JP2002013621A - Controller of power train - Google Patents

Abstract

PROBLEM TO BE SOLVED: To effectively prevent generation of switching shock in switching operation between a low mode and a high mode. SOLUTION: This controller of a power train comprises a target gear ratio setting means 311 for setting a target gear ratio of a transmission, a gear ratio control means 312 for executing the control of the gear ratio of a continuously variable transmission mechanism and the control of the switching of a passage to realize the target gear ratio set by the target gear ratio setting means 311, a torque-adjusting means 309 for adjusting the toque inputted to the transmission from an engine which corresponds to the acceleration operation, and the variation of a torque-inhibiting means 313 for inhibiting the variation of the torque inputted to the transmission from an output source during the execution of the switching control of the passage, when the target gear ratio accompanying the switching of the passage, is set by the target gear ratio setting means.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a vehicle part train, and more particularly to a power train control device using a continuously variable transmission mechanism.

[0002]

2. Description of the Related Art Conventionally, for example, Japanese Patent Application Laid-Open No. 9-210191
As shown in the publication, in a power train of a vehicle including a toroidal-type continuously variable transmission mechanism and a planetary gear mechanism, as a power transmission path between an engine and drive wheels, the continuously variable transmission mechanism and the planetary gear mechanism And a second path that passes only through the continuously variable transmission mechanism.
By selectively engaging the friction element for path switching, the output of the engine is transmitted to the driving wheels using one of the first and second paths.

[0003] By controlling the speed ratio of the continuously variable transmission mechanism to a predetermined speed ratio in a state where the first path is selected, a gear neutral state in which output transmission to the drive wheel side is zero is obtained. At the same time, by increasing or decreasing the speed ratio of the continuously variable transmission mechanism from the predetermined speed ratio, a forward mode or a reverse mode in a low mode in which the final speed ratio of the power train is relatively low can be obtained. On the other hand, when the second path is selected, a high-mode forward state in which the final reduction ratio of the power train changes only according to the speed ratio of the continuously variable transmission mechanism in a region where the speed ratio is relatively high is obtained. It has become.

In any of the above-mentioned routes, the control of the final speed ratio as the power train is performed by controlling the speed ratio of the continuously variable transmission mechanism. In general, in the low mode, the speed ratio of the continuously variable transmission mechanism is controlled. As the speed increases (decreases to the low side), the reduction ratio of the power train decreases (increases to the high side). In the high mode, the speed ratio of the continuously variable transmission mechanism decreases (increases to the high side). ) Is controlled so that the reduction ratio of the power train becomes small (the speed is increased to the high side). For this reason, in the low mode and the high mode, even if the final speed ratio is controlled to the high side or the low side, the speed ratio of the continuously variable transmission mechanism changes in opposite directions, and as a result, Regardless of whether the mode is switched from the low mode to the high mode or from the high mode to the low mode, a predetermined mode switching point existing on the low side in the speed ratio of the stepped transmission mechanism is interposed, The transmission ratio is converted from a change to the low side to a change to the high side.

In this case, the characteristic representing the relationship between the change in the speed ratio of the continuously variable transmission mechanism in the low mode and the final gear ratio of the power train, and the same characteristic in the high mode are the same as those in the low speed mode. At the predetermined mode switching point existing on the side, at which point the same final gear ratio of the power train is obtained at the same gear ratio of the continuously variable transmission mechanism in both the low mode and the high mode. Therefore, theoretically, when the speed ratio of the continuously variable transmission mechanism reaches the speed ratio at the mode switching point, the final speed ratio is significantly increased before and after mode switching by operating the friction element for path switching. The fluctuation is suppressed, and the occurrence of the switching shock can be avoided.

However, actually, when the operation of the friction element is started only when the speed ratio of the continuously variable transmission mechanism reaches the mode switching point, the speed ratio of the continuously variable transmission mechanism is changed during the operation of the friction element. The point may further change beyond the point, the final gear ratio of the power train may deviate from the switching point, and a switching shock accompanying a gear ratio fluctuation may occur after mode switching. That is, when the gear ratio has passed the switching point by the time the path switching operation by the friction element has been completed, the power roller constituting the continuously variable transmission operates so as to eliminate the deviation due to the excessive ratio. As a result, the torque transmission amount fluctuates, and the driver feels this as a switching shock.

To cope with this, during the mode switching operation, the speed ratio of the continuously variable transmission mechanism may be maintained at the speed ratio at the mode switching point. In the transmission control device described in the above publication, The frictional element for achieving the low mode and the friction element for achieving the high mode are simultaneously set in the engaged state, and the above-described speed ratio is maintained. By bringing both friction elements into the fastened state at the same time,
Since the switching point where the characteristics of the low mode and the high mode match can be maintained, at the end of the path switching,
The occurrence of a shock when one friction element is disengaged is avoided.

[0008]

By the way, in the toroidal type continuously variable transmission disclosed in the above publication, a power roller interposed between an input disk having a toroidal surface and an output disk is generally inclined with respect to the two disks. The gear ratio control is performed by changing the angle. For example, from the predetermined neutral position where the power roller does not receive the rotational force of the disk and the tilt does not advance, the power roller is tilted by a predetermined angle in a predetermined direction by receiving the rotational force of the disk, The tilt angle of the power roller is changed by moving a support member supporting the roller with respect to the disk. Further, in the case where the movement amount control of the support member is achieved by controlling the hydraulic pressure supplied to the support member, and the hydraulic control is performed by controlling the sleeve of a three-layer valve for generating hydraulic pressure, By performing the feedback control of the sleeve position of the three-layer valve, it is possible to realize the speed ratio control of the continuously variable transmission mechanism and the control of the final speed ratio as the power train.

However, generally, in the vicinity of the mode switching point, the speed ratio of the continuously variable transmission mechanism is approaching the switching point, and the power roller continues to tilt from the high side to the low side. is there. Therefore, when the sleeve of the three-layer valve reaches the position corresponding to the switching point, or when the actual speed ratio reaches the speed ratio at the switching point, the sleeve is fixed to the position corresponding to the switching point. However, the tilting of the power roller does not stop at the switching point, but rather goes too low due to inertia. As a result, the mode switching operation is started at a point shifted from the original switching point. On the other hand, when the friction element on the engagement side is engaged, the gear ratio is forcibly moved to the theoretical value of the above-mentioned switching point, and a significant change in the gear ratio occurs, so that a switching shock occurs.

[0010] Particularly, in a transient operation state in which the torque input from the engine to the transmission changes, the torque change occurs when both of the friction elements for achieving the high mode and the low mode are simultaneously engaged. As a result, research has shown that a shock caused by this occurs. For example, when the input torque increases in a state where the engagement-side friction element and the release-side friction element are simultaneously engaged, the power roller starts to tilt in accordance with the increase in the input torque. Further, since the transmission of power in the low mode via the first path is fixed at a predetermined speed ratio, a shift of the speed ratio corresponding to the torque change is absorbed as a toroidal slip and a clutch slip. Therefore, there has been a problem that the occupant senses a change in the speed ratio caused by the instantaneous release after the end of the switching operation as a switching shock.

SUMMARY OF THE INVENTION The present invention has been made in view of the above points, and has a power train control device capable of effectively preventing a switching shock from occurring during a switching operation between a low mode and a high mode. It is intended to provide.

[0012]

The invention according to claim 1 is
A first path for transmitting the torque of the output source to the driven section via the continuously variable transmission mechanism and the gear mechanism, and a second path for transmitting the torque to the driven section via only the continuously variable transmission mechanism. In a power train having a transmission provided, target speed ratio setting means for setting a target speed ratio of the transmission based on a traveling state of a vehicle, and a target speed ratio set by the target speed ratio setting means are realized. Speed ratio control means for executing speed ratio control and path switching control of the continuously variable transmission mechanism, and torque adjustment for adjusting torque input from the output source to the transmission in accordance with accelerator operation. Means and the target speed ratio setting means sets a target speed ratio with the switching of the path, and during the execution of the switching control of the path, the fluctuation of the torque input to the transmission from the output source during the execution of the switching control of the path. Suppress That it is obtained by a torque variation suppressing means.

According to the above configuration, the target transmission ratio is set by the target transmission ratio setting means along with the switching of the path, and is input from the output source to the transmission when the switching control of the path is executed. By executing the control for suppressing the fluctuation of the torque by the torque fluctuation suppressing means, the occurrence of the switching shock due to the fluctuation of the input torque is prevented.

According to a second aspect of the present invention, in the power train control device according to the first aspect, the torque fluctuation suppressing means transmits the torque via at least both the first path and the second path. When the state is established, the system is configured to suppress the fluctuation of the torque input from the output source to the transmission.

According to the above configuration, when the friction element for the first path and the friction element for the second path are simultaneously engaged and the state of transmitting the torque via both paths is established, the output source is reduced. The control for suppressing the fluctuation of the torque input from the transmission to the transmission is executed by the torque fluctuation suppressing means, so that the occurrence of the switching shock due to the fluctuation of the input torque is effectively prevented.

