 
That is, upon an occurrence of a switch to the reverse drive state from the 
normal drive state in the low mode, the toroidal continuously variable 
gear mechanism 30 turns the transfer of torque in direction to the normal 
direction from the reverse direction with a result of turning the shift 
distance D of the roller 33 to the plus side or the higher speed side from 
the minus side or the lower speed side, as indicated by the arrow Da in 
FIG. 15. 
On the other hand, upon an occurrence of a switch to the reverse drive 
state from the normal drive state in the high mode, the toroidal 
continuously variable gear mechanism 30 turns the transfer of torque in 
direction to the reverse direction from the normal direction with a result 
of turning the shift distance D of the roller 33 to the minus side or the 
lower speed side from the plus side or the higher speed side, as indicated 
by the arrow Db in FIG. 15. By virtue of the control, when it is detected 
that the accelerator pedal is released for example, the stepping motor 251 
is actuated to shift the roller 33 through operation of the shift valve 
220 prior to inclination of the roller 33, preventing generation of shocks 
due to steep shift and inclination of the roller 33 resulting from 
directional reversal of the transfer of torque. 
FIG. 22 shows a flow chart illustrating a sequence subroutine of the shift 
shock restrained gear ratio control during a reverse-to-normal state 
switch for the microprocessor of the control unit 300. 
When the low-to-high mode switchover control subroutine commences and 
control proceeds to a function block at step S501 where a judgement is 
made as to whether the toroidal continuously variable gear mechanism 30 is 
in the high mode or in the low mode. When the toroidal continuously 
variable gear mechanism 30 is in the low mode, a target shift distance of 
the roller 33 of the toroidal continuously variable gear mechanism 30 
toward the lower speed side is determined based on the input torque and 
gear ratio at step S502. Subsequently, the number of drive pulses for the 
stepping motor 251 necessary to move the valve sleeve 222 of the forward 
shift valve 220 by a distance corresponding to the target shift distance D 
of the roller 33 is determined at step S503, and the stepping motor 251 is 
driven with the pulses to shift the roller 33 to the lower speed side at 
step S504. 
On the other hand, when the toroidal continuously variable gear mechanism 
30 is in the high mode, a target shift distance of the roller 33 of the 
toroidal continuously variable gear mechanism 30 toward the higher speed 
side is determined based on the input torque and gear ratio at step S505. 
Subsequently, the number of pulses for the stepping motor 251 necessary to 
cause the target shift distance D of the roller 33 is determined at step 
S506, and the stepping motor 251 is driven with the pulses to shift the 
roller 33 to the higher speed side at step S507. 
That is, upon an occurrence of a switch to the normal drive state from the 
reverse drive state in the low mode, the toroidal continuously variable 
gear mechanism 30 turns the transfer of torque in direction to the reverse 
direction from the normal direction with a result of turning the shift 
distance D of the roller 33 to the minus side or the lower speed side from 
the plus side or the higher speed side, as indicated by the arrow Db in 
FIG. 15. On the other hand, upon an occurrence of a switch to the normal 
drive state from the reverse drive state in the high mode, the toroidal 
continuously variable gear mechanism 30 turns the transfer of torque in 
direction to the normal direction from the reverse direction with a result 
of turning the shift distance D of the roller 33 to the plus side or the 
higher speed side from the minus side or the lower speed side, as 
indicated by the arrow Da in FIG. 15. By virtue of the control, when it is 
detected that the accelerator pedal is stepped on for example, the 
stepping motor 251 is actuated to shift the roller 33 through operation of 
the shift valve 220 prior to inclination of the roller 33, preventing 
generation of shocks due to steep shift and inclination of the roller 33 
resulting from directional reversal of the transfer of torque. 
While, in the above embodiment, the double-slider shift valves 220 and 230 
are employed as pressure control valves equipped with a feedback feature 
which are driven by the stepping motors 251 and 252 to control hydraulic 
pressure PH for speed increase and hydraulic pressure PL for speed 
reduction, another type of pressure control valve 320 such as shown in 
FIG. 23 may be incorporated. 
Referring to FIG. 23, the valve 320 has a valve sleeve 322 secured to a 
valve body 321 and a valve spool 323 received to rotate in the valve 
sleeve 322. The valve spool 323 at one of its opposite ends is connected 
to an output shaft of a stepping motor 326 by means of a connecting rod 
324 and a connecting pin 325 and forced toward the opposite end by a 
spring 327. The valve spool 323 at another end is formed with internal 
threads 323a which are engaged by a rod 328 formed with external threads 
328a and an end slot 328b. A rod 331 integral with the trunnion (not 
shown) is provided with a cam 332. There is a crank lever 333 pivoted on 
the valve body 331, one arm of the crank lever 333 being in engagement 
with the cam 332 and another arm of the crank lever 333 being received in 
the end slot 328b of the rod 328. By means of the crank lever 333, 
rotation of the cam 332 is translated into axial movement of the valve 
spool 323. On the other hand, the valve spool 323 is rotated and axially 
moved by the stepping motor 326. During rotation of the valve spool 323, 
the rod 328 is prevented from rotation by means of the engagement with the 
crank lever 333. The valve sleeve 322 is formed with an inlet port 322a, 
and a speed increase pressure port 322b and a speed reduction pressure 
port 322c disposed on opposite sides of the inlet port 322a. The valve 
spool 323 is formed with an external groove 323b which brings the inlet 
port 322a into communication with the speed increase and speed reduction 
pressure ports 322b and 322c according relative positions between the 
valve sleeve 322 and the valve spool 323. 
When the stepping motor 326 is driven to cause predetermined turns of the 
valve spool 323 in a normal direction, the valve spool 323 axially shifts 
from the balanced or neutral position shown in FIG. 23 to bring the inlet 
port 322a into communication with the speed increase pressure port 322b or 
the speed reduction pressure port 322c. Communication of the inlet port 
322a with the speed increase pressure port 322b causes a rise in hydraulic 
pressure PH for speed increase with the result of inclining the roller 33 
toward the higher speed side. The inclination of the roller 33 reactively 
turns the rod 332 and hence the cam 332, which is accompanied by a turn of 
the crank lever 333. As a result, the valve spool 323 is returned to the 
balanced or neutral position and the inlet port 322a is brought out of 
communication with the speed increase pressure port 322b or the speed 
reduction pressure port 322c. In this manner, the angle of inclination of 
the roller 33, i.e. the controlled gear ratio of the toroidal continuously 
variable gear mechanism 20, 30 is maintained. 
As described above, according to the control system for the continuously 
variable transmission incorporating a toroidal continuously variable gear 
mechanism of the invention which causes an estimated shift of the roller 
of the toroidal continuously variable gear mechanism in an estimated 
direction prior to directional reversal of a shift of the roller from the 
neutral position due to directional reversal of the transfer of torque 
through the toroidal continuously variable gear mechanism when there is an 
occurrence of a switch of gear ratio control modes or an occurrence of a 
switch of vehicle drive states, generation of shocks due to steep shift 
and inclination of the roller 33 resulting from directional reversal of 
the transfer of torque is prevented or significantly restrained, which is 
always desirable for operational quality and performance of the 
continuously variable transmission. 
It is to be understood that although the present invention has been 
described with regard to preferred embodiments thereof, various other 
embodiments and variants may occur to those skilled in the art, which are 
within the scope and spirit of the invention, and such other embodiments 
and variants are intended to be covered by the following claims.