JPH11247964A - Transmission control device for transmission ratio infinite continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To control torque transmitting force with high accuracy, and to maintain a geared neutral point of a motive power circulating mode by controlling the differential pressure of a piston of a hydraulic cylinder capable of changing torque transmitting force of a power roller by first/second pressure control valves to control the reduction of supply pressure. SOLUTION: A power roller of this toroidal continuously variable transmission is connected to the lower end to a hydraulic cylinder through a trunnion, torque transmitting force is continuously controlled by adjusting the difference of pressures communicated with the oil chamber 30A and oil chamber 30B of the cylinder. A (+) torque control valve 40 and a (-) torque control valve 45 composed of pressure control valves, are connected to a line pressure circuit 101 of an oil pressure control device. These control valves 40, 45 respectively supply control pressure to the oil chambers 30A, 30B of the hydraulic cylinder according to output pressure from a (+) torque solenoid 50 and a (-) torque solenoid 55 to thereby control the torque transmitting force with high accuracy.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to an improvement of a shift control device for a continuously variable transmission with an infinite speed ratio, which is employed in a vehicle or the like.

[0002]

2. Description of the Related Art A belt-type or toroidal-type continuously variable transmission is conventionally known as a vehicle transmission. In order to further expand the shift range of such a continuously variable transmission, a continuously variable transmission is known. There is known a continuously variable transmission with an infinitely variable transmission ratio, which is capable of controlling the transmission ratio to infinity by combining a constant transmission and a planetary gear mechanism, such as Japanese Patent No. 2557269.

[0003] This is a full toroidal type continuously variable transmission capable of continuously changing the gear ratio to a unit input shaft of a continuously variable transmission having an infinitely variable gear ratio connected to an engine, and a constant transmission (reduction gear). Are connected in parallel and these output shafts are selectively connected by a planetary gear mechanism.The output shaft of the continuously variable transmission is connected to the sun gear of the planetary gear mechanism. It is connected to the carrier of the planetary gear mechanism via the clutch.

[0004] The continuously variable transmission output shaft connected to the sun gear
While being connected to the unit output shaft, which is the output shaft of the infinitely variable speed ratio transmission, via the direct connection mode clutch, the ring gear of the planetary gear mechanism is also connected to the unit output shaft.

In such a continuously variable transmission with an infinite transmission ratio,
As shown in FIG. 9, while the power circulation mode clutch is connected and the direct connection mode clutch is disconnected, the unit speed ratio (IVT ratio in the diagram) is changed according to the speed ratio difference between the continuously variable transmission and the fixed transmission. A power circulation mode in which the unit input shaft rotation speed / unit output shaft rotation speed is continuously changed from a negative value to a positive value including infinity (= geared neutral point), and a power circulation mode clutch. On the other hand, the direct connection mode in which the direct connection mode clutch is connected and the shift control is performed according to the speed ratio of the continuously variable transmission can be selectively used while the disconnection is performed.

In a full toroidal type continuously variable transmission, a power roller held between an input / output disk is driven by a hydraulic actuator controlled by a bleed-off circuit.

In this hydraulic actuator, two systems of hydraulic pumps communicate with respective oil chambers in opposed oil chambers for applying oil pressure to the front and back of the piston, and a flow control valve is provided downstream of each oil chamber. The power roller is tilted by changing the pressure difference between the front and the back applied to the piston by the downstream flow control valve through the two oil chambers as the entire discharge flow rate from the hydraulic pump.

[0008]

However, in the above conventional example, the two oil chambers provided on the front and back of the piston are respectively provided with an upstream hydraulic pump and a downstream flow control valve via oil passages. Because of the communication, the hydraulic pressure applied to the oil chamber increases by the resistance of the oil passage from the hydraulic pressure set by the flow control valve, and the hydraulic pump discharge flow rate is low in the low rotation range, and the hydraulic oil viscosity is low, and the high oil temperature is low. At times, the effect of the oil passage resistance is small, but at high rotations where the discharge flow rate of the hydraulic pump is large or at a low oil temperature where the viscosity of the hydraulic oil increases, the influence of the oil passage resistance becomes large.