According to a third aspect of the present invention, in the control apparatus for a power train according to the second aspect, the torque fluctuation suppressing means is configured such that a change in torque input to the transmission from the output source falls within an allowable range. Thus, the torque adjusting means is controlled, and the allowable range is set to a value corresponding to the operating state.

According to the above construction, the target speed ratio setting unit sets the target speed ratio accompanied by the switching by the target speed ratio setting means. When the switching control of the route is executed, the target speed ratio is input from the output source to the transmission. By executing the control for limiting the torque change within the allowable range by the torque fluctuation suppressing means, the occurrence of the switching shock due to the fluctuation of the input torque is prevented.

According to a fourth aspect of the present invention, in the power train control device according to the second or third aspect, the torque fluctuation suppressing means is configured to control the torque variation via both the first path and the second path. Before and after the transmission state, the torque control means is controlled so that the change in the torque input from the output source to the transmission falls within the first allowable range, and the transmission state of the torque via the two paths is determined. The torque input from the output source to the transmission is changed by the first
The torque adjusting means is controlled so as to be within a second allowable range which is set to be smaller than the allowable range.

According to the above configuration, when the control for switching the path is executed, the input from the output source to the transmission is made before the torque is transmitted via both the first path and the second path. The control for suppressing the change in the applied torque to be equal to or less than the first allowable value is executed by the torque fluctuation suppressing means, and the change in the input torque is further reduced at the time when the torque is transmitted through the two paths. When the suppression control is executed and the transmission state of the torque via both paths is released, the change in the input torque is suppressed by the torque change suppression unit so that the change in the input torque is allowed to some extent. Will be controlled.

[0020]

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS FIG. 1 schematically shows the overall structure of a power train 10 which comprises an input shaft 11 connected to an output shaft 2 of an engine 1 via a torsional damper 3. , This shaft 11
Hollow primary shaft 12 loosely fitted outside
And a secondary shaft 13 arranged in parallel with these shafts 11 and 12.
To 13 are arranged so as to extend in the width direction of the vehicle.

The toroidal first and second continuously variable transmission mechanisms 20 and 30 are provided on the axis of the input shaft 11 and the primary shaft 12, and an axial load is applied to them to transmit power. And a planetary gear mechanism 50, a low clutch 60 and a high clutch 70 are disposed on the axis of the secondary shaft 13, and a transmission is constituted by these. I have. Further, a low mode gear train 80 and a high mode gear train 90 are provided between the axis of the input shaft 11 and the primary shaft 12 and the axis of the secondary shaft 13.

The first and second continuously variable transmission mechanisms 20, 30
Have input disks 21 and 31 and output disks 22 and 32 each having a toroidal surface facing each other, and are provided between the disks 21 and 22 and between the disks 31 and 32. Is provided with two power rollers 23 and 33 for power transmission, respectively.

In the first continuously variable transmission mechanism 20 disposed farther from the engine 1, an input disk 21 is disposed on the opposite side of the engine 1 and an output disk 2
2 is arranged on the engine 1 side. Also, Engine 1
The input disk 31 is disposed on the engine 1 side, and the output disk 32 is disposed on the opposite side of the engine 1 of the second continuously variable transmission mechanism 30 disposed closer to the engine 1. Input disk 2 of both continuously variable transmission mechanisms 20 and 30
Reference numerals 1 and 31 are respectively connected to both ends of the primary shaft 12. The output disk 22,
32 is formed integrally and is rotatably supported at an intermediate portion of the primary shaft 12.

A first gear 81 constituting the low-mode gear train 80 is connected to one end of the input shaft 11, that is, an end located on the opposite side of the engine 1. The loading cam mechanism 40 is provided between the input disc 21 of the continuously variable transmission mechanism 20 and the input cam 21.
Is interposed. Further, a first gear 91 constituting a high mode gear train 90 is provided on the outer periphery of the output disks 22 and 32 integrated as described above.

On the other hand, a second gear 82 constituting the low mode gear train 80 is rotatably supported at one end of the secondary shaft 13 (an end located on the opposite side of the engine 1). Is connected to the first gear 81 via an idle gear 83, and a planetary gear mechanism 50 is disposed at an intermediate portion of the secondary shaft 13. And, the pinion carrier 51 of the planetary gear mechanism 50 and the second
A low clutch 60 for engaging or disengaging them is interposed between the gear 82 and the gear 82.

The engine 1 of the planetary gear mechanism 50
On the side, a second gear 92 forming a high mode gear train 90
Are rotatably supported, and the second gear 92 is
Output disks 22, 32 of the second continuously variable transmission mechanisms 20, 30
While being engaged with the first gear 91 provided on the
The second gear 92 is connected to the sun gear 52 of the planetary gear mechanism 50, and the internal gear 53 of the planetary gear mechanism 50 is connected to the secondary shaft 13. On the engine 1 side of the planetary gear mechanism 50, a high clutch 70 for fastening or disengaging the second gear 92 of the high mode gear train 90 and the secondary shaft 13 is provided.

Further, the other end of the secondary shaft 13 has first and second gears 4a and 4b and an idle gear 4a.
c, a differential unit 5 is connected via an output gear train 4 composed of a pair of left and right driven shafts 6a and 6b extending from the differential unit 5 to right and left ends of driven shafts (not shown) composed of left and right driving wheels. Are connected.

An oil pump 100 is disposed at the other end of the input shaft 11, and the driving force of the input shaft 11 is applied to the first mode of the low mode gear train 80.
The oil pump 100 is driven by being transmitted to the oil pump 100 via the gear 81.

Next, the first and second continuously variable transmission mechanisms 20,
The configuration of 30 will be described in more detail by taking the first continuously variable transmission mechanism 20 as an example. The pair of power rollers 23, 23
As shown in FIG. 3, the input / output disks 21 and 22 are supported by trunnions 25 and 25 via shafts 24 and 24 extending substantially in the radial direction of the input / output disks 21 and 22, respectively.
The output disks 21 and 22 are arranged vertically and parallel in a substantially horizontal posture on the 180 ° opposite side on the circumference of the toroidal surfaces opposed to each other, and the two disks 21 and 22 are located at two places on the opposite side of the peripheral surface by 180 °. 22 are in contact with the toroidal surfaces, respectively.

The trunnions 25, 25 are disposed between the left and right support members 26, 26 attached to the case 101 of the power train 10, and the power rollers 2 in the tangential direction of the disks 21, 22.
The shafts 3 and 23 are supported so as to be capable of rotating around a horizontal axis X, X orthogonal to the shafts 24, 24 and reciprocating linearly in the X, X directions. Rods 27, 27 extending to one side along the axes X, X are connected to the trunnions 25, 25, respectively.
These rods 27, 2 are provided on the side of the case 101.
A transmission control unit 110 for tilting the power rollers 23 via the trunnions 25 and the trunnions 25 is mounted.

The transmission control unit 110 has a hydraulic control unit 111 and a trunnion drive unit 112. The trunnion drive unit 112 has upper and lower trunnions 25, 25.
The pistons 113 and 114 for speed increase and deceleration attached to the rods 27 and 27 are respectively opposed to each other. The pistons 113 and 114 form the hydraulic chamber 115 for speed increase and the hydraulic chamber 115 for deceleration. Each is formed.

As for the trunnion 25 located above, the speed increasing hydraulic chamber 115 is disposed on the power roller 23 side, and the deceleration hydraulic chamber 116 is disposed on the opposite side of the power roller 23. As for the trunnion 25 located below, the speed-increasing hydraulic chamber 115 is set on the opposite side of the power roller 23, and the deceleration hydraulic chamber 116 is disposed on the power roller 23 side.

The speed-increasing hydraulic pressure PH generated by the hydraulic pressure control unit 111 is transmitted through oil passages 117 and 118 to the speed-increasing hydraulic chambers 115 and 115 of the upper and lower trunnions 25 and 25.
The deceleration hydraulic pressure PL, also generated by the hydraulic control unit 111, is supplied to deceleration hydraulic chambers 116, 116 of the upper and lower trunnions 25, 25 via an oil passage (not shown). The gear ratio of the first and second continuously variable transmission mechanisms 20 and 30 is controlled by controlling the PL.

The specific operation of the speed ratio control for the first continuously variable transmission mechanism 20 will be described below. The speed increasing hydraulic pressure PH supplied to the speed increasing hydraulic chambers 115, 115 of the upper and lower trunnions 25, 25 by the hydraulic control unit 111 shown in FIG.
Hydraulic pressure P for deceleration supplied to hydraulic chambers 116 for deceleration
When the height is relatively higher than L, the upper trunnion 25 horizontally moves to the right and the lower trunnion 25 horizontally moves to the left in FIG.

At this time, the output disk 2 shown in FIG.
2 is rotated clockwise (direction c) in FIG. 3, the upper power roller 23 receives a downward force from the output disk 22 due to the rightward movement, and the upper power roller 23 is positioned in front of the drawing. Input disk 2 rotating counterclockwise
1 will receive an upward force. The lower power roller 23 receives an upward force from the output disk 22 by moving to the left, and
1 will receive a downward force. As a result, the upper and lower power rollers 23 are tilted so that the contact position with the input disk 21 moves outward in the radial direction and the contact position with the output disk 22 moves inward in the radial direction. As a result, the speed ratio of the first continuously variable transmission mechanism 20 is reduced, and the speed is increased.

