JP2000234670A - Shift control device of transmission ratio infinite continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To accurately control the transmission ratio even at low oil temperature time by driving a shift control valve on the basis of the target transmission ratio and the actual transmission ratio from a mechanical feedback means, and controlling a flow rate and the direction of a hydraulic fluid supplied to a hydraulic cylinder connected to a trunnion. SOLUTION: A toroidal continuously variable transmission 2 and a constant transmission 3 including a power roller sandwiched between input and output discs so as to be freely tiltable are connected to an output shaft 1 in a parallel state, and output shafts of these transmissions 2, 3 are connected to an output shaft 6 via a planetary gear mechanism 5, a motive power circulating mode clutch 9 and a dirctly coupled mode clutch 10. In controlling a shift of such transmissions, a shift control valve using a step motor 36 as a displacement driving source is provided to control the supply direction and hydraulic pressure supplied to a hydraulic cylinder for tilting the power roller. When a vehicle driving state becomes a preset state, the step motor 36 is controlled in driving according to preset creep torque and the actual transmission ratio.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift control device for a continuously variable transmission having an infinitely variable speed ratio employed in a vehicle or the like, and more particularly to a shift control device employing a toroidal type continuously variable transmission as a continuously variable transmission. It is.

[0002]

2. Description of the Related Art A continuously variable transmission of a toroidal type has been known as a transmission of a vehicle. In order to further expand the shift range of such a continuously variable transmission, a continuously variable transmission has been required. A continuously variable transmission with an infinitely variable transmission ratio is known, which is capable of controlling the transmission ratio to infinity by combining a transmission with a planetary gear mechanism, for example, Japanese Patent Application Laid-Open No. H10-267117.

[0003] This is a toroidal-type continuously variable transmission capable of continuously changing the gear ratio on a unit input shaft of an infinitely variable gear ratio continuously variable transmission connected to an engine, and a constant transmission (reduction gear). The output shafts are connected in parallel, and these output shafts are selectively connected by a planetary gear mechanism. As shown in FIG. 18, while the power circulation mode clutch is connected, the direct connection mode clutch is cut off, The unit speed ratio ii (unit input shaft speed / unit output shaft speed in the IVT ratio in the figure) is changed from a negative value to a positive value according to the difference between the gear ratio of the stepped transmission and the fixed transmission. A power circulating mode in which gear shifting control is continuously performed including a large (= geared neutral point GNP), and a power circulating mode clutch is cut off while a direct connection mode clutch is connected to change the gear ratio ic of the continuously variable transmission. The direct mode which performs shift control in response can be selectively used.

In the above-described continuously variable transmission with an infinite transmission ratio, the hydraulic cylinder that drives the trunnion of the toroidal-type continuously variable transmission in the axial direction is controlled by a step motor, and the geared neutral point GNP is controlled by a step motor. By instructing the number of steps to be shifted by a fixed amount from the number of steps corresponding to the geared neutral point GNP, and increasing or decreasing the differential pressure applied to the front and back of the piston of the hydraulic cylinder by using a relief valve, The creep torque transmitted to the drive wheels at low vehicle speed is controlled according to the operation state of the brake.

[0005]

However, in the above conventional example, since the differential pressure of the hydraulic cylinder is controlled by a pressure control valve such as a relief valve, the response of the hydraulic control is low when the temperature of the hydraulic oil is low. Therefore, there is a problem that the differential pressure of the hydraulic cylinder cannot be accurately controlled, and the control accuracy of the creep torque decreases.

The present invention has been made in view of the above problems, and has as its object to control the creep torque near the geared neutral point in the power circulation mode with high accuracy regardless of the oil temperature of the transmission. .

[0007]

A first aspect of the present invention is directed to a toroidal type continuously variable transmission and a constant transmission which continuously change the gear ratio by tilting a power roller held between input and output disks. An infinitely variable transmission ratio in which the output shafts of the toroidal type continuously variable transmission and the fixed transmission are connected to the unit output shaft via a planetary gear mechanism, power circulation mode clutch and direct connection mode clutch, while being connected to the unit input shaft. A transmission control apparatus for a continuously variable transmission with an infinite transmission ratio, comprising: a transmission, a hydraulic cylinder connected to a trunnion that supports the power roller, and transmission control means for controlling the hydraulic pressure of the hydraulic cylinder. A shift control means for controlling a hydraulic pressure supplied to the hydraulic cylinder and a supply direction; an actuator for driving the shift control valve; Mechanical feedback means for feeding back the actual gear ratio of the continuously variable transmission to the shift control valve; operating state detecting means for detecting an operating state of the vehicle; and a preset state when the operating state matches a preset state. A shift commanding means for calculating a drive amount of the actuator based on a creep torque and a gear ratio of the toroidal-type continuously variable transmission, and controlling the actuator based on the drive amount.

In a second aspect based on the first aspect, the driving state detecting means comprises means for detecting a vehicle speed and means for detecting an amount of depression of an accelerator pedal. When the detected vehicle speed is lower than the preset vehicle speed and the accelerator pedal depression amount is 0, it is determined that the vehicle speed matches the preset state.

In a third aspect based on the first aspect, the shift command means corrects a drive position of an actuator corresponding to a geared neutral point in a no-load state with a value corresponding to the creep torque. These are command values to the actuator.

In a fourth aspect based on the third aspect, the speed change command means sets a differential pressure of the hydraulic cylinder according to a speed ratio of the toroidal type continuously variable transmission and the creep torque. Differential pressure setting means and a command value for the actuator are set according to the differential pressure.

In a fifth aspect based on the third aspect, the speed change command means includes an input torque calculating means for calculating an input torque of the infinitely variable speed ratio transmission, and an infinitely variable speed ratio continuously variable transmission. A torque ratio calculating means for calculating a ratio between an input torque and an output torque of the transmission as a torque ratio; and setting a command value of the actuator based on the torque ratio obtained by using the creep torque as an output torque.

In a sixth aspect based on the third aspect, the speed change command means corresponds to an input torque calculation means for calculating an input torque of a continuously variable transmission with an infinite speed ratio, and the creep torque. Drive amount setting means for determining a target speed ratio of the toroidal type continuously variable transmission based on the output torque and setting a drive amount of the actuator so as to achieve the target speed ratio is provided.

[0013] In a seventh aspect based on the sixth aspect, the speed change command means uses the input torque of the infinitely variable speed ratio continuously variable transmission as a parameter and an output torque of the toroidal type continuously variable transmission. A map in which the relationship between the gear ratios of the toroidal-type continuously variable transmission is set in advance is provided.

In an eighth aspect based on the sixth aspect, the speed change command means uses the speed ratio of the continuously variable transmission with an infinite gear ratio as a parameter to set the input torque and the output of the toroidal type continuously variable transmission. It has a map in which the relationship of torque is set in advance.

In a ninth aspect based on the eighth aspect, the speed change control means includes a predetermined input torque and an output torque which are set in advance in consideration of an internal loss of a continuously variable transmission with an infinite gear ratio. Based on the relationship, a target speed ratio of the toroidal-type continuously variable transmission is set from an output torque corresponding to the creep torque.

