JPH11247984A - Transmission control device for indefinite gear ratio type continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To make control of a tilt angle of a power roller consistent with control of a torque transmission value quickly and with a high degree of accuracy, irrespective of unevenness of a hydraulic control device. SOLUTION: This transmission control means for controlling a hydraulic cylinder which drives a power roller in a troidal type continuously variable transmission through a trunnion so as to obtain a target gear ratio in accordance with an operating condition of a vehicle, is composed of a first piston control means for controlling pressures in a first and a second oil chamber 30A, 30B defined in the hydraulic cylinder by a piston, and a shift control valve 46 for controlling the hydraulic pressure fed into the hydraulic cylinder, a second piston control means for feeding back a tilt angle of the power roller to the shift control valve 46, and a servo change-over valve 48 for selectively changing over the first and second piston control means.

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 Application Laid-Open No. 9-42428.

[0003] In this system, a toroidal type continuously variable transmission capable of continuously changing the gear ratio and a constant transmission (reduction gear) are arranged in parallel on a unit input shaft of a continuously variable transmission with an infinitely variable gear ratio connected to an engine. The output shaft of the continuously variable transmission is connected to the sun gear of the planetary gear mechanism, and the output shaft of the fixed transmission is connected to the planetary gear via a power circulation mode clutch. It is connected to the carrier of the gear mechanism.

[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. 17, while the power circulation mode clutch is connected and the direct connection mode clutch is disconnected, the unit speed ratio (IVT ratio in the figure) is changed according to the speed ratio difference between the continuously variable transmission and the fixed transmission. The power input mode / unit output axis speed is controlled from negative value to positive value including infinity (= geared neutral), and the power-circulation mode and the power-circulation mode in which gearshift control is continuously performed. On the other hand, it is possible to selectively use 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.

[0006]

However, in the conventional shift control device for a continuously variable transmission with an infinite transmission ratio, the toroidal type continuously variable transmission controls the tilt angle of the power roller and the torque transmission force. However, since the trunnion supporting the power roller is driven by the hydraulic control device that drives the trunnion in accordance with the pressure difference between the front and rear of the hydraulic cylinder, the hydraulic responsiveness of the hydraulic cylinder and the like is insufficient due to variations in the dimensional tolerances of the hydraulic control device. And drive characteristics, making it difficult to quickly and accurately control the tilt angle toward the target gear ratio, especially in the geared neutral or power-circulation mode where the target gear ratio is infinite with the unit gear ratio. In order to maintain the vicinity, it is necessary to adjust the differential pressure across the hydraulic cylinder with high precision and quickly change the torque transmission force. If the accuracy of the pressure difference between the front and rear is reduced due to the variation in the characteristics of the pressure control device, the gear is shifted slightly from the geared neutral in the forward or backward direction, and a torque is generated in a direction opposite to the traveling direction intended by the driver. However, there is a problem that the driver feels strange.

In addition, it is difficult to accurately set the differential pressure across the piston of a hydraulic cylinder that drives a toroidal-type continuously variable transmission due to the lack of control response of the hydraulic pressure and the variation in actuator drive characteristics. Therefore, when the input torque of the continuously variable transmission changes suddenly, the actual differential pressure varies with the differential pressure command value, so that the power roller of the continuously variable transmission exceeds a predetermined tilt angle range, There is a possibility that the trunnion supporting the disk may hit the tilt stopper to generate a shock, or the disk and the power roller may slip.

SUMMARY OF THE INVENTION The present invention has been made in view of the above problems, and makes it possible to quickly and accurately control the tilt angle of the power roller and the torque transmission amount regardless of the variation of the hydraulic control device. The purpose is to:

[0009]

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 with an infinite transmission ratio coupled to a unit output shaft via a power circulation mode clutch and a direct connection mode clutch; a hydraulic cylinder for driving a power roller of the toroidal type continuously variable transmission via a trunnion; A transmission control unit for controlling the hydraulic cylinder so as to attain a target transmission ratio according to an operating state, wherein the transmission control unit comprises a piston of the hydraulic cylinder. First piston control means for controlling the pressures of the first and second oil chambers defined by:
A shift control valve for controlling the oil pressure to the hydraulic cylinder, a second piston control means for feeding back the tilt angle or the actual gear ratio of the power roller to the shift control valve, and the first and second piston control means. Control switching means for selectively switching.