Conversely, the deceleration hydraulic pressure PL supplied to the deceleration hydraulic chambers 116, 116 of the upper and lower trunnions 25, 25
Is relatively higher than the speed increasing hydraulic pressure PH supplied to the speed increasing hydraulic chambers 115, 115, the upper trunnion 2
5 moves horizontally to the left on the drawing, and the lower trunnions 25 move horizontally to the right, respectively.
The upper power roller 23 receives an upward force from the output disk 22 and a downward force from the input disk 21. The lower power roller 23 receives a downward force from the output disk 22 and an upward force from the input disk 21. As a result, the upper and lower power rollers 23, 23
While the contact position with 1 moves inward in the radial direction,
By tilting so that the contact position with the output disk 22 moves to the outside in the radial direction, the speed ratio of the first continuously variable transmission mechanism 20 is increased, and the speed is reduced.

The structure and operation of the first continuously variable transmission mechanism 20 are the same as those of the second continuously variable transmission mechanism 30. Then, as shown in FIGS. 1 and 2, the hollow primary shaft 12 loosely fitted to the input shaft 11.
The input disks 21 and 31 of the first and second continuously variable transmission mechanisms 20 and 30 are spline-fitted to both ends, respectively, so that these input disks 21 and 31 always rotate in the same manner. . Further, as described above, since the output disks 22 and 32 of the both continuously variable transmission mechanisms 20 and 30 are integrated, the rotational speeds on the output side of the both continuously variable transmission mechanisms 20 and 30 are always the same. Therefore, the first and second hydraulic control of the power rollers 23 and 33 as described above is performed.
The speed ratio control of the continuously variable transmission mechanisms 20 and 30 is also executed such that the speed ratio is always kept the same.

Next, a description will be given of a hydraulic control circuit of the power train 10 constituted by the shift control unit 110 and a clutch control unit 120 (see FIG. 3) attached to a lower portion of the case 101.

As shown in FIG.
The hydraulic pressure control circuit 200 includes a regulator valve 202 that adjusts the pressure of the hydraulic oil discharged from the oil pump 100 to a predetermined line pressure and outputs the adjusted line pressure to the main line 201.
And a relief valve 204 for generating a predetermined relief pressure using the line pressure supplied from the main line 201 as an original pressure and outputting the relief pressure to a relief pressure line 203, and a D range, an R range, A manual valve 205 that allows selection between the N range and the P range is provided.

Of the above valves, the manual valve 20
5 communicates the main line 201 with the first and second output lines 206 and 207 in the D range, and the first and third output lines 206 and 2 in the R range.
08 respectively, and operates so as to cut off the line pressure in the N range and the P range.

The hydraulic circuit communicating with the regulator valve 202 and the relief valve 204 includes a linear solenoid valve 209 for controlling line pressure and a linear solenoid valve 210 for controlling relief pressure, and the discharge pressure of the pump 100 as a source pressure. A reducing valve 211 that generates a constant pressure is provided. Based on the constant pressure generated by the reducing valve 211, the linear solenoid valves 209 and 210 generate control pressures. .

The control pressure is supplied to the control ports 202a and 204a of the regulator valve 202 and the relief valve 204, whereby the line pressure and the relief pressure are adjusted. That is, the regulator valve 20 is controlled in accordance with the control signals output to the linear solenoid valves 209 and 210.
2 and the relief pressure output from the relief valve 204 to the relief pressure line 203 are respectively adjusted.

Further, the reducing valve 211
The constant pressure generated in step (1) is also guided to an on / off solenoid valve 213 that operates the fail-safe valve 212. The on / off solenoid valve 213 is normally turned on to reduce the above-mentioned constant pressure to the fail-safe valve 2.
By supplying the control port 212a to the twelve control ports 212a, the spool of the fail-safe valve 212 is moved to the right. In a fail-safe state, the on / off solenoid valve 213 is turned off to drain the constant pressure off from the control port 212a of the fail-safe valve 212, thereby moving the spool of the fail-safe valve 212 to the left. It has become.

The hydraulic pressure control circuit 200 receives a speed-increasing hydraulic pressure PH for speed change control based on the line pressure and the relief pressure at the time of forward movement and at the time of reverse movement, respectively.
Three-layer valve 220 for forward and reverse three-layer valve 230 for generating hydraulic pressure PL for deceleration, and these three-layer valves 220 and 2
A shift valve 240 for selectively operating the shift valve 30 is provided.

The position of the spool of the shift valve 240 is determined by whether or not the line pressure is supplied to the control port 240a at one end. When the line pressure is not supplied, the spool is located on the right side. By doing
Line pressure supply line 241 leading to forward three-layer valve 220
The main line 201 communicates with the main line 201. When the line pressure is supplied, the spool of the shift valve 240 is located on the left side, so that the main line 201 communicates with the line pressure supply line 242 communicating with the three-way reversing valve 230. I have.

Here, the line pressure is supplied to the control port 240a of the shift valve 240 when the spool moves to the right in normal times, that is, via the fail-safe valve 212 and the third output line 208. , The manual valve 205 is located in the R range. On the other hand, even when the spool of the fail-safe valve 212 normally moves to the right, the line pressure is not supplied to the control port 240a of the shift valve 240 when the manual valve 205 is located in the D range. Further, at the time of fail-safe, the spool of the fail-safe valve 212 moves to the left, and the shift valve 240 and the third output line 208 are shut off. No line pressure is supplied to the control port 240a of 240.

The forward and backward three-layer valves 220 and 2
Numeral 30 has the same configuration, and sleeves 222 and 232 are fitted movably in the axial direction of the bores 221 and 231, and spools 223 and 233 are fitted movably in the axial direction of the sleeves 222 and 232, respectively. Both are housed in the valve body 111a of the hydraulic control unit 111 in the shift control unit 110 shown in FIG.

A line pressure port 2 connected to line pressure supply lines 241 and 242 communicating with the shift valve 240 is provided at the center of the two three-layer valves 220 and 230.
24, 234 are formed, and first and second relief pressure ports 225, 226, 235, 236 to which the above-described relief pressure lines 203 are connected are formed at both ends. Further, the line pressure ports 224, 234
And between the first relief pressure ports 225 and 235,
The pressure increasing ports 227 and 237 are provided, and the line pressure ports 224 and 234 and the second relief pressure port 22 are provided.
6, 236, deceleration pressure ports 228, 238 are provided, respectively.

The forward and backward three-layer valve 2
Lines 243 and 244 communicating with the speed increasing pressure ports 227 and 237 of the valves 20 and 230, and lines 245 and 246 communicating with the deceleration ports 228 and 238 of the three-way valves 220 and 230 for forward and backward respectively. It is connected to the shift valve 240. This shift valve 2
When the spool 40 is located on the right side, the speed-up pressure port 227 and the deceleration pressure port 22 of the three-way forward valve 220
Lines 243 and 245 communicating with No. 8 are speed-up lines 2
47, and the deceleration hydraulic chambers 115, 115 and the deceleration hydraulic chambers 116, 116 via the deceleration line 248.
Respectively.

When the spool of the shift valve 240 is located on the left side, the lines 244 and 246 communicating with the speed increasing pressure port 237 and the deceleration pressure port 238 of the three-way reversing valve 230 are connected to the speed increasing line 247 and Through the deceleration line 248, the speed-increasing hydraulic chamber 115,
115 and deceleration hydraulic chambers 116, 116, respectively.

The operation of the two three-layer valves 220 and 230 will now be described with reference to FIG. In FIG. 5, the directions of the two three-layer valves 220 and 230 are opposite to those in FIG. For example, when the sleeve 222 of the forward three-layer valve 220 moves relatively to the left (in the direction of arrow g) in the drawing from the illustrated neutral position, the communication between the line pressure port 224 and the pressure increasing pressure port 227 and The degree of communication between the second relief pressure port 226 and the deceleration pressure port 228 increases.

Conversely, when the sleeve 222 relatively moves to the right (in the direction of the arrow h), the communication between the line pressure port 224 and the deceleration pressure port 228 and the first relief pressure port 225 and the speed-increasing pressure port 227 Respectively increase. Therefore, in the former case, the speed increasing hydraulic pressure PH
Increases, and the deceleration hydraulic pressure PL decreases. In the latter case, the deceleration hydraulic pressure PL increases and the speed-increasing hydraulic pressure PH decreases.

The above operation is effected by the three-way valve 230 for retraction.
Similarly, step motors 251 and 252 for operating the sleeves 222 and 232 of the three-way valves 220 and 230 for forward and backward movement are provided, and the motors for forward and backward movement are provided via link members 253 and 254, respectively. It is connected to the sleeves 222, 232 of the three-layer valves 220, 230.

The hydraulic control unit 110 includes:
The sleeve 2 is driven by step motors 251 and 252.
22 and 232, the spool 22
There is provided a cam mechanism 260 that moves the shaft 3,233 in the axial direction against the spring force of the springs 229,239.