[0016]

According to the first aspect of the present invention, the gear ratio control is performed during normal running, and the gear ratio control valve is controlled in accordance with the target gear ratio commanded by the actuator and the feedback of the actual gear ratio from the mechanical feedback means. Since the flow rate and the direction of the hydraulic oil that is driven and supplied to the hydraulic cylinder connected to the trunnion are controlled, the gear ratio can be accurately controlled even when the oil temperature is low.

When the operating condition becomes a predetermined condition, for example, in the vicinity of the power circulation mode, the torque shift of the toroidal-type continuously variable transmission is used to determine the actuator based on the target creep torque and the gear ratio. By controlling the creep torque, the creep torque can be controlled in the vicinity of the geared neutral point by controlling the target creep torque according to the gear ratio, even at a low oil temperature. This enables accurate and stable creep torque control regardless of fluctuations in oil temperature, thereby greatly improving the operability of a continuously variable transmission with an infinite speed ratio.

Further, the second invention is provided when the vehicle speed is less than a predetermined vehicle speed and the accelerator pedal depression amount becomes zero, that is, when the accelerator pedal is released near the geared neutral point. It becomes possible to start the control of the creep torque.

According to a third aspect of the present invention, in the control of the creep torque, the driving position of the actuator corresponding to the geared neutral point in the no-load state is used as a reference.
By correcting according to the target creep torque,
Since the command value of the actuator can be easily determined, it is possible to control the creep torque with high accuracy even by, for example, open loop control.

According to a fourth aspect of the present invention, in the vicinity of a geared neutral point, a differential pressure of a hydraulic cylinder is determined in accordance with a speed ratio of a toroidal type continuously variable transmission and a target creep torque. Therefore, the creep torque can be easily controlled.

According to a fifth aspect of the present invention, in the vicinity of the geared neutral point, the actuator is driven by changing the torque ratio of the continuously variable transmission with an infinite gear ratio to drive the actuator so as to realize the creep torque with respect to the input torque. An arbitrary creep torque can be obtained regardless of the temperature.

According to a sixth aspect of the present invention, in the vicinity of the geared neutral point, a target gear ratio is determined such that the output torque becomes a target creep torque based on a preset relationship between the input torque and the output torque. Since the actuator is driven according to the target gear ratio, an arbitrary creep torque can be realized with high accuracy even at a low oil temperature.

The seventh or eighth invention uses a map which defines the relationship between the output torque of a toroidal type continuously variable transmission and the gear ratio, using the input torque of the infinitely variable speed ratio transmission as a parameter. By setting input torque = engine torque and output torque = target creep torque, the target gear ratio can be easily obtained. By setting the drive position of the actuator corresponding to the target gear ratio as a command value, Any creep torque can be realized with high accuracy even at low oil temperature.

According to the ninth aspect of the present invention, since the internal loss of the continuously variable transmission with an infinite transmission ratio is added to the relationship between the input torque and the output torque, the control accuracy of the creep torque can be further improved.

[0025]

An embodiment of the present invention will be described below with reference to the accompanying drawings.

FIGS. 1 and 2 show an example in which a double-cavity toroidal type continuously variable transmission 2 composed of a half toroid is used to form an infinitely variable speed ratio continuously variable transmission.

As shown in FIGS. 1 and 2, the continuously variable transmission with an infinite transmission ratio continuously changes the transmission ratio to unit input shafts 1a and 1b connected to a crankshaft (not shown) of the engine. Possible toroidal type continuously variable transmission 2 and gear 3
a, a constant transmission 3 (reduction gear) composed of a counter gear 3b is connected in parallel, and these output shafts 4 and 3c are arranged coaxially with the unit output shaft 6 and the planetary gear mechanism 5 The output shaft 4 of the toroidal type continuously variable transmission 2 is connected to a sun gear 5 a of a planetary gear mechanism 5.
The output shaft 3c of the constant transmission 3 is connected to the carrier 5b of the planetary gear mechanism 5 via the power circulation mode clutch 9.

The continuously variable transmission output shaft 4 connected to the sun gear 5a receives the driving force of the toroidal type continuously variable transmission 2 from the gear 4a and the chain 40, and receives an infinitely variable speed ratio through the direct connection mode clutch 10. The ring gear 5c is also connected to the unit output shaft 6, while being connected to the unit output shaft 6 which is the output shaft of the machine.

A transmission output gear 7 is provided on the right side of the unit output shaft 6 in the drawing, and the transmission output gear 7 meshes with the final gear 12 of the differential gear 8 and is connected to the drive shaft 11 a connected to the differential gear 8. , 11b
The driving force is transmitted at a predetermined total reduction ratio. ｛1. Transmission mechanism｝ Toroidal-type continuously variable transmission 2 has two sets of input discs 21 and output discs 2 as shown in FIG.
2, a double-cavity half-toroidal type that sandwiches and presses the power rollers 20, 20;
Output gear 2a interposed between a pair of output disks 22
Is connected via a chain 40 to a gear 4a formed on the continuously variable transmission output shaft 4 of the unit output shaft 6 arranged in parallel with the unit input shafts 1a and 1b.

As shown in FIG. 2, the unit input shafts 1a and 1b are arranged coaxially and connected in a rotational direction via a loading device (not shown).
The unit input shaft 1a is coupled to the crankshaft of the engine and forms a gear 3a of the constant transmission 3,
The unit input shaft 1b is connected to two sets of input discs 21 and 21, and the axial pressure generated by the loading device in response to the input torque from the unit input shaft 1a causes the power rollers 20 and 21 shown in FIG. 20 is pinched and pressed between the input and output disks.

As shown in FIG. 3, the lower portions of the power rollers 20, 20 which are arranged at the opposing positions are connected to a hydraulic cylinder 30 to form a trunnion 23 which is axially displaceable and rotatable around the axis. Each of the plurality of trunnions 23 is rotatably supported at a lower end of one of the plurality of trunnions 23 by a rotation angle of the trunnion 23 (hereinafter, referred to as a tilt angle φ of the power roller 20).
A precess cam 35 is provided for synthesizing the actual gear ratio and the axial displacement of the trunnion 23 and feeding them back.

The hydraulic cylinder 30 has upper and lower oil chambers 30A and 30B defined by a piston 31, and as shown in FIG.
The hydraulic cylinders 30, 30 are set such that the arrangement of the oil chambers 30A, 30B is reversed to each other, and the opposing trunnions 23, 23 are driven in mutually opposite directions.

Therefore, when the oil pressure in the oil chamber 30B is reduced at the same time as the oil pressure in the oil chamber 30A is increased, the trunnion 23 on the right side in FIG.
Is lowered, and the power rollers 20, 20 are tilted to the Lo side (speed ratio ic = large side) of the toroidal type continuously variable transmission 2 to perform the speed change.

On the other hand, when the oil pressure in the oil chamber 30A is reduced and the oil pressure in the oil chamber 30B is increased at the same time, the trunnion 23 on the right in FIG. The shift is performed by tilting to the Hi side (speed ratio ic = small side) of the toroidal type continuously variable transmission 2.

As shown in FIG. 3, the precess cam 35 has a cam groove 35A having a predetermined inclination in a circumferential direction, and one end of a swingable feedback link 38 is engaged with the cam groove 35A. I do.

The feedback link 38 is, for example, L
And is supported so as to be swingable about a swing shaft 39. One end of the trunnion 23 engages with the cam groove 35A and the other end engages with one end of the transmission link 37. , That is, the tilt angle and the axial displacement amount are transmitted to one end of the speed change link 37.