Here, the oil chambers on the front and back of the piston are controlled hydraulically by independent hydraulic circuits. Therefore, if the efficiency of the two hydraulic pumps and the oil path resistance are completely the same, the differential pressure between the front and back of the piston is different. Control can be performed accurately, but in reality, due to variations in the efficiency of the hydraulic pump and errors in the dimensional tolerances of the cross-sectional area, length, or curvature of the oil passage, the two hydraulic circuits are controlled by the flow control valve. Even if the settings are the same,
The hydraulic pressure applied to the oil chamber does not match due to errors in the oil passage resistance and the discharge flow rate, and there has been a problem that the differential pressure between the front and back of the piston cannot be accurately controlled.

In particular, in a continuously variable transmission with an infinite transmission ratio,
As shown in FIG. 9, even if an attempt is made to maintain a geared neutral point at which the unit speed ratio becomes infinite, the vehicle is stopped by setting the differential pressure between the front and back of the piston to 0 and the transmission torque of the power roller to 0. Even if an attempt is made to make the difference, the pressure difference will not be completely zero due to the erroneous calculation of the oil passage resistance, and the power roller will transmit torque to the forward side or the reverse side in accordance with the error in the pressure difference. Therefore, when the vehicle deviates from the geared neutral point and a creep torque is generated in a traveling direction opposite to the driver's intention, there is a problem that an uncomfortable feeling is given.

Further, in the above conventional example, since two hydraulic circuits are required, there are problems that the number of parts is increased, the structure is complicated, and the manufacturing cost is increased.

SUMMARY OF THE INVENTION The present invention has been made in view of the above-mentioned problems, and in addition to making it possible to control the torque transmission force of the power roller with high accuracy, to maintain the geared neutral point in the power circulation mode, The purpose is to promote cost reduction.

[0013]

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. The output shafts of the continuously variable transmission and the constant transmission are connected to the unit input shaft, and the planetary gear mechanism,
A continuously variable transmission having an infinitely variable transmission ratio connected to a unit output shaft via a power circulation mode clutch and a direct connection mode clutch, and a torque transmission force acting on the power roller being provided between a first oil chamber and a second oil chamber. A shift control device for an infinitely variable gear ratio transmission having a hydraulic cylinder that can be changed via a piston that responds to pressure and a shift control unit that controls a differential pressure of the hydraulic cylinder in accordance with a driving state of the vehicle. The shift control means includes first and second pressure control valves that communicate with the first and second oil chambers and that can continuously reduce the supply pressure from the hydraulic pressure source.

In a second aspect based on the first aspect, the shift control means supplies a signal pressure to the first and second pressure control valves in accordance with a differential pressure command value, respectively. A second signal pressure supply unit, wherein when the differential pressure command value to the first or second pressure control valve is a minimum value, the first or second pressure control valve is connected to the first and second pressure control valves. The output ports which respectively communicate with the oil chambers are connected to the drain.

In a third aspect based on the first aspect, the shift control means supplies the first and second pressure control valves with a signal pressure in accordance with a differential pressure command value. A second signal pressure supply unit, wherein when the differential pressure command value to the first or second pressure control valve is a minimum value, the first or second pressure control valve is connected to the first and second pressure control valves. The output ports respectively communicating with the oil chambers of the first and second oil chambers communicate with the supply pressure side.

In a fourth aspect, the invention is directed to the second or third aspect.
In the invention, the first and second signal pressure supply means are constituted by solenoids, and set the differential pressure command value to a minimum value when no power is supplied.

In a fifth aspect based on the first aspect, the first and second pressure control valves receive supply pressure from a single hydraulic source, and control the output port control pressure and supply pressure. The differential pressure can be changed continuously.

In a sixth aspect based on the first aspect, the first and second pressure control valves allow the drain side to communicate with each other and reduce the valve opening pressure to a predetermined extremely low pressure downstream thereof. The pressure-holding valve set in is set.

In a seventh aspect, the invention is directed to the second or third aspect.
In the invention, one of the first and second signal pressure supply means variably controls the signal pressure in accordance with the differential pressure command value, and the other sets the signal pressure to a minimum value.