[0010] In a second aspect based on the first aspect, the control switching means is configured to control the first switching in accordance with an operating state.
And the second piston control means is selectively switched.

In a third aspect based on the first aspect, the shift control means detects at least a vehicle speed as an operating state, and the control switching means includes a first piston control means at a predetermined vehicle speed or less. On the other hand, when the vehicle speed exceeds a predetermined vehicle speed, the second piston control means is selected.

In a fourth aspect based on the first aspect, the shift control means includes means for detecting or estimating a tilt angle of the power roller, and the control switching means includes a first piston control means. Is selected, when the tilt angle exceeds or is about to exceed the predetermined range, switching to the second piston control means is performed.

In a fifth aspect based on the first aspect, the control switching means selects the second piston control means when the target speed ratio exceeds a predetermined value, while the target speed ratio is set to a predetermined value. In the following cases, the first piston control means is selected.

In a sixth aspect based on the third aspect, the shift control means includes means for detecting or setting a target driving force or a load, and the control switching means includes a means for detecting a target driving force or a load. The predetermined vehicle speed is changed to a higher speed according to the increase.

In a seventh aspect based on the first aspect, the shift control means selectively engages the power circulation mode clutch and the direct connection mode clutch to selectively switch between the power circulation mode and the direct connection mode. Power switching mode switching means for switching between the first piston control means in the power circulation mode and the second piston control means in the direct connection mode.

In an eighth aspect based on the seventh aspect, the shift control means synchronously switches between the power circulation mode and the direct connection mode and switches between the first and second piston control means. A switching valve is provided.

In a ninth aspect based on the first aspect, the second piston control means always sets the target at any time even when the control switching means selects the first piston control means. Control is continued so as to follow the speed ratio or the actual speed ratio.

[0018] In a tenth aspect based on the first aspect, the first piston control means always sets the target shift speed even when the control switching means selects the second piston control means. Control is continued to follow the ratio.

In an eleventh aspect based on the fourth aspect, the speed change control means includes means for detecting a tilt angular velocity of the power roller, and the control switching means is configured to control the first piston control means from the first piston control means. The range of the tilt angle for switching to the two-piston control means is reduced in accordance with the absolute value of the tilt angular velocity.

In a twelfth aspect based on the fourth aspect, the speed change control means includes means for detecting a tilt angular velocity of the power roller, and the control switching means is configured to detect the tilt angle of the power roller from the first piston control means. The upper limit of the range of the tilt angle to be switched to the two-piston control means is reduced in accordance with the increase of the tilt angular velocity in the positive direction, while the lower limit of the range of the tilt angle is reduced in the negative direction of the tilt angular velocity. Increases with decreasing.

[0021]

According to the first aspect of the present invention, the hydraulic cylinder control means is configured to be capable of switching between the first means and the second means, so that the hydraulic cylinder control means can operate at the neutral point (geared neutral point) in the power circulation mode or in the vicinity thereof. Since the control by the first piston control means becomes possible, the neutral point can be maintained by the torque transmission force control, and the second piston control means can be provided in an operating condition requiring accurate speed ratio control. , It is possible to quickly and accurately control the tilt angle = actual gear ratio toward the target gear ratio, regardless of manufacturing errors or tolerances of the hydraulic control device. And the control of the amount of torque transmission can be achieved both quickly and accurately.

Further, in the second invention, since the first and second piston control means can be automatically switched according to the operation state, the operation characteristics of the first and second piston control means are determined. By selecting one suitable for the state, the operability of the continuously variable transmission with an infinite speed ratio can be improved.