The cam mechanism 260 is attached to the end of the rod 37 of the trunnion 35 located above the second continuously variable transmission mechanism 30, as shown in FIGS. The cam mechanism 260 includes a precess cam 261 having a spiral cam surface 261a formed on one end surface, and spools 223, 2 of three-way valves 220, 230 for forward and backward movement.
A shaft 262 that is disposed at one end of the shaft 33 in a direction perpendicular to these shafts and is rotatably supported by the valve body 111a of the hydraulic control unit 111;
2 and a driven lever 263 having a swinging end abutting on a cam surface 261a of the precess cam 261; and a driven lever 263 attached to the shaft 262 and having a swinging end having the forward and backward three-layer valve 220. ,
A forward drive lever 264 and a reverse drive lever 265 engaged with cuts 223a, 233a provided at one end of spools 223, 233 of 230 are provided.

When the upper power roller 33 in the second continuously variable transmission mechanism 30 is tilted by the control of the speed-increasing hydraulic pressure PH and the deceleration hydraulic pressure PL, the trunnion 35 and the rod When the 37 rotates integrally around the axis X, the precess cam 261 also rotates integrally therewith. In accordance with the rotation of the precess cam 261, the driven lever 263, whose swing end abuts on the cam surface 261 a, swings by a predetermined amount, and drives the forward and backward drive levers 264 and 265 via the shaft 262. Is also displaced by the same angle, so that the spools 22 of the forward and backward three-layer valves 220 and 230 are moved by an amount corresponding to the swing angle.
3,233 will move in the axial direction.

Therefore, the spools 223, 233
Is the tilt angle of the power roller 33 provided in the second continuously variable transmission mechanism 30 and the power roller 23 provided in the first continuously variable transmission mechanism 20, in other words, the continuously variable transmission mechanism 2
This corresponds to a gear ratio of 0,30.

Here, the first and second continuously variable transmission mechanisms 2
The control operation of the gear ratio (toroidal ratio Rt) of 0, 30 will be described by taking a forward movement as an example. First, the control pressure of the regulator valve 202 and the relief valve 204 is generated by the line pressure control linear solenoid valve 209 and the relief pressure control linear solenoid valve 210 in the hydraulic control circuit 200, and the line pressure corresponding to the control pressure is generated. And a relief pressure are generated.

The line pressure is supplied from the main line 201 to the line pressure port 224 of the forward three-layer valve 220 via the shift valve 240 and the line 241. The relief pressure is supplied to the first and second relief pressure ports 2 of the forward three-layer valve 220 through the relief pressure line 203.
25, 226. Then, based on the line pressure and the relief pressure, the sleeve position of the forward three-layer valve 220 is controlled by the step motor 251, so that the speed increasing hydraulic chamber 11 of the shift control unit 110 is controlled.
5, 115 and the pressure difference ΔP (= PH−PL) between the speed-increasing hydraulic pressure PH and the deceleration hydraulic pressure PL supplied to the deceleration hydraulic chambers 116, 116, respectively.

By executing this differential pressure control, as shown in FIG. 6, the power rollers 23, 33 are not rotated by the respective discs 21, 22, 31, 32, and the predetermined tilting does not proceed. In the neutral position, the two continuously variable transmission mechanisms 2
0,30 trunnions 25,35 or power rollers 2
3, 33 are held respectively. The trunnions 25, 35 or the power rollers 23, 33 move along the axis X, X direction from the neutral position, so that the power rollers 23, 33
Receiving the rotation of 32, the tilting proceeds (the toroidal ratio changes).

Now, for example, the input torque from the engine 1 is applied to the input disks 21, 30 of the continuously variable transmission mechanisms 20, 30.
When the power is transmitted from the side 31 to the output disks 22 and 32, the power rollers 23 and 33 are driven in the directions b and b by the rotation of the input disks 21 and 31 in the directions c and c. 23 and 33 and the trunnions 25 and 35 that support them,
A force acts to move the rotation directions 1 and 31 in the same direction as the rotation directions a and a.

Further, b of the power rollers 23 and 33,
Since the output disks 22 and 32 are driven in the directions c and c by the rotation in the direction b, the output disk 2
The forces in the directions opposite to the rotation directions c, c of the power rollers 2, 32 act on the power rollers 23, 33 or the trunnions 25, 35. As a result, the traction forces T1, T1 in the direction approaching the trunnion drive unit 112 act on the power rollers 23, 33 and the trunnions 25, 35.

Conversely, for example, the continuously variable transmission mechanisms 20 and 30
Input disks 21 and 3 from the output disks 22 and 32
When the input torque is transmitted to the first side, the rotation of the output disks 22 and 32 in the c and c directions causes the power roller 2 to rotate.
Since the power rollers 3 and 33 are driven in the directions b and b, the power rollers 23 and 33 and the trunnions 25 and
A force acts on 35 to move them in the same direction as the rotation directions c, c of the output disks 22, 32.

Further, b of the power rollers 23 and 33,
The input disks 21 and 31 are driven in the directions a and a by the rotation in the direction b.
The forces in the directions opposite to the rotational directions a, a of the power rollers 1, 31 act on the power rollers 23, 33 or the trunnions 25, 35. As a result, traction forces T2 and T2 in a direction away from the trunnion drive unit 112 act on the power rollers 23 and 33 and the trunnions 25 and 35.

In order to hold the power rollers 23, 33 at the neutral position against the traction forces T1, T2,
The speed increasing hydraulic chamber 115 and the speed reducing hydraulic chamber 116 provided in each of the trunnions 25 and 35 are provided with a speed increasing hydraulic pressure PH and a speed reducing hydraulic PL so that the differential pressure ΔP is equal to the traction force T. May be supplied respectively.

When the toroidal ratio is reduced from the neutral state, for example, in order to increase the speed, the sleeve 222 of the forward three-layer valve 220 is moved by the stepping motor 251 to the left (g direction) in FIGS. To increase the degree of communication between the line pressure port 224 and the speed-up pressure port 227 and the degree of communication between the second relief pressure port 226 and the deceleration pressure port 228 of the three-way valve 220 for forward movement. I do.

As described above, the speed increasing pressure line 24 shown in FIG.
7, the pressure-increasing hydraulic pressure PH supplied to the pressure-increasing hydraulic chambers 115, 115 is increased, and the deceleration pressure line 24
When the deceleration hydraulic pressure PL supplied from 8 to the deceleration hydraulic chambers 116, 116 is reduced, the differential pressure ΔP is increased. As a result, the trunnions 25, 35 or the power rollers 23, 33 are shown in FIG. It will move in the d1 and d1 directions.

The movement of the trunnions 25 and 35 through the power rollers 23 and 33 causes the power roller 2 to move.
3 and 33 are tilted in directions in which the contact positions with the input disks 21 and 31 are displaced radially outward and the contact positions with the output disks 22 and 32 are displaced radially inward, respectively.
The speed of the first and second continuously variable transmission mechanisms 20 and 30 is increased, and the toroidal ratio is reduced.

Further, as the power roller 33 of the second continuously variable transmission mechanism 30 is tilted as described above, the precess cam 261 of the cam mechanism 260 rotates by the same angle in the same direction (direction e shown in FIG. 5). Accordingly, the driven lever 263 and the shaft 2 of the cam mechanism 260
Both 62 and the drive lever 264 rotate in the f direction shown in FIG.

The shaft 262 and the drive lever 26
4, the spool 223 of the forward three-layer valve 220
Moves in the g direction, that is, the left direction in FIGS. 5 and 6 due to the spring force of the spring 229. Since this direction is the direction in which the sleeve 222 is moved by the step motor 251 as described above, Once increased the communication between the line pressure port 224 and the speed-up pressure port 227 and the second relief pressure port 226 and the deceleration pressure port 2
28 will return to the initial neutral state.

As a result, the differential pressure ΔP is reduced again, and the speed change operation is completed. After the speed ratio of the continuously variable transmission mechanisms 20, 30, ie, the toroidal ratio has changed by a predetermined amount, the power rollers 23, 33 is returned to the neutral position again and held.

The above shifting operation ends when the spool 223 of the three-way forward valve 220 moves to a position where the spool 223 becomes a predetermined neutral state in relation to the sleeve 222. The position is determined by the step motor 251. Is the position where the sleeve 222 has been moved by the
60, the positions are associated with the tilt angles of the power rollers 23, 33 and the trunnions 25, 35 via
The position of the sleeve 222 corresponds to the tilt angle of the power rollers 23, 33 and the trunnions 25, 35. as a result,
The control amount of the step motor 251 corresponds to the speed ratio of the first and second continuously variable transmission mechanisms 20 and 30, and the pulse control of the step motor 251 controls the toroidal ratio.

The above operation is performed by the step motor 251.
5, the sleeve 222 of the three-way valve 220 for advance
The same applies to the case of moving to the right (h direction) in the opposite direction in FIG. 6, in which case the trunnions 25,
35 to the power rollers 23 and 33 correspond to d shown in FIG.
By moving in the 2 and d2 directions, the toroidal ratio increases and the vehicle is decelerated.

On the other hand, as shown in FIG. 4, the hydraulic control circuit 200 includes two duty solenoids for controlling the low clutch 60 and the high clutch 70 in addition to the above-described structure for controlling the gear ratio. Vanoleb 271,
272 is provided, and the manual valve 205 is provided.
A first output line 206 derived from the second output line 206 is connected to a low clutch duty solenoid valve 271.
07 is the high clutch duty solenoid valve 27
2 respectively.