As shown in FIG. 4, the speed change link 37 has a spool 46S of the shift control valve 46 at the center.
While the other end opposite to the end connected to the feedback link 38 is connected to a step motor 36, and the speed change link 37 displaces the shift control valve 46 in the axial direction by the drive of the step motor 36. At the same time, the shift control valve 46 is displaced in the axial direction according to the rotation of the trunnion 23 and the axial displacement.

As shown in FIG. 2, the speed change control of the continuously variable transmission having an infinite speed ratio is performed by a speed change controller 80 mainly constituted by a microcomputer, and the rotation speed Ni () of the unit input shaft 1a or 1b is controlled. Output from the input shaft speed sensor 81 for detecting the engine speed Ne), output from the continuously variable transmission output shaft speed sensor 82 for detecting the speed Nco of the continuously variable transmission output shaft 4, and a unit. The output from the vehicle speed sensor 83 that detects the vehicle speed VSP from the rotation speed of the output shaft 6 and the like, the depression amount Acc of the accelerator pedal detected by the accelerator opening sensor 86, the brake signal from the brake switch 87, and the like are input, and the speed is changed. The controller 80 processes these detected values as operating states, and according to the operating states, the solenoids 91, 92
, The power circulation mode clutch 9 and the direct connection mode clutch 10 are selectively engaged to switch between the power circulation mode and the direct connection mode, and the stepping motor 36 is controlled so that the unit speed ratio ii according to the operation state is obtained. By driving, the speed ratio ic of the toroidal type continuously variable transmission 2 is controlled.

In this speed change control, the accelerator pedal depression amount A during normal running is the same as in the conventional speed ratio control.
While the gear ratio control is performed so that the gear ratio is in accordance with the cc and the vehicle speed VSP, the control of the transmission torque is performed as described later so that an arbitrary creep torque is obtained in the vicinity of the geared neutral point GNP in the power circulation mode. . {2. Hydraulic Control Circuit Next, the hydraulic control device will be described in detail with reference to FIG.

First, in the hydraulic control device, the hydraulic pressure supplied from the hydraulic pump is adjusted to a predetermined pressure by a pressure regulator 100 controlled by a PL solenoid 90, and is supplied to a line pressure circuit 101 as a line pressure PL.

The line pressure circuit 101 is connected to the shift control valve 46 for controlling the flow rate and the supply direction to the hydraulic cylinder 30 for driving the trunnion 23. The spool 4 is moved according to the displacement of the step motor 36 or the feedback link 38 controlled by the controller 80.
6S is displaced, and the line pressure PL is changed to two oil chambers 30A, 30B of the hydraulic cylinder 30 according to the displacement amount of the spool 46S.
Supply to one of the

The line pressure circuit 101 is provided with a solenoid 91 for controlling the direct connection mode clutch 10 and a solenoid 92 for controlling the power circulation mode clutch 9.

In response to the increase in the signal pressure from the solenoid 91, the control valve 93 supplies the line pressure PL from the manual valve 60 to the direct connection mode clutch 10 and engages, while controlling when the signal pressure from the solenoid 91 decreases. Valve 9
3 connects and disconnects the direct connection mode clutch 10 to the drain.

Similarly, in response to an increase in the signal pressure from the solenoid 92, the control valve 94 supplies the line pressure PL from the manual valve 60 supplied via the shuttle valve 121 to the power circulation mode clutch and engages. On the other hand, when the signal pressure from the solenoid 92 decreases, the control valve 94 connects the power circulation mode clutch 9 to the drain and releases it.

One of the power circulation mode clutch 9 and the direct connection mode clutch 10 is engaged by the solenoids 91 and 92, and the power circulation mode and the direct connection mode are selectively switched.

Here, the shift control valve 46
Is a supply port 46P communicating with the line pressure circuit 101.
, A Lo port 46L communicating with the oil chamber 30A of the hydraulic cylinder 30, a Hi port 46H communicating with the oil chamber 30B of the hydraulic cylinder 30, and two drain ports 46D.
Are provided with the supply port 46P interposed therebetween, and the supply port 46P moves from the supply port 46P in accordance with the axial displacement of the spool 46S.
The line pressure PL is regulated and supplied to one of the o-side port 46L and the Hi-side port 46H, while the other port communicates with the drain port 46D.

That is, when the spool 46S is at the neutral position, the supply port 46P, the drain port 46D,
The Lo side port 46L and the Hi side port 46H are each sealed, and the oil pressure of the oil chamber 30A and the oil chamber 30B is held.

From this neutral position, the step motor 36 is
When driven to the o-side, first, the spool 46S is displaced upward in the drawing, and the supply port 46P and the Lo-side port 46L communicate with each other.
D and a flow rate corresponding to the opening amount of the supply port 46P is supplied to the Lo-side port 46L.

Conversely, when the step motor 36 is driven to the Hi side, the spool 46S is displaced downward from the neutral position in the figure, and the supply port 46P and the Hi port 46H communicate with each other, while the Lo port 46L is connected to the drain port. 46D, and the flow rate according to the opening amount of the supply port 46P is Hi.
It is supplied to the side port 46H.

Here, the speed change control of the toroidal type continuously variable transmission 2 is performed by the speed change controller 80 by the step motor 3.
6 is driven according to a target value such as a speed ratio or a differential pressure. The relationship between the speed ratio ic of the toroidal type continuously variable transmission 2 and the number of steps STEP of the step motor 36 is as shown in FIG. It is set in advance and the number of steps STE
The minimum position of P becomes the minimum Lo gear ratio of the toroidal-type continuously variable transmission 2 and the tilt angle φ of the power roller 20 becomes the minimum, while the maximum position of the number of steps STEP becomes the maximum Hi gear ratio side (in the figure). (To the right of the rotation synchronization point RSP), the tilt angle φ of the power roller 20 increases, and in the power circulation mode, as shown in FIG. As the speed ratio ic of No. 2 becomes higher, the unit speed ratio i of the continuously variable transmission with infinite speed ratio is increased.
i is Lo side and geared neutral point GN
P (speed ratio ic = ic0, tilt angle φ = φ0)
Furthermore, on the reverse side, the geared neutral point GN
As the speed ratio ic of the toroidal type continuously variable transmission 2 decreases from P, the unit speed ratio ii becomes Hi in the reverse direction.

When the target gear ratio changes to the Lo side, the step motor 36 moves one end of the gear change link 37 to the position shown in FIG.
When the tilt angle φ of the power roller 20 is in a steady state at this time, since the precess cam 35 has stopped, the spool 46S is also displaced upward, and the supply port 46P is displaced upward. While the Lo-side port 46L communicates, the Hi-side port 46H communicates with the drain port 46D, and the oil pressure in the oil chamber 30A increases in accordance with the flow rate supplied from the supply port 46P via the Lo-side port 46L. The oil pressure in the chamber 30B is discharged from the drain port 46D, the right trunnion 23 shown in FIG. 3 rises in the direction of the arrow shown by the solid line, and the power roller 20 tilts with the rise of the trunnion 23.

By driving the hydraulic cylinder 30, the trunnion 23 is displaced in the axial direction and around the axis, and the displacement of the trunnion 23 is transmitted to the speed change link 37 via the feedback link 38, and the power roller 20 is tilted to the Lo side. In response to the roll, the feedback link 38 displaces the left end of the transmission link 37 downward in FIG.