Further, in the third invention, when the vehicle speed exceeds a predetermined vehicle speed, the actual speed ratio is controlled quickly and with high accuracy toward the target speed ratio by switching to the second piston control means. By switching to the first piston control means, the transmission ratio at or near the neutral point in the power circulation mode is reliably performed by accurate torque transmission force control, thereby controlling the tilt angle of the power roller and torque transmission. Control of the amount can be easily realized.

According to the fourth aspect of the present invention, when the first piston control means is selected, when the tilt angle of the power roller is going to be out of the predetermined range or beyond the predetermined range,
Switching to the second piston control means makes it possible to prevent the tilt angle of the power roller from hitting the tilt stopper beyond a predetermined range.

In the fifth aspect, the switching between the first and second piston control means is performed at a predetermined target gear ratio, so that the neutral point of the power circulation mode in which the target gear ratio is equal to or less than a predetermined value. And at a speed ratio in the vicinity thereof, the torque transmission force control by the first piston control means becomes possible, and in other target gear ratio ranges, the second piston control means enables quick and accurate speed ratio control.

In the sixth invention, when switching between the first and second piston control means is performed in accordance with the target driving force or the load, the higher the load or the target driving force, the higher the vehicle speed. By switching from the first piston control means to the second piston control means, the range of the torque transmission force control can be expanded, and the accuracy of the torque transmission force control in the vicinity of the neutral point of the power circulation mode can be further improved.

According to the seventh aspect of the present invention, the switching of the power transmission mode and the switching of the piston control means can be performed simultaneously, so that the control contents can be simplified.

Further, in the eighth invention, since the mode switching valve for simultaneously switching the power transmission mode and switching the piston control means is provided, the control contents are simplified, the number of parts is reduced, and the manufacturing cost is reduced. Reduction can be achieved.

Further, in the ninth aspect, even when the control is performed by the first piston control means, the second piston control means is made to follow the target speed ratio or the actual speed ratio, so that when the control is switched, Switching shock can be reduced.

Further, in the tenth aspect, even when the control is performed by the second piston control means, the first piston control means is made to follow the hydraulic pressure to achieve the target gear ratio, so that the control switching is performed. Therefore, the switching shock can be reduced.

According to an eleventh aspect of the present invention, the tilt angle range of the power roller is detected, and the range of the tilt angle for switching from the first piston control means to the second piston control means is determined by the magnitude of the absolute value of the tilt angle speed. By reducing accordingly, even if the tilting speed increases due to a sudden change in the input torque or the like, it is possible to reliably prevent collision with the tilting stopper.

According to a twelfth aspect of the present invention, one of the upper limit value and the lower limit value of the range of the tilt angle at which the first piston control means is switched to the second piston control means is changed according to the positive or negative of the tilt angular velocity. Therefore, even if the tilting speed increases due to a sudden change in the input torque or the like, the collision with the tilting stopper can be reliably prevented without unnecessarily reducing the tilting angle range.

[0033]

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 toroidal type continuously variable transmission 2 is used to form an infinitely variable speed ratio continuously variable transmission.

FIG. 1 schematically shows a continuously variable transmission with an infinitely variable speed ratio, which is constructed in the same manner as in the conventional example.
In addition, a toroidal-type continuously variable transmission 2 capable of continuously changing the gear ratio, and a constant transmission 3 composed of gears 3a and 3b
(Reduction gears) 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 a sun gear 5a of the planetary gear mechanism 5,
The output shaft 3c of the constant transmission 3 is connected to a power circulation mode clutch
Through the carrier 5b of the planetary gear mechanism 5.

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 the 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 of a double-cavity toroidal type in which a pair of input disk 21 and output disk 22 sandwich and press a power roller 20, respectively. 3 and 4, the lower end is connected to the hydraulic cylinder 30 so that the trunnion 2 is axially displaceable 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 in 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 = Pinc−Pdec of the piston 31 is the torque transmitting force of the toroidal type continuously variable transmission 2.