Then, when the vehicle is running in the low mode, the duty solenoid valve 271 for the low clutch operates
The line pressure from the first output line 206 is adjusted to generate the engagement pressure (low clutch pressure) of the low clutch 60. Under normal conditions, this low clutch pressure becomes low through the fail-safe valve 212 and the low clutch line 274. When the low clutch 60 is supplied to the hydraulic chamber of the clutch 60, the low clutch 60 is engaged with an engagement force according to the size. As a result, the torque input from the engine 1 is transmitted to the driven portion via the first path that passes through both the continuously variable transmission mechanisms 20, 30 and the planetary gear mechanism 50.

When the vehicle is running in the high mode, the line pressure from the second output line 207 is adjusted by the operation of the high clutch duty solenoid valve 272 to generate the engagement pressure of the high clutch 70 (high clutch pressure). The high clutch pressure is applied to the high clutch line 2
Is supplied to the hydraulic chamber of the high clutch 70 via the high clutch 75 with a fastening force corresponding to the size thereof.
0 is concluded. As a result, the torque input from the engine 1 is transmitted to the driven portion via the second path that passes only through the continuously variable transmission mechanisms 20 and 30.

The duty solenoid valve 271,
272 outputs no clutch pressure when the duty ratio of the control signal is 0% (fully closed), and outputs the supplied line pressure as it is as the clutch pressure when it is 100% (fully open). At the intermediate duty ratio, a clutch pressure corresponding to the duty ratio is generated.

The low clutch line 274 and the high clutch line 275 are provided with accumulators 276 and 277 for gently supplying the engagement pressure to the low clutch 60 and the high clutch 70, respectively. , 70 at the time of fastening is suppressed.

Further, the third output line 208 communicating with the manual valve 205 is connected to the control port 240a of the shift valve 240 via the fail-safe valve 212 at the time of normal operation as described above. When the line pressure is shifted to the position of the R range, the line pressure becomes lower than the control port 2 of the shift valve 240.
The spool is supplied to the shift valve 40a to move the spool of the shift valve 240 to the left side, that is, a position for retreating.

Further, at the time of fail-safe, etc., the on / off solenoid valve 213 for operating the fail-safe valve 212 is turned off, and the spool of the fail-safe valve 212 moves to the left. 271 and the low clutch line 274, and between the third output line 208 and the shift valve 240, respectively. At this time, especially the low clutch 6
The low clutch line 274 communicating with the zero hydraulic chamber is connected to the drain port 212b of the fail-safe valve 212, and the low clutch pressure is quickly discharged from the drain port 212b.

The hydraulic control circuit 200 shown in FIG. 4 is provided with a lubrication line 281 led from the drain port of the regulator valve 202 in addition to the above configuration. Then, a relief valve 282 for adjusting the lubricating oil pressure to a predetermined value and a first
The second opening / closing valves 283, 284, etc. are arranged, and the first and second opening / closing valves 283, 284, etc.
The supply of lubricating oil to each part of the power train such as the second continuously variable transmission mechanisms 20 and 30 and the planetary gear mechanism 50 is controlled.

The power train 10 according to this embodiment
Has the above-described mechanical configuration and a hydraulic control circuit 200, and uses this hydraulic control circuit 200 to control the speed ratio of the first and second continuously variable transmission mechanisms 20 and 30 and to operate the clutch 6.
A control unit is provided for controlling the gear ratio (unit ratio Ru) of the entire power train 10 by performing the engagement control of 0, 70.

As shown in FIG. 7, the control unit 300 includes a vehicle speed sensor 30 for detecting the running speed of the vehicle.
1. An engine speed sensor 302 for detecting the number of revolutions of the engine 1, a throttle opening sensor 303 for detecting a throttle opening, a selection range sensor 304 for detecting a range selected by a driver, and detecting an operation amount of an accelerator. 305, an oil temperature sensor 306 for detecting the temperature of the hydraulic oil, an input speed sensor 307 and an output speed sensor 308 for detecting the speed of the input disks 21, 31 and the output disks 22, 32 (see FIG. 1). And the like are input.

Then, the control unit 300
A target gear ratio setting means 311 for setting a target gear ratio corresponding to the running state of the vehicle based on the detection signals input from the sensors 301 to 308; and Linear solenoid valves 209 and 210 for line pressure control and relief pressure control, on / off solenoid valve 213, duty solenoid valve 2 for low clutch 60 and high clutch 70
By outputting control signals to the step motors 251 and 252 for driving the sleeves 222 and 232 of the three-layer forward valve 220 and the reverse three-layer valve 230, the speed of the continuously variable transmission mechanisms 20 and 30 is changed. Gear ratio control means 312 for performing ratio control and switching control of a path through which torque is transmitted is provided.

Further, the control unit 300
When the target gear ratio setting unit 311 sets the gear ratio with the above-mentioned path switching, the output adjusting unit including the electric throttle 309 is controlled during the execution of the path switching control. A torque fluctuation suppressing unit 313 for suppressing a change in output transmitted from the engine 1 to the transmission is provided. The above electric throttle 30
Numeral 9 is configured to adjust the torque input from the engine 1 to the transmission by electrically controlling the amount of intake air supplied to the combustion chamber of the engine 1 in response to a driver's accelerator operation at normal times. Have been.

The speed ratio control of the continuously variable transmission mechanisms 20, 30 executed by the speed ratio control means 312 will be described below. In the power train 10 equipped with the above-described continuously variable transmission mechanisms 20, 30, both the low clutch 60 and the high clutch 70 are disengaged when the N range is selected. Therefore, the power transmitted from the input shaft 11 to the secondary shaft 13 is not transmitted to the planetary gear mechanism 50 and the secondary shaft 13, and therefore, the power is not output from the differential device 5 to the driving wheels. Absent.

At this time, in the planetary gear mechanism 50, the sun gear 5 is driven by the power from the high mode gear train 90.
2 is driven, but the power from the low mode gear train 80 is supplied to the rotation member 60a on the input side of the low clutch 60 (FIG. 1).
), But not to the pinion carrier 51. Further, since the internal gear 53 is fixed to the secondary shaft 13, the pinion carrier 51 is rotating in a no-load state in conjunction with the sun gear 52.

By setting the toroidal ratio to a predetermined value in the above state, the pinion carrier 5
1 can be controlled to a speed at which the rotation speeds of the input / output rotation members 60a and 60b (see FIG. 1) of the low clutch 60 become equal. In other words, by controlling the toroidal ratio to the predetermined value, even when the low clutch 60 is connected, the rotation of the internal gear 53 or the secondary shaft 13 can be made zero. This allows the so-called geared neutral (G
The state of N) is obtained.

Here, the step motors 251, 25
The relationship between the number of pulses (N) of the control signal output to 2 and the toroidal ratio (Rt) has characteristics as shown in FIG. 8, for example. That is, when the number of pulses (N) increases (changes to the positive side), the toroidal ratio (R
t) becomes smaller (changes to the speed increasing side). At this time, the sleeves 222 and 232 of the three-layer valves 220 and 230 are used.
Moves in the direction of arrow g shown in FIGS. 5 and 6 as described above. The case where the sleeves 222 and 232 move in the direction g away from the pulse motors 251 and 252 is defined as a plus side.

Conversely, the step motors 251 and 252
When the number of pulses (N) of the control signal output to (1) decreases (changes to the negative side), the toroidal ratio (Rt) increases (changes to the deceleration side). At this time, the sleeves 222 and 232 of the two-layer valves 220 and 230 move in the direction of the arrow h shown in FIGS. 5 and 6 as described above. Note that the sleeves 222 and 232 are
The movement in the direction h approaching 1,252 is defined as a minus side.

When the geared neutral state is obtained, the toroidal ratio (GN ratio Rtn) is smaller than 1, and the number of pulses (GN pulse number Nn) for realizing the GN ratio Rtn is relatively on the positive side. On the other hand, the relationship between the number of pulses (N) of the control signal output to the step motors 251 and 252 and the unit ratio (Ru) has characteristics as shown in FIG. 9, for example.

That is, when the number of pulses (N) is equal to the number of GN pulses (Nn), the unit ratio (Ru)
Becomes infinity as shown by the code a or the code a. When the pulse number (N) decreases (changes to the minus side) from the GN pulse number (Nn) and the toroidal ratio (Rt) increases (changes to the deceleration side), the input rotation speed to the sun gear 52 decreases. As a result, the planetary gear mechanism 50
Starts rotating in the forward direction. That is, as the pulse number (N) decreases, the sleeve position moves to the negative side, and the toroidal ratio (Rt) increases, the unit ratio decreases (Ru) (changes to the speed increasing side), so that the vehicle moves forward. Low mode characteristics LF are realized.

Conversely, when the pulse number (N) increases (changes to the plus side) from the GN pulse number (Nn) and the toroidal ratio (Rt) decreases (changes to the speed increasing side), the sun gear 52 is moved to the sun gear 52. Is increased, the internal gear 53 of the planetary gear mechanism 50 is increased.
Begins to rotate in the reverse direction. That is, the number of pulses (N)
The unit ratio (Ru) increases (changes to the deceleration side) as the sleeve position increases, the sleeve position moves to the positive side, and the toroidal ratio (Rt) decreases. Is realized.