Therefore, the spool 46S that has been displaced upward is displaced downward toward the neutral position, and when the tilt angle φ of the power roller 20 matches the target gear ratio, the spool 46S is returned to the neutral position again. Return to hydraulic cylinder 3
The drive of 0 is stopped.

Thus, at the time of gear shifting, first, the spool 46S is driven by the step motor 36,
Hydraulic oil is supplied from the line pressure circuit 101 to the oil chamber 30A, while pressure oil in the oil chamber 30B is discharged to the tank, and the trunnion 23 is displaced, so that the tilt angle φ of the power roller 20 is shifted to the Lo side. Then, the tilt angle φ of the power roller 20 and the axial displacement of the trunnion 23 are
5. Since the feedback is provided to the shift control valve 46 via the feedback link 38 and the speed change link 37, the spool 46S gradually returns to the neutral position and maintains the state in which the step gear 36 matches the target speed ratio commanded by the step motor 36. can do.

On the other hand, when the target speed ratio changes to the Hi side, the step motor 36 connects one end of the speed change link 37 to
When the inclination angle φ of the power roller 20 is in a steady state at this time, the precess cam 35 is stopped, and the spool 46S is also displaced downward to supply the supply port. 46P communicates with the Hi port 46H, while the Lo port 46L connects with the drain port 46H.
D and supply port 46P to Hi-side port 46H
The oil pressure in the oil chamber 30B rises in accordance with the flow rate supplied through the
6D, the right trunnion 23 shown in FIG.
Falls in the direction of the arrow shown by the broken line, and the power roller 20 tilts with the fall of the trunnion 23.

By driving the hydraulic cylinder 30, the trunnion 23 is displaced in the axial direction and around the axis. The displacement of the trunnion 23 is transmitted to the speed change link 37 via the feedback link 38, and the power roller 20 is tilted to the Hi side. In response to the rolling, the feedback link 38 displaces the left end of the speed change link 37 upward in FIG.

Therefore, the spool 46S that has been displaced downward is displaced upward toward the neutral position, and when the tilt angle φ of the power roller 20 matches the target gear ratio, the spool 46S is returned to the neutral position again. Return to hydraulic cylinder 3
The drive of 0 is stopped. {3. Next, an example of creep torque control performed by the shift control controller 80 near the geared neutral point GNP in the power circulation mode will be described in detail with reference to the flowchart of FIG. Note that the flowchart of FIG. 6 is executed every predetermined time, for example, every several tens msec.

First, in step S1, the vehicle speed sensor 83
, The vehicle input speed detected by the input shaft speed sensor 81, the unit input shaft speed Ni (= engine speed Ne) detected by the input shaft speed sensor 81, and the output of the toroidal type continuously variable transmission 2 detected by the output shaft speed sensor 82. Side rotation speed Nco, accelerator pedal depression amount Ac from accelerator opening sensor 86
c) Read the brake signal from the brake switch 87.

Then, in a step S2, it is determined whether or not the vehicle speed VSP is lower than a predetermined vehicle speed VSP0. Here, the vehicle speed VSP0 is set to a value close to 0, for example, 10 km / h. Therefore, the fact that the vehicle speed is lower than VSP0 means that the vehicle is in a low vehicle speed region near the geared neutral point GNP in the power circulation mode.

If the vehicle speed VSP is lower than the predetermined vehicle speed VSP0, the routine proceeds to step S3, where it is determined whether or not the accelerator pedal depression amount Acc is 0, in other words, whether or not the accelerator pedal is released.

If the accelerator pedal depression amount Acc is 0, the process proceeds to creep torque control after step S4.

On the other hand, in step S2, the vehicle speed VSP becomes V
If it is not less than SP0, or if the accelerator pedal depression amount Acc is not 0 in step S3, the process proceeds to step S9 to execute normal shift control, and ends the process.

In step S4 for performing the creep torque control, the input shaft speed Ni and the output shaft speed Nco of the toroidal type continuously variable transmission 2 are read, and the current actual speed ratio i
c or the tilt angle φ is calculated. Here, ic = Ni / Nco (1).

Note that instead of using the gear ratio ic of the toroidal type continuously variable transmission 2, the unit gear ratio ii of the infinitely variable gear ratio continuously variable transmission can be used. However, at the geared neutral point GNP, the unit output shaft Since the number of revolutions of No. 6 is 0, it is difficult to grasp an accurate gear ratio. On the other hand, in the toroidal type continuously variable transmission 2, even at the geared neutral point GNP, when the engine (not shown) is operated, both the input side and the output side are rotating. Fluctuations can always be read accurately.

Next, at step S5, the target creep torque Tc is read. This target creep torque Tc is
For example, the position of the shift lever (not shown) is set to the forward position (D
Different values are read depending on whether the vehicle is in the range or the Ds range for sports running) or the reverse position (R range) and whether the brake is depressed.

The target creep torque Tc is, for example,
As shown in FIG. 7, when the vehicle is moving forward and the brake is not depressed, that is, when the brake signal is OFF, a large creep torque T3 is read. Also, when the brake is depressed, that is, when the brake signal is ON
, A small creep torque T1 is read.

Also, at the time of reversing, similarly, different target creep torques -T3 and -T1 are read in accordance with the operation state of the brake.

In step S6, a target step number St for generating the target creep torque Tc at the current gear ratio ic is calculated from the creep torque Tc map shown in FIG.

In FIG. 7, the bold solid line in the figure is a line at which the output torque To of the unit output shaft 6 becomes 0 and the vehicle can be stopped, and the creep torque Tc at this time becomes 0.

In the vicinity of the geared neutral point GNP, the creep torque Tc increases as T1, T2, T
3 and sequentially increases in the forward direction (forward direction, the same applies hereinafter), and To
The creep torque Tc increases in the negative direction (reverse direction, the same applies hereinafter) in the order of -T1, -T2, and -T3 as going downward in the figure from the line of "= 0".

The torque shift of the toroidal type continuously variable transmission is based on the creep torque Tc in the power circulation mode.
The positive direction is the Hi side of the speed ratio ic, and the negative direction is the Lo side of the speed ratio ic.

In step S7, the difference between the target step number St and the current step number S0 is calculated as the command step number Sx. In step S7, the command step number Sx is output to drive the step motor 36.

By the above control, in the vicinity of the geared neutral point GNP in the power circulation mode, while the current gear ratio ic is maintained, only the step motor 36 is driven with the number of steps STEP corresponding to the target creep torque Tc. Can be generated.

In FIG. 7, when the current speed ratio ic is at the geared neutral point GNP = ic0, the current step number S0 of the step motor 36 is Tc
At the point A intersecting with ic = ic0 on the line of = 0, if the target creep torque Tc = T1 is given at this time, the number of steps of the step motor 36 is driven up to Sx at the point B in the figure according to the target number of steps St. Will be done.

Therefore, the number of steps of the step motor 36 is driven in a decreasing direction, and in a normal operation state,
The spool 46S of the shift control valve 46 shown in FIG. 4 is displaced so as to be on the gear ratio Lo side (larger side), and a hydraulic pressure according to the opening amount of the supply port 46P is supplied to the oil chamber 30A. .