Here, since the torque transmitting 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 transmitting 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 ΔP is reduced so as to be in proportion to the transmission torque, the power roller returns to the neutral position, and the 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, selectively engages the power circulation mode clutch 9 and the direct connection mode clutch 10 according to the operating state, and outputs the power circulation mode. And the direct connection mode, and the speed ratio control of the continuously variable transmission 2 is performed so as to obtain the unit speed ratio it according to the driving state, and the torque transmission force is controlled near the vehicle speed VSP = 0 in the power circulation mode. .

For this reason, the hydraulic control device for performing the shift control includes first piston control means capable of controlling the differential pressure ΔP of the piston 31 with high precision, and the tilt angle of the power roller 20 with high precision. A possible second piston control means and a control mode switching means for selectively switching between the first piston control means and the second piston control means are provided.
, The shift control controller 80 sends a command value corresponding to the target driving force to the + torque solenoid 50 and the −torque solenoid 55 as the first piston control means, and the step motor 36 as the second piston control means. A command value corresponding to the target shift is transmitted to a servo switching solenoid valve 48 as control mode switching means.
To switch the first or second piston control means according to the operation state.

Next, the first piston control means, the second piston control means, and the control mode switching means will be described in detail with reference to the schematic circuit diagram of the hydraulic control device shown in 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 supplied to the line pressure circuit 101 as the supply pressure PL.

The line pressure circuit 101 has a + torque control valve 4 as first piston control means.
0 and −torque control valves 45 are connected, and these control valves 40 and 45 are connected to a + torque solenoid 50 driven by a shift control controller 80.
And a control pressure Pc is supplied to the oil passages 41 and 42 in accordance with the output pressure Psol from the torque solenoid 55.

Further, a shift control valve 46 as second piston control means is connected to the line pressure circuit 101. The shift control valve 46 has a target value of the step motor 36 driven by the shift control controller 80, Precess cam 35 and L link 38
Output ports 46b, 46b in accordance with the actual speed ratio fed back from the mechanical feedback means comprising
Control the flow to c.

Upstream of the hydraulic cylinder 30, the servo switching valve 4 responsive to the signal pressure of the servo switching solenoid 56 driven by the transmission control controller 80.
8 is provided to connect the oil chambers 30A and 30B to one of the oil passages 41 and 42 of the first piston control means and the output ports 46b and 46c of the second piston control means.

Here, the first piston control means includes a + torque control valve 40 comprising a pair of pressure control valves.
And a torque control valve 45 as a main component. The output pressure Psol from the + torque solenoid 50 and the −torque solenoid 55 is used as a signal pressure, and the supply pressure PL of the line pressure circuit 101 with respect to the signal pressure. The differential pressure of the output pressure (control pressure Pc) is controlled.

Here, the + torque control valve 4
Since the 0 side and the −torque control valve 45 side are configured in the same manner, hereinafter, the + torque control valve 40
The side will be described.

This + torque control valve 40 is
The output pressure Psol of the + torque solenoid 50 is connected to one end. The + torque solenoid 50 is of a normally closed type in which the output pressure Psol becomes 0 when no power is supplied.

The output pressure Psol of the + torque solenoid 50 is applied to the + torque control valve 40 via the port 40a.
Urges the spool 40s downward in the figure.
The control pressure Pc is fed back to the port 40b so as to urge the spool 40s downward.

A port 40f is formed so as to urge the spool 40s upward against the output pressure Psol, so that the supply pressure PL is fed back and a spool 40s is provided on the port 40f side in the figure. A spring 40r biasing upward is provided.

When the output pressure Psol is within a predetermined value,
The supply pressure port 40c communicating with the line pressure circuit 101 is
The spool 40s is configured to communicate with the oil passage 41 via the output port 40d, and when the output pressure Psol further increases and the spool 40s is displaced downward in the drawing against the spring 40r, the output port 40d is connected to the drain port. 4
The control pressure Pc is configured to be connected to the drain port 40e in communication with 0e.

Here, the pressure receiving area at the port 40b side where the spool 40s receives the feedback of the control pressure Pc and the pressure receiving area at the lower portion of the spool 40s receiving the supply pressure PL are set to the same value As. The differential pressure between PL and the control pressure Pc urges the spool 40s upward in the drawing.