Further, after the vehicle starts in the forward low mode LF, the unit ratio (Ru) is reduced by decreasing the number of pulses (N), and a predetermined switching is performed as shown by a symbol c in FIGS. When the point (number of pulses Nm, toroidal ratio Rtm, unit ratio Rum) is reached, the low clutch 60 is disengaged and the high clutch 70 is engaged, so that the clutch 6
Replacement of 0 and 70 is performed. As a result, the power from the input shaft 11 is transmitted to the secondary shaft 13 via the first and second continuously variable transmission mechanisms 20 and 30, the high mode gear train 90 and the high clutch 70.

As shown in FIG. 10, the mode switching point (c) is the same toroidal ratio (Rt) in both the low mode LF and the high mode HF.
This is the only point where the same unit ratio (Rum) is obtained in m). Therefore, by switching the mode at the above point (c), there is no significant change in the unit ratio before and after the switching, and a smooth mode switching without switching shock is realized.

The switching point pulse number (N) for realizing the switching point toroidal ratio (Rtm)
Various physical quantities such as m: when torque is zero) or a switching point sleeve position (Sm: when torque is zero) are theoretically determined to correspond to only one point. Note that the switching point pulse number (Nm) or the switching point sleeve position (Sm) for realizing the switching point toroidal ratio (Rtm) changes according to the torque.

Therefore, in theory, the actual toroidal ratio (Rt) is equal to the switching point toroidal ratio (Rtm).
When the pulse number or the sleeve position reaches the switching point pulse number (Nm) or the switching point sleeve position (Sm) and the control of the pulse number or the sleeve position is stopped, the actual toroidal ratio is increased. It becomes stable at the switching point toroidal ratio (Rtm).

Next, a specific control operation at the time of switching between the forward low mode LF and the forward high mode HF executed by the gear ratio control means 312 will be described. This control is basically control in the vicinity of the mode switching point (c), and the switching between the low clutch 60 and the high clutch 70 as friction elements for achieving each mode is performed.

That is, when the mode is switched from the low mode LF to the high mode HF, the low clutch 60 is disengaged and the high clutch 70 is engaged. Conversely, high mode H
When switching from F to the low mode LF, the high clutch 70 is disengaged and the low clutch 60 is engaged. When the actual toroidal ratio is changed to the switching point toroidal ratio (Rt
m), and during the switching operation, the number of pulses to the forward stepping motor 251 is controlled so that the actual toroidal ratio is kept constant at the switching point toroidal ratio (Rtm).

On the other hand, during normal times other than such mode switching, basically, as shown in FIGS. 11 to 13, in each of the forward low mode LF, the forward high mode HF, and the reverse low mode LR, Feedback control of a gear ratio (a toroidal ratio and a unit ratio) is performed based on a shift diagram that is set in advance using running conditions of the vehicle such as a vehicle speed (V) and a throttle opening (TVO) as parameters.

In the speed ratio control, as shown in FIG. 12, the target engine speed (Neo) is first determined by applying the actual vehicle speed (V) and the actual throttle opening (TVO) to the speed diagram. Then, a target unit ratio (Ruo) is calculated from the target engine speed (Neo) and the actual vehicle speed (V), and a target toroidal ratio (Rto) at which the target unit ratio (Ruo) is obtained.
Then, the toroidal ratio (Rto) is feedback-controlled by executing pulse control (sleeve position control) on the step motors 251 and 252 so that the target toroidal ratio (Rto) is realized. In each of the shift diagrams, a mode switching line (M) having a slope of the mode switching point unit ratio (Rum) is shown.

Next, the control operation at the time of mode switching will be described with reference to the time chart of FIG. Note that this time chart shows a case where the vehicle speed gradually increases due to, for example, continued depression of the accelerator pedal in the normal driving state, and switching from the low mode LF to the high mode HF occurs. During the period up to the time point t1 and the period after the time point t4 shown in FIG. 14, feedback (F / B) of the normal gear ratio based on the traveling state and the shift diagram is performed. t4
During this period, an area in which mode switching control is performed is shown.

During the period up to the time point t1, the low mode LF is attained, and the duty ratio for the low clutch duty duty solenoid valve 271 is set to 10 while the on / off solenoid valve 213 is turned on.
The duty ratio for the high clutch duty solenoid valve 272 is set to 0%. As a result, the line pressure is directly supplied to the hydraulic chamber of the low clutch 60 as the low clutch pressure (EL),
While this low clutch 60 is in a completely engaged state,
The high clutch pressure (EH) is applied to the hydraulic chamber of the high clutch 70.
Is not supplied, and the high clutch 70 is completely released.

During the period up to the time point t1, the target unit ratio changes to the speed increasing side by the feedback control of the gear ratio. Therefore, the target toroidal ratio changes to the deceleration side, and the toroidal ratio (Rt) is feedback-controlled so that the target toroidal ratio is realized. As a result, the actual toroidal ratio changes to the deceleration side so as to follow the target toroidal ratio. I do. At this time, the number of pulses or the sleeve position of the three-way forward valve 220 changes to the minus side.

At time t1 when the actual toroidal ratio reaches the switching point toroidal ratio (Rtm), the feedback control of the speed ratio is stopped, and the number of pulses or the sleeve position is changed to the predetermined number of pulses and the predetermined number of pulses at the time t1. Fixed in the sleeve position. The number of pulses is fixed for a predetermined time T from time t1.
This is continued until a has elapsed.

The actual toroidal ratio is stably fixed at the switching point toroidal ratio at least until the time point t4 at which the normal mode ratio feedback control is resumed after the mode switching is completed. Next, pulse control for the forward step motor 251 is performed. As a result, smooth shifting of the clutches 60 and 70 without significant speed ratio fluctuation and switching shock is realized.

On the other hand, at the time t1 when the actual toroidal ratio reaches the switching point toroidal ratio (Rtm).
Thus, the duty ratio for the low clutch is set to 0%, and the duty ratio for the high clutch is set to a predetermined duty ratio. As a result, the low clutch pressure (EL) gradually decreases and the low clutch 60 starts disengaging, while the high clutch pressure (EH) gradually increases and the high clutch 70 starts to be engaged. That is, when the actual toroidal ratio reaches the switching point toroidal ratio (Rtm), the switching operation of the clutches 60 and 70 is started.

Here, the predetermined number of pulses or the predetermined time Ta during which the sleeve position is kept fixed at the predetermined pulse number and the predetermined sleeve position at the time point t1 substantially coincides with the start of the shifting operation of the clutches 60 and 70,・
The time required for both high clutches 60 and 70 to be simultaneously engaged at time t2 is set. When the low and high clutches 60 and 70 are simultaneously engaged at the time t2 when the predetermined time R1 has elapsed, the toroidal ratio is changed to the switching point toroidal ratio regardless of the number of pulses or the sleeve position. (Rtm).

The predetermined time Ta is set to be shorter as the oil temperature is higher. This is because when the oil temperature is high, the viscosity of the working oil or the lubricating oil is low, the engagement operation of the engagement-side friction element (in this case, the high clutch 70) proceeds with good responsiveness, and the actual toroidal ratio is changed at the switching point. This is because the toroidal ratio (Rtm) is quickly reached. Thereby, the control time required for the mode switching can be reduced.

After the time point t2 when the predetermined time Ta has elapsed, the high clutch pressure (EH) further increases by further increasing the duty ratio for high clutch toward a duty ratio of 100%. On the other hand, the low clutch pressure (EL) is maintained at a constant pressure until a predetermined time Tb elapses. As a result, the low and high clutches 60 and 70 are simultaneously maintained in the engaged state.

At time t3 when the second predetermined time Tb has elapsed, the low clutch pressure (EL) starts to fall below the engagement pressure, whereby the disengagement operation of the low clutch 60 and the engagement operation of the high clutch 70 are started. We will go further. That is, the transition from the low mode operation state in which torque is transmitted via the first and second paths to the high mode operation state in which torque is transmitted via only the second path is started. You.

Then, a predetermined time Tc from the time t3
At the time point t4 when the high clutch 70 is completely engaged and the low clutch 60 is completely switched to the high mode HF in which the low clutch 60 is completely disengaged, and the actual gear ratio is set to the target gear ratio set by the target gear ratio setting means 311. The feedback control of the gear ratio that matches the toroidal ratio is restarted.

Next, the control operation of the torque fluctuation suppressing means 313 will be described. This torque fluctuation suppressing means 31
Reference numeral 3 denotes a case where the target gear ratio setting means 311 sets the target gear ratio with the above-described path switching, and suppresses the fluctuation of the torque input from the engine to the transmission during the execution of the path switching control. As above electric throttle 3
09 is configured to control the torque adjusting means.

That is, from time t1 when the actual toroidal ratio reaches the mode switching point toroidal ratio (Rtm), the high clutch 70 is completely engaged and the low clutch 60 is completely disengaged. By the time t4 when the switching of the route is completed,
When the driver performs an accelerator operation, the amount of change in the opening of the throttle valve corresponding to the accelerator operation amount is calculated as:
The fluctuation of the input torque is suppressed by, for example, executing a correction for decreasing the input torque as compared with a normal state.