[0020]

According to the first invention, the differential pressure of the piston of the hydraulic cylinder is controlled by the first and second pressure control valves for controlling the supply pressure to be reduced, so that the differential pressure of the piston is controlled with high precision. However, the torque transmission force of the power roller can be performed with high accuracy, and at the geared neutral point in the power circulation mode, the differential pressure can be reliably set to 0, and the hydraulic cylinder and oil path resistance and the like can be set. Regardless of manufacturing variations, the geared neutral can be maintained with high accuracy by setting the torque transmission force of the power roller to 0, and the torque can be reduced in the direction opposite to the traveling direction intended by the driver, as in the conventional example. The vehicle can be reliably stopped and the vehicle can be maintained in a stopped state, and the operability and reliability of the infinitely variable speed continuously variable transmission using the toroidal type continuously variable transmission can be greatly improved. A.

According to a second aspect of the present invention, the first and second pressure control valves are configured to control a control pressure from an output port in accordance with a signal pressure from first and second signal pressure supply means in accordance with a differential pressure command value. And when the differential pressure command value is the minimum value, the first or second oil chamber is connected to the drain, so that the differential pressure command values of the first and second signal pressure supply means are set to the minimum value. This ensures that the differential pressure of the piston can be set to zero, the transmission torque of the power roller can be set to zero, and the geared neutral point can be maintained.

In a third aspect of the present invention, the first and second pressure control valves are configured to control the control pressure from an output port in accordance with a signal pressure from the first and second signal pressure supply means in accordance with a differential pressure command value. And the supply pressure is supplied to the first or second oil chamber when the differential pressure command value is the minimum value, so that the differential pressure command value of the first and second signal pressure supply means is set to the minimum value. With this setting, the differential pressure of the piston can be reliably set to 0, and the transmission torque of the power roller can be set to 0 to maintain the geared neutral point.

According to a fourth aspect of the present invention, the first and second signal pressure supply means are constituted by solenoids, and when no power is supplied, the differential pressure command value is set to a minimum value or a value equal to the minimum value. If the power is turned off due to a failure such as a disconnection, the signal pressure becomes the minimum value or a value equal to the minimum value, the piston differential pressure can be reliably set to 0, and the transmission torque of the power roller is set to 0. Therefore, even if a failure such as a disconnection occurs during traveling, a sudden shift or torque change can be prevented to ensure fail-safe.In a geared neutral where the transmission torque of the power roller is set to 0, the solenoid must be turned off. Since the power can be supplied, the power consumption in the idling state of the engine can be reduced, and downsizing of the generator and the battery can be promoted.

In the fifth aspect, the first and second pressure control valves can continuously change the differential pressure between the control pressure of the output port and the supply pressure by the supply pressure from a single hydraulic pressure source.
It is possible to accurately control the differential pressure of the piston irrespective of variations such as manufacturing errors while reducing the number of parts by reducing the number of parts by using a single hydraulic pressure source such as a hydraulic pump.

According to the sixth aspect of the present invention, the drain side of the first and second pressure control valves is communicated with each other, and a pressure keeping valve whose valve opening pressure is set to a predetermined extremely low pressure is interposed downstream thereof. By doing so, it is possible to prevent air from being mixed into the first and second oil chambers and to ensure control accuracy of the hydraulic cylinder.

According to a seventh aspect of the present invention, the first and second signal pressure supply means are configured such that one of the first and second signal pressure supply means variably controls the signal pressure in accordance with the differential pressure command value and the other sets the signal pressure to a minimum value. By controlling only one of them, the differential pressure control of the piston can be performed with high accuracy, and the control can be simplified.

[0027]

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

FIGS. 1 to 4 show an example in which a half-toroidal type continuously variable transmission 2 is used to form a continuously variable transmission with an infinite speed ratio.

FIG. 1 schematically shows a continuously variable transmission with an infinitely variable transmission ratio. A unit input shaft 1 of the continuously variable transmission with infinitely variable transmission ratio is connected to an engine. Continuously variable transmission 2 and a constant transmission 3 (reduction gear) composed of gears 3a and 3b are connected in parallel, and these output shafts 4 and 3c are connected by a planetary gear mechanism 5. The output shaft 4 of the continuously variable transmission 2 is connected to the sun gear 5a of the planetary gear mechanism 5 and the output shaft 3c of the constant transmission (reduction gear) 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 is connected to a unit output shaft 6 which is an output shaft of an infinitely variable speed ratio transmission via a direct connection mode clutch 10, while a planetary gear mechanism. The fifth ring gear 5c is also coupled to the unit output shaft 6.