At this time, as shown in FIG.
A is pressurized with respect to the oil chamber 30B, and a force acts on the trunnion 23 on the right side in the figure in the upward direction indicated by the solid line.

When the shift lever (not shown) is in the forward position of the D range or the Ds range and the driver operates the brake to stop the vehicle at a predetermined brake torque, the rotation speed of the unit output shaft 6 Is 0.

In this case, the speed ratio ic of the continuously variable transmission 2
Will forcibly maintain the speed ratio ic0 corresponding to the geared neutral point GNP. That is, even if a force in the solid line direction in FIG. 3 is applied to the trunnion 23, the tilt angle of the power roller 20 does not change, and the creep torque Tc corresponding to the differential pressure between the oil chambers 30A and 30B of the hydraulic cylinder 30 is transmitted. Will be.

Of course, even when the gear ratio ic deviates from ic0, which is the geared neutral point GNP, an arbitrary creep torque Tc can be obtained in the same manner as described above. ｛4. Principle of creep torque control by shift control valve バ ル ブ Here, the principle of controlling the creep torque Tc of the infinitely variable transmission by the shift control valve 46 that controls the supply direction and flow rate of the hydraulic pressure to the hydraulic cylinder 30 will be described. Details will be described below.

When the number of steps of the step motor 36 is changed, the speed ratio ic of the toroidal type continuously variable transmission 2 is changed.
Changes the gear ratio ic by changing the tilt angle φ of the power roller 20 according to the characteristics shown in FIG. 5, but the geared neutral point G in the power circulation mode is changed.
In the vicinity of NP, the unit speed ratio ii becomes almost infinite and the vehicle is stopped. If the shift position (not shown) is in the travel range (D range or R range), the driver operates the brake to stop the vehicle. The state is maintained.

The geared neutral point G in the power circulation mode in which the vehicle is stopped by operating a brake (not shown)
Considering the relationship between the output torque To of the toroidal-type continuously variable transmission 2 and the number of steps STEP of the step motor 36 and the speed ratio ic around NP, the following is obtained.

Now, as shown in FIG.
c is geared neutral point GNP ic0
When the step number STEP of the step motor 36 is also at the no-load position S0 intersecting with the line of the output torque To = 0, the driver operates the brake with an arbitrary brake torque Tb (unit output shaft 6 conversion). I have.

In this stopped state, as shown in FIG. 8B, the output torque To is in a state of being balanced with the brake torque Tb, and the output torque To is reduced from + Tb (forward side) to −Tb.
If it is within the range of Tb (reverse side), the vehicle can maintain the stopped state.

On the other hand, in the toroidal type continuously variable transmission 2, as disclosed in Japanese Patent Application Laid-Open No. 8-338492 proposed by the applicant of the present invention, the power roller is controlled in accordance with the direction and magnitude of the input torque fluctuation. The torque shift characteristic in which the tilting angle φ of the power roller 20 changes and the gear ratio ic shifts due to the elastic deformation generated in the support member (the rotating shaft and the trunnion 23), the backlash of each part, and the axial displacement of the trunnion 23. There is.

Using this torque shift, the gear ratio ic
When the vehicle is stopped near the geared neutral point GNP, the creep torque Tc is adjusted by intentionally creating a torque shift state by adjusting the supporting force of the trunnion 23, that is, the oil pressure of the hydraulic cylinder 30. Can be adjusted.

That is, if the hydraulic pressure of one of the oil chambers 30A or 30B is increased with the gear ratio ic fixed, the power roller 20 does not tilt and the trunnion 2
3 cannot be displaced in the axial direction, and the above-mentioned torque shift state occurs due to the elastic deformation of the trunnion 23 in the axial direction or the change in play of each part, and the torque transmitted by the power roller 20 can be changed. By changing the hydraulic pressure, this transmission torque is changed to creep torque Tc.
It can be controlled as

That is, the oil chamber 30A of the hydraulic cylinder 30
Alternatively, the spool 46S of the shift control valve 46 for selectively supplying the line pressure PL to 30B is driven,
By controlling the amount of opening of the supply port 46P, the amount of axial displacement of the trunnion 23 is changed in a stopped state of the vehicle to change the transmission torque of the power roller 20, that is, the creep torque Tc.

Now, in FIG. 8A, as described above, the gear ratio is at the geared neutral point GNP with the gear ratio ic = ic0, and the number of steps ST of the step motor 36 is ST.
The EP is also at the no-load position S0 crossing the line of the output torque To = 0, and the driver operates the brake with an arbitrary brake torque Tb.

In this state, when the number of steps STEP of the step motor 36 is driven in the decreasing direction to Sf, FIG.
As shown in FIG.
By driving the spool 46S upward in the drawing so as to achieve the target gear ratio on the Lo side, the supply port 46P communicates with the oil chamber 30A, the drain port 46D communicates with the oil chamber 30B, and the opening of the supply port 46P opens. The oil pressure according to the amount is applied to the oil chamber 30
A.

As the oil pressure in the oil chamber 30A increases, the trunnion 23 attempts to displace in the axial direction upward in FIG. 3, but since the vehicle is in a stopped state, the rotational speed of the unit output shaft 6 becomes zero. The tilt angle φ of the power roller 20 cannot be changed from the position φ0 corresponding to the geared neutral point GNP, and the output torque To transmitted to the unit output shaft 6 is at the point A (= 0) in FIG. ) To the point Af to change the gear ratio ic to the geared neutral point ic.
While maintaining the output torque To at 0, the output torque To is increased to + Tb to generate the creep torque Tc in the forward direction, so that the creep torque Tc and the brake torque Tb can be in a state of being balanced.

After the stepping motor 36 stops at the step number STEP = Sf, the tilt angle φ of the power roller 20 is maintained at φ0 if the stopped state of the vehicle is maintained. The mechanical feedback mechanism including the precess cam 35 shown in FIG.
Is a position corresponding to the command value Sf of the step motor 36 and the amount of deformation of the trunnion 23, and can continuously generate the creep torque Tc.

That is, as shown in FIG. 8A, the number of steps STEP of the step motor 36 is changed from S0 to a point A where no load is applied at the geared neutral point GNP and a point Af where the brake torque Tb is balanced on the forward side. As shown in FIG. 8 (B), by changing the value up to the step number Sf corresponding to the brake torque Tb,
In accordance with the amount of deformation of the trunnion 23, that is, the magnitude of the torque shift, the creep torque Tc of the infinitely variable speed ratio transmission using the toroidal type continuously variable transmission 2 can be set to any value even at a low oil temperature. It is possible to control the time.

Further, as shown in FIG. 8A, Af in the figure where the creep torque Tc and the brake torque Tb are balanced.
At this point, when the brake is released, the vehicle moves from the point Af to the point C in the drawing and moves forward, and then stops at the point C where the output torque matches the running resistance.

This is because the engine (not shown) performs idle speed control, and increases the engine load according to the increase in the engine load, that is, the creep torque Tc, while increasing the engine torque according to the decrease in the engine load. In order to maintain a predetermined idle speed, the creep torque T from the point A to the point Af in the figure is increased.
The engine load increases as c increases, and, for example, the fuel injection amount of the engine increases. On the other hand, when going from point Af to point C in the figure, the creep torque Tc decreases and the engine increases. The load is also reduced, and the fuel injection amount of the engine is reduced to a value where the driving resistance = load and the engine output are balanced.