More specifically, before and after the torque is transmitted through both the first path and the second path, that is, from the time t1 to the time t2 and from the time t1
From the time point 3 to the time point t4, the control is performed such that the change in the torque input from the engine 1 to the transmission is within the predetermined first allowable range, and the torque is transmitted through the two paths. From the time point t2 to the time point t3 when the clutches 60 and 70 are simultaneously engaged, the change in the torque input from the engine 1 to the transmission is set to a range smaller than the first allowable range. The control to be within the second allowable range is executed.

The control operation of the torque fluctuation suppressing means 313 will be described with reference to the flowcharts shown in FIGS. When the control operation starts, first, in step S1, based on the signals from the sensors 301 to 308, the current vehicle speed, the throttle opening of the engine 1, the accelerator operation amount, the selected range,
After detecting various state quantities such as the actual toroidal ratio and the oil temperature, in step S2, the current time is switched to the low / high mode when the target speed ratio with the above-described path switching is set by the target speed ratio setting means 311 in step S2. (Time t in FIG. 14)
It is determined whether or not 1).

If YES is determined in the step S2, the count value TM of the control time measuring timer is reset in the step S3, and the throttle opening tvo corresponding to the current accelerator operation amount is set in the step S4. After setting as the first reference opening degree tvo * 1 of the electric throttle 309, in step S5,
First allowable range of throttle opening tvmax1, tvo
In addition to setting a first threshold value ΔTVO1 for setting min1 and setting the first threshold value ΔT
By adding and subtracting VO1 to and from the first reference opening degree tvo * 1, the first allowable range tvmax1, tvo is set.
Calculate min1.

As shown in FIG. 17, the first threshold value ΔTVO1 is equal to the torque tq input from the engine 1 to the transmission.
1 is set based on a map using 1 as a parameter, and as the input torque tq1 increases, the first threshold value ΔTVO
1 is set to a small value.

If it is determined NO in step S2 and it is confirmed that the current time is not the switching point, then in step S7, it is determined whether or not the low / high mode switching control is currently being executed. Is determined. If NO is determined in step S7 and it is confirmed that the state is not in the state where the low / high mode switching control is being executed, the process proceeds to step S24 to be described later and the electric throttle 309 is controlled.

If it is determined as YES in step S7 and it is confirmed that the mode switching control has already been started, it is determined in step S8 whether the count value TM of the timer is smaller than t2. I do. If YES is determined in the above step S8 and it is confirmed that the present time is within the range from the time point t1 to the time point t2 shown in FIG. 14, in a step S9, the throttle opening degree tvo corresponding to the current accelerator operation amount is set. Is larger than the maximum value tvmax1 of the first allowable range. If YES is determined in step S9, the control value TVO for the throttle opening is corrected to the maximum value tvmax1 in step S10.

If NO is determined in step S9 and it is confirmed that the throttle opening degree tvo corresponding to the current accelerator operation amount is smaller than the maximum value tvmax1 of the first allowable range, then in step S11, It is determined whether or not the throttle opening degree tvo corresponding to the current accelerator operation amount is smaller than the minimum value tvomin1 of the first allowable range. If YES is determined in this step S10, in step S12, the control value TVO of the throttle opening is corrected to the minimum value tvomin1.

If NO is determined in step S8 and it is confirmed that the current time is out of the range from time t1 to time t2 shown in FIG. 14, then in step S13,
By determining whether or not the count value TM of the timer is equal to or greater than t3, the current time is determined as the time t3 shown in FIG.
It is determined whether or not it is after. If this determination result is YES, in step S14, it is determined whether or not the count value of the timer is t3, that is, whether or not the present time is the above-mentioned time t3.

If it is determined YES in step S14 and it is confirmed that the current time is the time point t3, the flow shifts to step S4 to change the current throttle opening degree tvo to the first reference opening degree tvo * 1. And the first threshold ΔTVO1 and the first throttle opening degree.
Allowable ranges tvmax1 and tvmin1 are set.

If NO is determined in step S14 and it is confirmed that the current time is within the range from time t3 to time t4 shown in FIG. 14, the process proceeds to step S9 to control the throttle opening control value. Control is performed to limit TVO within the first allowable range tvmax1 to tvmin1.

On the other hand, it is determined NO in step S13, and the present time is within the range from time t2 to time t3 shown in FIG. 14, that is, the torque of both the first path and the second path has passed. If it is confirmed that the transmission state is established, it is determined in step S15 whether the count value of the timer is t2, that is, whether the current time is the time t2.

If the determination in step S15 is YES, the current throttle opening tvo is set as the second reference opening tvo * 2 in step S16.
In step S17, a second threshold value ΔTVO2 for setting the second allowable ranges tvmax2 and tvmin2 of the throttle opening is set based on the input torque.
As shown in FIG. 18, the two thresholds ΔTVO2 are set based on a map using the torque tq2 input from the engine 1 to the transmission as a parameter, and the input torque tq
As the value of 2 is larger, the second threshold value ΔTVO2 is set to a smaller value.

Next, in step S18, the second threshold ΔTVO2 is added to and subtracted from the second reference opening degree tvo * 2, thereby obtaining the second allowable range tvmax.
After calculating 2, vomin2, in step S19, it is determined whether or not the current throttle opening degree tvo is larger than the maximum value tvmax2 of the second allowable range. If YES is determined in step S19, the control value TVO of the throttle opening is corrected to the maximum value tvmax2 in step S20.

If NO is determined in step S19 and it is confirmed that the current throttle opening tvo is smaller than the maximum value tvmax2 of the second allowable range, then in step S21, the current throttle opening tvo is determined.
It is determined whether or not vo is smaller than the minimum value tvmin2 of the second allowable range. YES in this step S21
Is determined in step S22, the control value TVO of the throttle opening is set to the minimum value tvomin2
To be corrected.

Next, in step S23, after the count value TM of the timer is incremented by one,
In step S24, the electric throttle 309 is determined based on the throttle opening tvo corresponding to the current accelerator operation amount or the corrected throttle opening control value TVO.
The input torque is controlled by changing the opening of the motor.

When the above control is executed, for example, as shown in FIG.
In the state where the target opening degree tvoT is gradually increasing, when the start time t1 of the mode switching control is reached, the change in the opening degree of the electric throttle 309 based on the throttle opening degree tvo * 1 at this time point t1 is determined. The first threshold value ΔTVO1 for suppressing is set. Then, as a result of prohibiting a change in the throttle opening exceeding a first allowable range set based on the first threshold ΔTVO1, the amount of increase in the torque input from the engine 1 to the transmission is reduced by the first threshold. It will be suppressed within the first allowable range corresponding to ΔTVO1.

When the clutches 60 and 70 are simultaneously engaged, that is, when the torque is transmitted via the first and second paths, the throttle opening at time t2 is set. A second threshold value ΔTVO2 for suppressing a change in the opening degree of the electric throttle 309 is set based on tvo * 2. This second threshold value ΔTV
Since O2 is set to a value smaller than the first threshold value ΔTVO1, an increase in the input torque in a state where the both clutches 60 and 70 are simultaneously engaged is further suppressed, and corresponds to the second threshold value ΔTVO2. It will be kept within the second allowable range.

It should be noted that, from the state in which both clutches 60 and 70 are simultaneously engaged, the point in time when the engagement of low clutch 60 is released, that is, the point in time t3 when the state shifts to the power transmission state via only the second path, Until the time point t4 at which the mode switching control ends, the first allowable increase in the torque input from the engine 1 to the transmission is performed based on the first threshold value ΔTVO1 set based on the throttle opening at the time point t3. Control for suppressing the range is performed.

As shown in FIGS. 19B and 19C, when the target opening degree tvoT of the electric throttle 309 corresponding to the accelerator depression amount once increases and then decreases, the first path Before and after the torque is transmitted via both the second path and the second path (between time t1 and time t2 and between time t3 and time t4), the increase and decrease of the torque input from the engine 1 to the transmission are determined. The electric throttle 309 is controlled so as to fall within a first allowable range corresponding to the first threshold value ΔTVO1.

When the torque is transmitted through the two paths, the change in the torque input from the engine 1 to the transmission is changed to a second value smaller than the first threshold value ΔTVO1. The second corresponding to the threshold value ΔTVO2
By controlling so as to be within the allowable range, the increase and decrease of the input torque are further suppressed.

As shown in FIG. 20A, when the target opening degree tvo of the electric throttle 309 corresponding to the operation amount of the accelerator continues to decrease, the decrease amount of the input torque is set to the first amount. 20B, the target opening degree tvo of the electric throttle 309 corresponding to the depression amount of the accelerator is set to the first ΔTVO1 and the second threshold ΔTV, as shown in FIG.
If it fluctuates within the range of O2 or less, the change in the input torque is not suppressed and the throttle opening tv is not changed.
The input torque changes corresponding to the target opening degree tvoT.