A transmission output gear 7 is provided on the unit output shaft 6. The transmission output gear 7 meshes with a final gear 12a of the differential gear 8, and has a predetermined total reduction ratio.
The driving force is transmitted to the drive shafts 11a and 11b connected to the drive shafts 11a and 11b.

As shown in FIG. 1, the continuously variable transmission 2 is a double-cavity toroidal type in which a power roller 20 is sandwiched and pressed by two sets of input disks 21 and output disks 22, respectively. As shown in FIGS. 3 and 4, the trunnion 2 is coupled at its lower end to a hydraulic cylinder 30 so as to be displaceable in the axial direction and rotatable around the axis.
3 is provided at the lower end of the trunnion 23 with a precess cam 35 for feeding back a tilt angle, that is, an actual gear ratio, to a shift control valve 46 described later.

When the input disk 21 rotates as shown in FIGS. 3 and 4, the torque transmission force of the power roller 20 is reduced by increasing the oil pressure of the oil chamber 30A, while the oil chamber 30B is rotated.
, The torque transmission force of the power roller 20 increases, and the torque transmission force is continuously controlled by adjusting the differential pressure between the oil chambers 30A and 30B.

Therefore, the oil pressure of the oil chamber 30A is Pdec,
Assuming that the oil pressure in the oil chamber 30B is Pinc, the differential pressure ΔP applied to the piston 31 is ΔP = Pinc−Pdec, and this differential pressure ΔP is the torque transmitting force of the toroidal type continuously variable transmission 2.

Here, since the torque transmission force of the toroidal type continuously variable transmission 2 is considered to be positive from the engine side to the sun gear 5a side of the planetary gear mechanism 5, the torque transmission force of the entire transmission is in this case. As the driving torque increases, in the forward state of the power circulation mode, the driving torque increases as the differential pressure ΔP of the piston 31 becomes negative, that is, as the differential pressure ΔP decreases (however, the absolute value increases). The opposite is true.

However, the differential pressure Δ that does not match the transmission torque
When the piston 31 is displaced by giving P, the trunnion 23
Is displaced in the axial direction, so that the power roller 20 is displaced from the neutral point (a position where the rotation axis of the power roller 20 intersects with the rotation axis of the input / output disks 21 and 22 as shown in FIGS. 3 and 4). When the piston 31 is displaced by increasing the differential pressure ΔP, the power roller 20 tilts the speed ratio of the toroidal type continuously variable transmission to the Hi side. Also,
When the increased differential pressure ΔP is reduced to match the transmission torque, the power roller returns to the neutral position, and the gear shift stops.

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 composed of a microcomputer.
Output from the input shaft speed sensor 81 for detecting the engine speed Nt (= engine speed Ne) of the continuously variable transmission output shaft 4
, The output from the continuously variable transmission output shaft rotation speed sensor 82 for detecting the rotation speed No., the output from the vehicle speed sensor 83 for detecting the vehicle speed VSP from the rotation speed of the unit output shaft 6, etc., and the depression of an accelerator pedal (not shown). The shift control controller 80 processes these detected values as an operating state, and according to the operating state, the solenoid 9
1 and 92, the power circulating mode clutch 9 and the direct connection mode clutch 10 are selectively engaged to switch between the power circulating mode and the direct connection mode, and to increase the unit torque ratio it according to the operating state by + torque. The gear ratio control of the continuously variable transmission 2 is performed by selectively driving the solenoid 50 or the −torque solenoid 55.

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 by the pressure regulator 100 controlled by the PL solenoid 90, and is supplied to the line pressure circuit 101 as the supply pressure PL.

The line pressure circuit 101 is connected to a + torque control valve 40 and a −torque control valve 45 constituted by pressure control valves. These control valves 40 and 45 are driven by a shift control controller 80. + Torque solenoid 50 and-
The control pressure P is applied to the oil chambers 30A and 30B of the hydraulic cylinder 30 in accordance with the output pressure Psol from the torque solenoid 55.
supply c. The + torque solenoid 50 and the -torque solenoid 55 are driven so that the output pressure Psol can be continuously changed by, for example, duty control.