In FIG. 8, an example in which the creep torque Tc is controlled within the range of the brake torque Tb has been described. However, as shown in FIG. 7, the relationship between the creep torque Tc and the number of steps STEP is set in a map or the like in advance. , Any required creep torque Tc = 0 to |
Step motor 3 to step number STEP according to T3 |
6, the desired creep torque Tc can be obtained easily and reliably.

On the other hand, when the creep torque Tc is generated to the retreat side under the same conditions as above, FIG.
As shown in (B), by changing the step number STEP of the step motor 36 in the range from S0 to Sr, the step number STEP can be arbitrarily set in the range from the point A to the point Ar in balance with the brake torque -Tb as in the forward side in the drawing. Can be set to the creep torque Tc. ｛5. Operation and Effect of Creep Torque Control Using Shift Control Valve As shown in FIG. 7, near the geared neutral point GNP, the creep torque Tc = 0 to | Tn | and the number of steps STEP of the step motor 36 or the driving position Is set in advance, and only by driving the step motor 36 to the number of steps STEP corresponding to the desired creep torque Tc, the creep torque can be easily and reliably utilized utilizing the torque shift of the toroidal type continuously variable transmission 2. Tc can be controlled.

The creep torque Tc is controlled by changing the oil pressure at which the hydraulic cylinder 30 supports the trunnion 23 while maintaining the gear ratio ic, so as to obtain an arbitrary creep torque Tc by setting a torque shift state. In this control, since the shift control valve 46 is controlled to adjust the supply direction and flow rate of the pressure oil, it is possible to constantly control the transmission torque stably regardless of the oil temperature.

On the other hand, during normal running, a command for the target gear ratio based on the number of steps STEP of the step motor 36,
Due to the feedback of the actual gear ratio by the mechanical feedback mechanism, accurate gear ratio control can always be performed, and in particular, the viscosity of the hydraulic oil increases, as in the case of the differential pressure control by the pressure control valve as in the conventional example. The control of the gear ratio and the control of the creep torque at low oil temperature are reliably prevented from becoming unstable, and the control of the creep torque in the vicinity of the geared neutral point GNP in the power circulation mode;
Speed ratio control during traveling can be performed with high accuracy at all times in a stable manner, and the operability of the continuously variable transmission with an infinite speed ratio can be greatly improved.

The control of the creep torque Tc is performed by controlling the geared neutral point GN as shown in FIG.
Using the creep torque Tc (output torque To) near P as a parameter, the step number ST of the step motor 36
Since only a map or the like in which the relationship between the EP and the gear ratio ic is set in advance is used, it is possible to simplify the control contents and reduce the manufacturing cost, and in particular, the oil temperature at which the accuracy of the feedback control decreases is extremely low. In some cases, it is possible to obtain an arbitrary creep torque Tc by open loop control.

FIGS. 9 to 12 show a second embodiment.
In addition to the configuration of the first embodiment, a differential pressure sensor 1 for detecting a differential pressure ΔP between oil chambers 30A and 30B of a hydraulic cylinder 30
10 and feedback control is performed so that the pressure difference ΔP according to the creep torque Tc is obtained.
Other configurations are the same as those of the first embodiment.

As shown in FIG. 10, the control of the creep torque Tc is as follows. First, in step S10, the vehicle speed VSP detected by the vehicle speed sensor 83 shown in FIG.
1, the rotation speed Ni of the unit input shaft, the rotation speed Nco on the output side of the toroidal type continuously variable transmission 2 detected by the output shaft rotation speed sensor 82, and the accelerator pedal depression Acc detected by the accelerator opening sensor 86. Read.

Then, in step S11, the vehicle speed VSP
Is lower than a predetermined speed VSP0,
If it is less than the predetermined value, the process proceeds to step S12 to determine whether the accelerator pedal depression amount Acc is zero.

If the accelerator pedal depression amount Acc is 0, the process proceeds to the creep torque control from step S13.

In step S11, the vehicle speed is VSP
If the value is equal to or greater than 0, or if the accelerator pedal depression amount Acc exceeds 0 in step S12, the process proceeds to step S12.
Proceeding to 19, normal shift control is performed.

Step S13 for performing creep torque control
Then, the input shaft speed Ni and the output shaft speed Nco of the toroidal type continuously variable transmission 2 are read, and the current actual speed ratio ic or tilt angle φ is calculated.

Next, at step S14, the target creep torque Tc is read. This target creep torque Tc
As in the first embodiment, for example, a value set in advance according to whether the position of a shift lever (not shown) is in a forward position (D range) or a reverse position (R range) is read. .

In step S15, based on the map shown in FIG. 11, the target differential pressure ΔP0 for generating the target creep torque Tc is calculated from the output torque To, ie, the creep torque Tc and the current gear ratio ic. The target differential pressure ΔP0 is represented by ΔP0 = Phi−Plo (2) where Plo is the oil pressure of the oil chamber 30A and Phi is the oil pressure of the oil chamber 30B.

In the map of FIG. 11, the solid line in the figure is a map or a function for setting the target differential pressure ΔP0 corresponding to the output torque To when the speed ratio ic is the geared neutral point ic0. When it is smaller than ic0, a line of ic <ic0 shown by a chain line in the figure is used, and conversely, the gear ratio ic
Is smaller than the geared neutral point ic0, a line of ic> ic0 indicated by a dashed line in the figure is used. In addition, maps other than ic = ic0 are set for each speed ratio, not shown, in addition to the above.

Next, in step S16, the current differential pressure ΔP (= Phi−Pl) is detected from the differential pressure sensor 110 shown in FIG.
o) is read, and in step S17, the number of steps STEP of the step motor 36 according to the deviation between the target differential pressure ΔP0 and the current differential pressure ΔP is determined by a map or function set in advance as shown in FIG. Alternatively, it is calculated as the command step number Sx based on the gain.

The map shown in FIG. 12 shows that the stepping motor 3 is located near the geared neutral point GNP.
6 shows the relationship between the number of steps STEP (position) and the differential pressure ΔP. When the differential pressure ΔP increases in the positive direction, the number of steps STEP also increases. Conversely, when the differential pressure ΔP decreases in the negative direction, The step number STEP is also set to decrease in the negative direction, and when the differential pressure ΔP increases in the positive direction, the oil chamber 3
When the differential pressure ΔP decreases in the negative direction while the hydraulic pressure Phi of 0B increases, the hydraulic pressure Plo of the oil chamber 30A increases.

In step S17, the target differential pressure Δ
The number of command steps S based on the difference between P0 and the actual differential pressure ΔP
x is output to drive the step motor 36.

By the above control, in the vicinity of the geared neutral point GNP in the power circulation mode, based on the target differential pressure ΔP, a differential pressure ΔP corresponding to the target creep torque Tc is maintained while maintaining the current gear ratio ic. As in the first embodiment, the step motor 36 is feedback-controlled, and the spool 46S of the shift control valve 46 is displaced so that an arbitrary creep torque Tc is reduced near the geared neutral point GNP in the power circulation mode. Irrespective of this, it can be generated with high accuracy.