As described above, the first path for transmitting the torque of the output source composed of the engine 1 to the driven portion via the first and second continuously variable transmission mechanisms 20 and 30 and the planetary gear mechanism 50, In a power train having a transmission provided with a second path for transmitting to a driven portion via only the first and second continuously variable transmission mechanisms 20 and 30, the target transmission ratio setting means 311 controls the transmission of the path. When the target gear ratio with the switching is set, the torque fluctuation suppressing means 313 for suppressing the fluctuation of the torque input from the engine 1 to the transmission during the execution of the path switching control is provided. In the transient operation state in which the torque input from the transmission to the transmission changes, the switching shock caused by the switching of the path can be effectively prevented.

That is, if the input torque greatly increases or decreases during the above-described path switching operation, that is, during the low / high mode switching control, the power rollers 23 and 33 tilt according to the increase or decrease in the input torque. Continuously variable transmission mechanism 2
Since the gear ratios 0 and 30 greatly deviate from the switching point, there is a possibility that a sudden change in the gear ratio at the end of the above-mentioned path switching may cause a switching shock. In contrast to this, while executing the path switching control as described above,
In the case where the fluctuation of the torque inputted from the engine 1 to the transmission is suppressed, it is possible to effectively prevent the switching shock caused by the sudden change of the gear ratio at the time when the switching of the path is completed. Can be.

In particular, when the engagement-side friction element and the disengagement-side friction element including the low clutch 60 and the high clutch 70 are simultaneously engaged, the transmission of power in the low mode via the first path is not performed. Since the gear ratio is fixed to the predetermined gear ratio, the gear ratio deviation corresponding to the torque change is caused by slip on the toroidal surface and the clutch 60,
70 is absorbed by slips and the like, and these are instantaneously released after the end of the above-mentioned switching operation, so that a switching shock occurs. Therefore, as described above, at least the low clutch 60 and the high clutch 70
At the same time, that is, when the torque is transmitted through both the first path and the second path, the fluctuation of the torque input from the engine 1 to the transmission is determined by the torque With the configuration in which the fluctuation is suppressed by the fluctuation suppressing unit 313, the occurrence of the switching shock can be more effectively prevented.

In the above embodiment, the electric throttle 309 is controlled so that a change in torque input from the engine 1 to the transmission falls within an allowable range.
Since the allowable range is set to a value corresponding to the operation state, that is, the input torque, the value of the input torque is large, and the change in the input torque causes the speed ratio of the continuously variable transmission mechanisms 20, 30 to be remarkable. The occurrence of the switching shock can be effectively suppressed in an operating state in which the switching shock is easily changed. In addition, in an operating state in which the value of the input torque is small and the speed ratio of the continuously variable transmission mechanisms 20 and 30 does not greatly change due to the change, the operating range intended by the driver is expanded by increasing the allowable range. There is an advantage that can be maintained.

In particular, as shown in the above embodiment, before and after the torque is transmitted via both the first path and the second path (from time t1 to time t2 and from time t3 to time t)
4) The electric throttle 3 is controlled so that a change in torque input from the engine 1 to the transmission falls within a first allowable range.
09 when the torque is transmitted through the two paths, the change in the torque input from the engine 1 to the transmission is set to a second range smaller than the first allowable range. Electric throttle 30 within the allowable range
9, the shift of the gear ratio corresponding to the torque change is absorbed by the slip of the toroidal surface and the slip of the clutches 60 and 70 in the state of transmission of the torque via the two paths. It is possible to suppress the occurrence of the switching shock caused by the operation. In addition, before and after the transmission of the torque via the two paths, the fluctuation of the input torque is allowed to some extent, so that the driving state intended by the driver can be maintained.

In the above-described embodiment, an example will be described in which the torque fluctuation suppressing means 313 controls the torque adjusting means including the electric throttle 309 to suppress the fluctuation of the torque input from the engine to the transmission. However, in place of the electric throttle 309, a torque adjusting means such as an ignition timing control means for controlling the ignition timing of the engine 1 is provided, and the fluctuation of the input torque is suppressed by controlling the retard amount of the ignition timing. You may do so. Further, the change of the input torque may be prohibited while the switching control of the path is executed.

Further, as shown in FIGS. 19 (b) and 19 (c), at the time point t4 when the path switching is completed, the electric throttle 30
9 and the target opening tv based on the accelerator depression amount
When oT is greatly different, by setting a coefficient corresponding to the engine output to make the two coincide with each other,
If it is configured to match the target value while limiting the change in the actual opening amount, a sudden change in the engine output can be prevented.

[0143]

As described above, according to the present invention, only the first path for transmitting the torque of the output source to the driven portion via the continuously variable transmission mechanism and the gear mechanism, and only the continuously variable transmission mechanism are used. A target transmission ratio setting means for setting a target transmission ratio of the transmission based on a running state of the vehicle, in a power train having a transmission provided with a second path for transmitting to a driven portion via the transmission; Speed ratio control means for executing speed ratio control and path switching control of the continuously variable transmission mechanism so as to achieve the target speed ratio set by the target speed ratio setting means; and A torque adjusting means for adjusting the input torque in accordance with the accelerator operation, and a time when the target speed ratio accompanying the switching of the route is set by the target speed ratio setting means, while executing the switching control of the route. , The above output source And the torque fluctuation suppressing means for suppressing the fluctuation of the torque input to the transmission from the above, the above-mentioned path switching is performed in a transient operation state in which the torque input from the output source to the transmission changes. There is an advantage that the switching shock can be effectively prevented.

[Brief description of the drawings]

FIG. 1 is an explanatory diagram showing an overall configuration of a power train.

FIG. 2 is a plan sectional view showing a specific configuration of the transmission.

FIG. 3 is a sectional view taken along line AA of FIG. 2;

FIG. 4 is an explanatory diagram showing a hydraulic circuit configuration of a transmission.

FIG. 5 is a sectional view of the transmission control three-layer valve as viewed from a direction B in FIG. 3;

6 is a cross-sectional view of the speed ratio control mechanism as viewed from a direction C in FIG.

FIG. 7 is a block diagram showing a control system of the entire power train.

FIG. 8 is a characteristic diagram showing a relationship between the number of pulses of a step motor and a toroidal ratio.

FIG. 9 is a characteristic diagram showing a relationship between the number of pulses of a step motor and a unit ratio.

FIG. 10 is a characteristic diagram showing a relationship between a unit ratio and a toroidal ratio.

FIG. 11 is a shift diagram in a forward low mode.

FIG. 12 is a shift diagram in a forward high mode.

FIG. 13 is a shift diagram in a reverse low mode.

FIG. 14 is a time chart showing a mode switching operation.

FIG. 15 is a flowchart showing a first half of a mode switching operation.

FIG. 16 is a flowchart showing the latter half of the mode switching operation.

FIG. 17 is a characteristic diagram showing a relationship between a first threshold value and an input torque.

FIG. 18 is a characteristic diagram illustrating a relationship between a second threshold value and an input torque.

FIG. 19 is a time chart showing how the slot opening changes during the switching operation.

FIG. 20 is a time chart showing a change state of a slot opening during a switching operation.

[Explanation of symbols]

 DESCRIPTION OF SYMBOLS 1 Engine (power source) 10 Power train 20 and 30 Continuously variable transmission mechanism 50 Planetary gear mechanism (Gear mechanism) 60 Low clutch (friction element) 70 High clutch (friction element) 309 Electric throttle (torque adjusting means) 311 Target gear ratio Setting means 312 Gear ratio control means 313 Torque fluctuation suppression means

Claims (4)

[Claims] A first path for transmitting a torque of an output source to a driven part via a continuously variable transmission mechanism and a gear mechanism;
In a power train having a transmission provided with a second path for transmitting to a driven portion via only a continuously variable transmission mechanism, a target for setting a target transmission ratio of the transmission based on a running state of the vehicle Speed ratio setting means; speed ratio control means for executing speed ratio control and path switching control of the continuously variable transmission mechanism so that the target speed ratio set by the target speed ratio setting means is realized; A torque adjusting means for adjusting the torque input from the output source to the transmission in accordance with the accelerator operation; and a target speed ratio with the switching of the path set by the target speed ratio setting means, A control device for a power train, comprising: torque fluctuation suppressing means for suppressing fluctuation of torque input from the output source to the transmission during execution of the switching control. 2. The power train control device according to claim 1, wherein the torque fluctuation suppressing unit is configured to transmit torque when at least the first path and the second path are transmitted. A control device for a power train, wherein the control device is configured to suppress a change in torque input to the transmission from the output source. 3. The power train control device according to claim 2, wherein the torque fluctuation suppressing means controls the torque adjusting means such that a change in torque input from the output source to the transmission falls within an allowable range. A control device for a power train, wherein the control device is configured to control and set the allowable range to a value corresponding to an operation state. 4. The power train control device according to claim 2, wherein said torque fluctuation suppressing means comprises a first power supply.
Before and after the transmission of the torque via both the first path and the second path, the torque adjusting means such that the change in the torque input from the output source to the transmission falls within a first allowable range. And when the torque is transmitted through the two paths, the change in the torque input from the output source to the transmission is set to a second range smaller than the first allowable range. A power train control device configured to control the torque adjusting means so as to fall within an allowable range.