The line pressure circuit 101 is provided with a solenoid 91 for controlling the power circulation mode clutch 9 and a solenoid 92 for controlling the direct connection mode clutch 10. In response to the increase in the signal pressure from the solenoid 91, the control valve 93 supplies the line pressure from the manual valve 60 to the power circulation mode clutch 9 and engages. On the other hand, when the signal pressure from the solenoid 91 decreases, the control valve 93 turns on. The power circulation mode clutch 9 is connected to the drain and released.

Similarly, in response to the increase in the signal pressure from the solenoid 92, the control valve 94 supplies the line pressure from the manual valve 60 to the direct connection mode clutch 10 and engages, while the signal pressure from the solenoid 92 decreases. Then, the control valve 94 connects the direct connection mode clutch 10 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 by the hydraulic cylinder 30 is mainly performed by a + torque control valve 40 and a -torque control valve 45 composed of a pair of pressure control valves. The output pressure Psol is used as a signal pressure, and the supply pressure PL of the line pressure circuit 101 and +
It controls the differential pressure between the output pressure (control pressure Pc) of the torque control valve 40 and the −torque control valve 45.

Here, the + torque control valve 4
Since the 0 and -torque control valves 45 have the same configuration, only the + torque control valve 40 will be described below.

This + torque control valve 40 is
+ The output pressure Psol of the torque solenoid 50 is
s is connected to the port 40a at the upper end in the figure. In addition,
The + torque solenoid 50 is of a normally closed type in which the output pressure Psol becomes 0 when no power is supplied.

+ Output pressure Psol of torque solenoid 50
Urges the spool 40s of the + torque control valve 40 downward in the figure via the port 40a, and in addition, the port 40b applies the control pressure P from the output port 40d.
c is fed back to bias the spool 40s downward.

A cylindrical movable plug 40p is in contact with the lower end of the spool 40s in the figure, and an output pressure Psol is provided at a predetermined position facing the outer periphery of the movable plug 40p.
A port 40f for guiding the supply pressure PL is formed so as to urge the spool 40s upward in opposition to the above, and in addition to feeding back the supply pressure PL, the spool 40s is attached to the port 40f side upward in the figure. The energizing spring 40r is accommodated.

When the output pressure Psol is within a predetermined value,
The supply pressure port 40c communicating with the line pressure circuit 101 is
The output port 40d is configured to communicate with the oil passage 41 via the output port 40d. When the output pressure Psol increases, the spool 40s is displaced downward in the drawing against the spring 40r, and the output port 40d communicates with the drain port 40e. Thus, the control pressure Pc is configured to be connected to the drain port 40e.

Here, the pressure receiving area of the port 40b on which the spool 40s receives the feedback of the control pressure Pc and the pressure receiving area of the movable plug 40p abutting on the spool 40s receiving the supply pressure PL are set to the same value As. The differential pressure between the supply pressure PL and the control pressure Pc is fed back so as to urge the spool 40s upward in the drawing.

Here, assuming that the pressure receiving area in which the spool 40s receives the output pressure Psol from the port 40a is Asol and the urging force of the spring 40r is Fs, the balance equation is as follows: Psol · Asol = (PL−Pc) · As + Fs ... (1) Therefore, a = Asol / As, b = Fs / As
By transforming the above equation (1) as (constant), the following equation is obtained: PL−Pc = a · Psol−b (2) As shown in FIG. 6, the supply pressure PL corresponds to the output pressure Psol. And the control pressure Pc, the differential pressure PL-Pc can be controlled.

When the output pressure Psol = 0, the differential pressure PL
Although −Pc <0, the control pressure Pc does not exceed the supply pressure PL because the source pressure of the control pressure Pc is the supply pressure PL of the line pressure circuit 101. At this time, the spool 40s is in the pressure adjustment state. Rc is pushed upward in the drawing by the spring force Fs, and Pc = PL in which the supply pressure ports 40c and 40d communicate with each other.
State.

In other words, a dead zone before the start of pressure adjustment is created by the spring force Fs, and the characteristics of the control pressure Pc are shown in FIG. 6 when the supply pressure PL is assumed to be constant with respect to the output pressure Psol. And the output pressure Psol = 0
And the section between the broken lines in the figure is the dead zone.

That is, when the output pressure Psol from the + torque solenoid 50 increases, the differential pressure PL-Pc increases,
Further, by the spring force Fs, Psol = b / a = F
Below s / Asol, Pc = PL, as described above.