In this case, the output torque To (= creep torque T) near the geared neutral point GNP
The relationship between c) and the target differential pressure ΔP0 is previously set in a map or the like using the gear ratio ic as a parameter, and the target differential pressure ΔP0 near the geared neutral point GNP is further set.
By setting the relationship between 0 and the number of steps STEP in a map or the like in advance, and performing feedback control so that the detected differential pressure ΔP matches the target differential pressure ΔP0, more accurate creep torque control can be realized.

FIGS. 13 to 15 show a third embodiment. Instead of the map of the gear ratio ic and the number of steps STEP using the creep torque Tc of the first embodiment as a parameter, the gear ratio is infinitely variable. The creep torque Tc is controlled based on the torque ratio t of the transmission.
Other configurations are the same as those of the first embodiment.

The output torque To of the continuously variable transmission with infinite transmission ratio
Is the output torque of the toroidal type continuously variable transmission 2, and the output torque To of the unit output shaft 6 and the unit input shaft 1a
Assuming that the ratio of the input torque Ti is the torque ratio t, To = Ti × t (3)

As shown in FIG. 15, the torque ratio t becomes positive in the forward direction and negative in the reverse direction with the geared neutral point GNP at the speed ratio ic = ic0 as a boundary.
The absolute value is maximum at the geared neutral point GNP.

Therefore, the desired creep torque T
Since c is synonymous with the output torque To of the continuously variable transmission with infinite transmission ratio, it can be realized by controlling the torque ratio t of the continuously variable transmission with infinite transmission ratio to the input torque Ti.

On the other hand, the input torque Ti of the infinitely variable speed ratio continuously variable transmission is equal to the engine torque Te (not shown), and various methods have been conventionally proposed for detecting and estimating the engine torque Te. As shown in FIG. 13, the engine torque Te is obtained from the fuel injection amount Tp read from the engine controller 85,
It is known that the fuel injection amount Tp is treated as the engine torque Te. Here, the fuel injection amount Tp is treated as having the same meaning as the engine torque Te.

Note that the engine control controller 85
Perform idle speed control when the vehicle is stopped,
The engine speed Ne is controlled to be a predetermined value. Therefore, when the creep torque Tc is changed, the engine load also changes. Therefore, the fuel injection amount Tp is changed to maintain the predetermined idle speed. ing.

Hereinafter, the control of the creep torque Tc will be described in detail with reference to the flowchart of FIG.

First, in step S20, the vehicle speed sensor 8
3, the vehicle speed VSP detected by the input shaft speed sensor 81, the unit input shaft speed Ni (= engine speed Ne) detected by the input shaft speed sensor 81, and the output of the toroidal type continuously variable transmission 2 detected by the output shaft speed sensor 82. Side rotational speed Nco, accelerator pedal depression amount A detected by accelerator opening sensor 86
The fuel injection amount Tp is read from the engine controller 85 in addition to cc.

Then, in steps S21 and S22, it is determined whether the vehicle speed VSP is less than the predetermined value VSP0 and the accelerator pedal depression amount Acc is 0, and the vehicle speed is determined to be VSP.
If the value is less than 0 and the accelerator pedal depression amount Acc is 0, the process proceeds to the creep torque control after step S23. If not, the process proceeds to step S30 to perform the normal shift control, and ends the process.

Step S23 for performing creep torque control
Then, the input shaft speed Ni and the output shaft speed Nco of the toroidal type continuously variable transmission 2 are read, and the current actual speed ratio ic or tilt angle φ is calculated.

Next, in step S24, the fuel injection amount Tp read in step S20 is set as the input torque Ti of the infinitely variable speed ratio transmission.

In step S25, the target creep torque Tc is read in the same manner as in the first embodiment, and in step S26, the above-described step S2 is performed.
The target torque ratio tr is calculated as tr = Tc / Ti (3 ′) from the input torque Ti obtained in step 4 and the target creep torque Tc.

Next, in step S27, the target gear ratio i corresponding to the target torque ratio tr is obtained from the map shown in FIG.
cr is obtained, and in step S28, the step number STEP of the step motor 36 corresponding to the target speed ratio icr is calculated as the command step number Sx from the normal speed ratio control map shown in FIG. 5, and in step S29, The command step number Sx is output, and the step motor 36 is driven.

By the above control, in the vicinity of the geared neutral point GNP in the power circulation mode, the step motor 36 is controlled so that the target torque ratio tr corresponding to the target creep torque Tc is maintained while maintaining the current gear ratio ic.
Is controlled, the spool 46S of the shift control valve 46 is displaced similarly to the first embodiment, and an arbitrary creep torque Tc can be adjusted with high accuracy in the vicinity of the geared neutral point GNP in the power circulation mode regardless of the oil temperature. Can be generated.

Further, by using the fuel injection amount Tp of the engine as the input torque Ti of the continuously variable transmission with infinite transmission ratio, it is possible to more accurately control the creep torque Tc which causes the engine load to fluctuate. It is possible to improve the control accuracy of the torque Tc.

FIG. 16 shows the fourth embodiment. The map of the torque ratio t and the gear ratio ic shown in FIG. 15 of the third embodiment is obtained by using the output torque To and the gear ratio using the input torque Ti as a parameter. Instead of the map of the ratio ic, the other configuration is the same as that of the third embodiment.

In this case, the calculation of the target torque ratio tr in step S26 shown in the flowchart of FIG. 14 is omitted, and the input torque Ti obtained in step S24 and the target creep torque Tc read in step S25 are output. It is assumed that the torque is To and the target speed ratio ic is obtained from the map of FIG.
r can be obtained, and the calculation load on the shift control controller 80 can be reduced.

FIG. 17 shows a fifth embodiment, in which a map of the output torque To and the speed ratio ic is set using the input torque Ti shown in FIG. 16 of the fourth embodiment as a parameter, and the speed ratio ic is set as a parameter. Instead of the map of the output torque To and the input torque Ti, the other configuration is the same as that of the fourth embodiment.

In this case, the relationship between the input torque Ti and the output torque To corresponding to each speed ratio is not linear, and until the input torque Ti increases to some extent, the friction (internal loss) of each part of the transmission is reduced by the input torque Ti. Is relatively large with respect to the speed ratio ic.
The relationship of o (= creep torque Tc) can be made to substantially match the actual value.

Accordingly, in the flowchart of FIG. 14, the input torque Ti obtained in step S24 is
The target creep torque Tc read in step S25 is set as the output torque To, and the target gear ratio i is obtained from the map of FIG.
cr can be obtained, and the creep torque T can be more accurately determined.
c can be controlled.

In the above embodiment, an example is shown in which the position of the power circulation mode clutch 9 is disposed between the counter gear 3b and the carrier 5b. However, the power circulation mode clutch 9 is moved from the unit input shaft 1a to the unit. Output shaft 6
Can be disposed at any position between the transmission output gear 7 and the unit output shaft 6 (not shown), for example, disposed between the ring gear 5c and the unit output shaft 6.
a and the gear 3a of the fixed transmission 3 or may be interposed on the continuously variable transmission output shaft 4 connected to the sun gear 5a.

The shift control valve 46 is connected to the step motor 36 and the feedback mechanism via the speed change link 37.
A spool 46S driven directly by an actuator such as a step motor 36 by eliminating the transmission link 37;
A valve body that is relatively displaceable with respect to the spool 46S may be connected to the feedback link 38.