The characteristic of the differential pressure PL-Pc is as follows.
Does not change because the control pressure Pc also changes. However, since the value of Pc exists only within the range of 0 ≦ Pc ≦ PL, when the supply pressure PL decreases, the differential pressure PL−
The Pc value may be limited by the value of the supply pressure PL.

That is, the + torque control valve 40 can control the differential pressure between the supply pressure PL and the control pressure Pc.
Further, when the + torque solenoid 50 of the solenoid proportional valve is not energized, the control pressure Pc is equal to the supply pressure PL.

Incidentally, the torque control valve 45
Also, each of the ports 45a to 45f, the spool 45s, and the spring 45r are formed in the same manner as the + torque control valve 40.

These two pressure control valves, the + torque control valve 40 and the −torque control valve 4
5 is applied to the oil chamber 30 of the hydraulic cylinder 30.
A and 30B are supplied to the solenoids 50 and 55, and only one of the solenoids 50 and 55 is driven at this time. Both are not energized at the same time, and the non-energized control pressure Pc is, as described above, the supply pressure Pc. PL, the control pressure Pc of the + torque control valve 40 becomes equal to the piston chamber 30.
A, the control pressure Pc of the torque control valve 45 is supplied to the piston chamber 30B.

The control of the hydraulic cylinder 30 for driving the trunnion 23 in the axial direction is performed by, for example, a + torque solenoid 50.
When only the torque solenoid 55 is driven, the torque solenoid 55 is not energized, so that the output pressure Psol from the torque solenoid 55 is 0, and the control pressure Pc of the torque control valve 45 becomes equal to the supply pressure PL.

On the other hand, + the output pressure P of the torque solenoid 50
sol controls the pressure difference between the supply pressure PL and the control pressure Pc, which results in:
Control pressure Pc from 5 and + torque control valve 4
The differential pressure of the control pressure Pc of 0 is controlled.

Accordingly, the control pressure Pc of the + torque control valve 40 is controlled by the difference ΔP = Pinc−Pd between the pressure Pinc of the oil chamber 30A and the pressure Pdec of the oil chamber 30B.
ec (> 0), that is, the pressure difference ΔP across the piston 31 is controlled.

That is, the positive transmission torque from the input shaft side of the toroidal type continuously variable transmission 2 to the sun gear 5a side of the planetary gear mechanism 5 can be controlled.

Conversely, when only the -torque control valve 45 communicating with the oil chamber 30B is driven, the pressure difference ΔP = Pdec−Pinc (> 0) of the piston 31 is similarly controlled. Can control the reverse transmission torque, that is, the negative transmission torque from the sun gear 5 a side of the planetary gear mechanism 5 to the input shaft side of the continuously variable transmission 2.

By driving one of the + torque solenoid 50 and the −torque solenoid 55 in this manner, the differential pressure ΔP applied to the piston 31 of the hydraulic cylinder 30 is controlled with high precision, and the torque of the power roller 20 is controlled. Since the transmission force control can be performed with high accuracy, and as shown in FIG. 17, near the geared neutral point in the power circulation mode, the torque transmission force of the power roller 20 can be controlled with high accuracy. Regardless of manufacturing variations of the hydraulic circuit 30 and the hydraulic circuit, etc., it is possible to set and maintain the differential pressure ΔP = 0 between the front and rear of the piston 31, as in the above-described conventional example. Can reliably prevent the generation of torque in the reverse direction and maintain the stopped state, and use a toroidal-type continuously variable transmission 2 with an infinitely variable transmission ratio. Drivability and reliability of the transmission can be greatly improved.

Further, since the + torque solenoid 50 and the −torque solenoid 55 are normally closed, if both power supply are cut off at the geared neutral point, the differential pressure ΔP across the piston 31 becomes zero, and the power roller 20 Can be set to 0,
It is possible to reduce the power consumption when the vehicle is stopped, reduce the power consumption when the engine is idling, reduce the size of the generator and the battery, and when the + torque solenoid 50 and the -torque solenoid 55 are disconnected. Then, since the control pressure Pc = supply pressure PL and the differential pressure ΔP before and after the piston 31 becomes 0, even if the disconnection occurs during traveling,
As a result, the torque transmission force of the power roller 20 becomes zero, thereby preventing a sudden shift and ensuring fail-safe.