In the above embodiment, the target creep torque Tc is set to a different value according to the operation state of the brake torque. However, the target creep torque Tc is further changed as appropriate according to the road surface gradient and the state of the vehicle. Is also good.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram of an infinitely variable speed ratio continuously variable transmission showing one embodiment of the present invention.

FIG. 2 is a control conceptual diagram of a continuously variable transmission with an infinite speed ratio.

FIG. 3 is a conceptual diagram of a toroidal type continuously variable transmission.

FIG. 4 is a schematic circuit diagram illustrating a configuration of a hydraulic control device.

FIG. 5 is a map showing the relationship between the number of steps STEP of the step motor and the speed ratio ic of the toroidal-type continuously variable transmission, and mainly used for normal speed ratio control.

FIG. 6 is a flowchart illustrating an example of creep torque control performed by a shift control controller.

FIG. 7 shows a gear ratio ic near the geared neutral point GNP with the output torque To as a parameter.
4 is a map showing a relationship between the step number and the step number STEP of the step motor.

FIG. 8 is an explanatory diagram showing creep torque Tc control in the vicinity of a geared neutral point GNP;
7A is a map showing the relationship between the speed ratio ic and the number of steps STEP of the step motor using the output torque To as a parameter, and FIG. 7B is a graph showing the relationship between the output torque To and the number of steps STEP.

FIG. 9 shows a second embodiment, and is a control conceptual diagram of a continuously variable transmission with an infinite speed ratio.

FIG. 10 is a flowchart illustrating an example of a creep torque control performed by the shift control controller.

FIG. 11 shows an output torque T in the vicinity of a geared neutral point GNP with the gear ratio ic as a parameter.
A map showing the relationship between o and the differential pressure ΔP.

FIG. 12 is a map showing the relationship between the differential pressure ΔP and the number of steps STEP in the vicinity of the geared neutral point GNP.

FIG. 13 is a control conceptual diagram of a continuously variable transmission with an infinite speed ratio according to the third embodiment.

FIG. 14 is a flowchart illustrating an example of creep torque control performed by a shift control controller.

FIG. 15 is a map showing a relationship between the torque ratio t of the infinitely variable speed transmission and the speed ratio ic of the toroidal type continuously variable transmission.

FIG. 16 is a map showing the fourth embodiment and showing the relationship between the output torque To and the speed ratio ic of the toroidal-type continuously variable transmission, with the input torque Ti as a parameter.

FIG. 17 shows a fifth embodiment, in which an input torque Ti is set using a speed ratio ic of a toroidal type continuously variable transmission as a parameter.
4 is a map showing the relationship between the output torque To.

FIG. 18 shows a speed ratio (CVT) of a continuously variable transmission with an infinite speed ratio.
2 is a map showing the relationship between the ratio of the transmission and the reciprocal of the unit speed ratio (IVT ratio).

[Explanation of symbols]

 1a, 1b Unit input shaft 2 Continuously variable transmission 3 Constant transmission 4 Continuously variable transmission output shaft 5 Planetary gear mechanism 6 Unit output shaft 9 Power circulation mode clutch 10 Direct connection mode clutch 20 Power roller 21 Input disk 22 Output disk 23 Trunnion Reference Signs List 30 hydraulic cylinder 30A, 30B oil chamber 31 piston 35 precess cam 36 step motor 37 shift link 38 feedback link 46 shift control valve 46S spool 60 manual valve 80 shift control controller 110 differential pressure sensor

Claims (9)

[Claims] 1. A toroidal type continuously variable transmission for continuously changing a gear ratio by tilting a power roller sandwiched between an input / output disk and a constant transmission are connected to a unit input shaft, respectively. A continuously variable transmission having an infinitely variable transmission ratio in which the output shafts of a fixed type continuously variable transmission and a constant transmission are connected to a unit output shaft via a planetary gear mechanism, a power circulation mode clutch and a direct coupling mode clutch, and the power roller is supported. A transmission control device for a continuously variable transmission having an infinitely variable gear ratio, comprising: a hydraulic cylinder connected to a trunnion to be driven; and a transmission control device for controlling a hydraulic pressure of the hydraulic cylinder. A shift control valve for controlling a hydraulic pressure and a supply direction to be changed, an actuator for driving the shift control valve, and an actual speed ratio of the toroidal-type continuously variable transmission. A mechanical feedback means for feeding back to the speed change control valve, and the operating condition detecting means for detecting an operating condition of the vehicle, when the operating condition matches the state set in advance,
And a shift command means for calculating a drive amount of the actuator based on a preset creep torque and a gear ratio of the toroidal-type continuously variable transmission, and controlling the actuator based on the drive amount. Transmission control device for a continuously variable transmission with an infinitely variable transmission ratio. 2. The driving state detecting means comprises: means for detecting a vehicle speed; and means for detecting an amount of depression of an accelerator pedal, wherein the shift command means includes: the detected vehicle speed is less than a predetermined vehicle speed; The transmission control device for a continuously variable transmission according to claim 1, wherein when the accelerator pedal depression amount is zero, it is determined that the state matches a preset state. 3. The shift command means, wherein a command value to the actuator is obtained by correcting a drive position of an actuator corresponding to a geared neutral point in a no-load state with a value corresponding to the creep torque. The shift control device for a continuously variable transmission with an infinite gear ratio according to claim 1. 4. The speed change command means includes: a speed change ratio of a toroidal-type continuously variable transmission; and a differential pressure setting means for setting a differential pressure of the hydraulic cylinder in accordance with the creep torque; and an actuator in accordance with the differential pressure. The shift control device for an infinitely variable speed ratio transmission according to claim 3, wherein the command value is set as follows. 5. The speed change command unit includes: an input torque calculation unit configured to calculate an input torque of a continuously variable transmission with an infinite gear ratio;
A torque ratio calculating means for calculating a ratio of an input torque and an output torque of the infinitely variable speed ratio transmission as a torque ratio,
The shift control device for an infinitely variable speed ratio transmission according to claim 3, wherein a command value for an actuator is set based on a torque ratio obtained by using the creep torque as an output torque. 6. The speed change command means includes an input torque calculation means for calculating an input torque of a continuously variable transmission with an infinite gear ratio;
Based on the output torque corresponding to the creep torque,
4. The transmission according to claim 3, further comprising a drive amount setting unit that determines a target speed ratio of the toroidal-type continuously variable transmission and sets a drive amount of the actuator so as to achieve the target speed ratio. Transmission control device for a continuously variable transmission with infinite ratio. 7. The speed change command means sets a relationship between an output torque of the toroidal type continuously variable transmission and a speed ratio of the toroidal type continuously variable transmission in advance using the input torque of the infinitely variable speed ratio continuously variable transmission as a parameter. The shift control device for an infinitely variable speed ratio transmission according to claim 6, further comprising a set map. 8. The speed change command means includes a map in which a relationship between an input torque and an output torque of a toroidal-type continuously variable transmission is preset using a speed ratio of an infinitely variable speed ratio transmission as a parameter. The speed change control device for an infinitely variable speed ratio continuously variable transmission according to claim 6. 9. The speed change controlling means according to claim 6, wherein said output torque corresponding to said creep torque is based on a relationship between an input torque and an output torque preset in consideration of an internal loss of a continuously variable transmission with an infinite gear ratio. 9. The speed change control device for an infinitely variable speed ratio transmission according to claim 8, wherein a target speed ratio of the toroidal type continuously variable transmission is set from the following.