Further, at and around the geared neutral point of the continuously variable transmission with an infinite transmission ratio, by controlling the differential pressure ΔP across the piston 31 so that the power roller 20 transmits a small torque in the forward direction, Similar to a vehicle equipped with a conventional torque converter, creep can be generated, and further, this creep force can be set to an arbitrary value by controlling the pressure difference ΔP between the front and rear of the piston 31. When controlling to the geared neutral point and its vicinity, the differential pressure ΔP becomes small, but since the oil pressure in the oil chambers 30A and 30B can be controlled near the supply pressure PL, the air mixed with the hydraulic oil has It is possible to improve the control accuracy of the torque transmission force and the tilt angle of the power roller 20 by suppressing the influence of the decrease.

Further, the hydraulic supply may be a single circuit,
Unlike the conventional example, there is no need to provide a plurality of hydraulic supply sources, so that the number of components can be reduced and the manufacturing cost can be reduced.

FIGS. 7 and 8 show a second embodiment, in which the + torque control valves 40 and-of the first embodiment are shown.
The arrangement of the torque control valve 45 is changed, and a part of the configuration is changed. Other circuit configurations are the same as those of the first embodiment.

The connection between the + torque control valve 40 and the −torque control valve 45 and the oil chambers 30A and 30B of the hydraulic cylinder 30 is opposite to that of the first embodiment.
A port 40f for connecting the output port 40d of the + torque control valve 40 to the oil chamber 30B, connecting the output port 45d of the -torque control valve 45 to the oil chamber 30A, and feeding back the supply pressure PL to the spools 40s, 45s; 45f and movable plug 40
p and 45p are abolished.

Then, the + torque control valve 40
And the drain ports 40e and 45e of the torque control valve 45 communicate with each other,
Is connected to a tank (not shown).

The pressure holding valve 47 is set such that the valve opening pressure is set substantially equal to 0, and the + torque control valve 40 and-
Oil chambers 30B, 3 via torque control valve 45
This prevents the air from being mixed into 0A, thereby ensuring the control accuracy of the hydraulic cylinder 30.

The output pressure Psol of the + torque solenoid 50 is applied to the lower end of the spool 40s of the + torque control valve 40, and the spool 40s is moved by the spring 40r.
Is urged upward in the figure opposite to
The control pressure Pc from the output port 40d is fed back to 0b 'so as to urge the spool 40s downward.

Then, the output pressure Psol of the port 40a increases, and the spool 40s applies the urging force Fs of the spring 40r.
When the pressure is displaced upward in the drawing, the supply pressure port 40c communicates with the output port 40d, the control pressure Pc applied to the oil chamber 30B of the hydraulic cylinder 30 increases, and the relationship between the output pressure Psol and the control pressure Pc becomes As shown in FIG.

At this time, the -torque solenoid 55 is not energized as described above, and the control pressure Pc from the output port 45d of the normally closed -torque control valve 45 remains zero.

Therefore, the differential pressure ΔP applied to the piston 31 of the hydraulic cylinder 30 is determined by the oil pressure of the oil chamber 30A being Pdec,
Assuming that the oil pressure of the oil chamber 30B is Pinc, ΔP = Pinc−Pdec, and the control is performed so as to balance the torque transmission force of the power roller 20.

Conversely, when the -torque solenoid 55 is operated, the + torque solenoid 50 is not energized, and the differential pressure ΔP applied to the piston 31 of the hydraulic cylinder 30 becomes ΔP = Pdec−Pinc, and the torque transmission of the power roller 20 is performed. The force is the first
It will be controlled in the opposite direction to the embodiment.

Thus, the + torque control valve 4
0 and-spool 40 of torque control valve 45
The feed pressure PL need not be fed back to the s and 45s, and the feedback ports 40f and 45f
Since the structure can be simplified by eliminating the need for a movable plug and a movable plug, the manufacturing cost can be reduced, and the structure of the spools 40s and 45s can be simplified, and miniaturization can be promoted. It is possible to further improve the control accuracy by reducing the sliding resistance.

The drain port 4 of the + torque control valve 40 and the −torque control valve 45
0e and 45e communicate with each other, and are connected to a tank (not shown) via a pressure holding valve 47 whose valve opening pressure is set extremely low downstream thereof. And-it is possible to prevent air from entering the oil chambers 30B and 30A via the torque control valve 45, and to ensure the control accuracy of the hydraulic cylinder 30.

[Brief description of the drawings]

FIG. 1 is a schematic configuration diagram of a continuously variable transmission with an infinite gear ratio showing an 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 view shown by an arrow A in FIG. 3;

FIG. 5 is a schematic circuit diagram showing a configuration of a hydraulic control device.

FIG. 6 is a diagram showing a relationship among a solenoid output pressure Psol, a control pressure Pc, and a line pressure PL.

FIG. 7 is a schematic circuit diagram illustrating a configuration of a hydraulic control device according to a second embodiment.

FIG. 8 is a diagram showing a relationship between a solenoid output pressure Psol and a control pressure Pc.

FIG. 9 is a graph showing a conventional example and showing a relationship between a gear ratio of a continuously variable transmission and a unit gear ratio.

[Explanation of symbols]

 Reference Signs List 1 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 30 hydraulic pressure Cylinder 30A, 30B Oil chamber 31 Piston 40 + Torque control valve 40b, 40f Port 40c Supply pressure port 40d Output port 40e Drain port 40r Spring 40p Movable plug 40s Spool 45-Torque control valve 47 Holding pressure valve 50 + Torque solenoid 55-Torque solenoid 80 shift control controller 101 line pressure circuit

Claims (7)

[Claims] 1. A toroidal-type continuously variable transmission that continuously changes a gear ratio by tilting a power roller held between an input / output disk and a constant transmission are connected to a unit input shaft, respectively. An infinitely variable speed ratio transmission in which the output shafts of the step transmission and the fixed transmission are connected to the unit output shaft via a planetary gear mechanism, a power circulation mode clutch and a direct connection mode clutch, and torque transmission acting on the power roller A hydraulic cylinder capable of changing a force via a piston responding to a differential pressure between the first oil chamber and the second oil chamber; and a shift control means for controlling a differential pressure of the hydraulic cylinder in accordance with an operation state of the vehicle. In the shift control device for a continuously variable transmission with an infinite gear ratio, the shift control unit communicates with the first and second oil chambers, respectively.
A shift control device for a continuously variable transmission with an infinite gear ratio, comprising: first and second pressure control valves capable of continuously reducing a supply pressure from a hydraulic pressure source. 2. The speed change control means includes first and second signal pressure supply means for supplying a signal pressure to first and second pressure control valves in accordance with a differential pressure command value, respectively. Alternatively, when the differential pressure command value to the second pressure control valve is a minimum value, the first or second pressure control valve communicates an output port that communicates with the first and second oil chambers with a drain. The shift control device for a continuously variable transmission with an infinite speed ratio according to claim 1, wherein: 3. The speed change control means includes first and second signal pressure supply means for supplying a signal pressure to first and second pressure control valves according to a differential pressure command value, respectively. Alternatively, when the differential pressure command value to the second pressure control valve is a minimum value, the first or second pressure control valve sets an output port communicating with each of the first and second oil chambers to a supply pressure side. The speed change control device for an infinitely variable speed ratio transmission according to claim 1, wherein the speed change control device is in communication. 4. The first and second signal pressure supply means,
4. The continuously variable transmission according to claim 2, wherein the differential pressure command value is set to a minimum value or a value equivalent to the minimum value when the power is turned off. Transmission control device. 5. The pressure control valve according to claim 1, wherein the first and second pressure control valves are configured to receive a supply pressure from a single hydraulic pressure source and to continuously change a differential pressure between a control pressure of the output port and the supply pressure. The speed change control device for an infinitely variable speed ratio transmission according to claim 1, wherein: 6. The first and second pressure control valves have a drain side communicated with each other, and a pressure-holding valve having a valve opening pressure set to a predetermined extremely low pressure is interposed downstream thereof. The shift control device for a continuously variable transmission with an infinite gear ratio according to claim 1. 7. The first and second signal pressure supply means,
4. The continuously variable transmission according to claim 2, wherein one of the variable transmissions controls the signal pressure in accordance with the differential pressure command value, and the other controls the signal pressure to a minimum value. Transmission control device.