Here, assuming that the pressure receiving area in which the spool 40s receives the output pressure Psol 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), PL-Pc = a.Psol-b (2) is obtained, and the differential pressure between the supply pressure PL and the control pressure Pc corresponding to the output pressure Psol. PL-Pc can be controlled.

When the output pressure Psol = 0, the differential pressure PL
−Pc <0, but 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.
The pressure of the spool 40s is not adjusted, and the spring force Fs
, And the port 40c and 40d communicate with each other.
The state becomes PL.

Therefore, a dead zone is created before the start of pressure adjustment due to the spring force Fs. The characteristic of the control pressure Pc is shown in FIG. 7 when the supply pressure PL is assumed to be constant with respect to the output pressure Psol. Become like

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.

Note that the torque control valve 45
Is configured similarly to the + torque control valve 40 described above.

These two pressure control valves, the + torque control valve 40 and the −torque control valve 4
5 is supplied from the oil passages 41 and 42 to the servo switching valve 48, and when the transmission control controller 80 selects the power circulation mode, as described later, the signal pressure from the servo switching solenoid 56 is output. Is reduced, the spool of the servo switching valve 48 is displaced upward as shown in FIG. 5, and the control pressure Pc of the + torque control valve 40 is applied to the piston chamber 30A via the oil passages 41 and 42, while the -torque control valve is The control pressure Pc of 45 is applied to the piston chamber 30.
B respectively.

The hydraulic cylinder 3 for driving the trunnion 23
While 0 is being performed by the first piston control means, the + torque solenoid 50 or the -torque solenoid 55 is driven by a command from the transmission control controller 80, and only one of the solenoids is driven at this time. There is no energization at the same time, and the non-energized control pressure Pc is, as described above, the supply pressure Pc.
L.

For example, when only the + torque solenoid 50 is driven, the -torque solenoid 55 is not energized, the output pressure Psol from the -torque solenoid 55 is 0, and the -control pressure P
c is 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:
The differential pressure between the control pressure Pc of No. 5 and the control pressure Pc of the + torque control valve 40 is controlled.

Accordingly, controlling the control pressure Pc of the + torque control valve 40 is equivalent to the differential pressure ΔP = Pinc− between the pressure Pinc of the oil chamber 30A and the pressure Pdec of the oil chamber 30B.
Pdec (> 0), that is, the differential pressure Δ across the piston 31
P will be controlled.

That is, 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 differential pressure 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.

In this way, the first piston control means controls the differential pressure ΔP across the hydraulic cylinder 30 with high accuracy by driving either of the solenoids.
A torque transmission force of 0 can be performed with high accuracy.

Next, the step motor 36 and the I-link 3
7, the second piston control means including the shift control valve 46 and the mechanical feedback means will be described.

The selection of the second piston control means is based on the fact that the signal pressure from the servo switching solenoid 56 is increased and the servo switching valve 48 is stroked downward in the drawing, so that the output ports 46b and 46 of the shift control valve 46 are provided.
c communicates with the oil chambers 30B and 30A, respectively.
The oil passages 41 and 42 of the piston control means are shut off.

The shift control valve 46 for controlling the oil pressure supplied to the oil chambers 30B and 30A has a spool 46s as shown in FIGS. 5 and 6, and the spool 46s is provided in the middle of the swingable I-link 37. Be linked.

One end of the I-link 37 is connected to a step motor 36. The step motor 36 drives the I-link 37 through a feed screw mechanism (not shown) in accordance with a command from the transmission control controller 80. Then, the spool 46s is driven.

On the other hand, the other end of the I-link 37 is connected to an L-link 38 which is in sliding contact with a precess cam 35 formed on a rod protruding from the lower end of the trunnion 23, thereby displacing the trunnion 23 around its axis, that is, tilting. Angle = mechanical feedback means for feeding back the actual gear ratio and the axial displacement (gear speed) is connected.

The tilt angle of the power roller 20 is transmitted from the L link 38 to the I link 37 via the trunnion 23 and the precess cam 35, and the actual gear ratio is fed back to the spool 46s.

A supply pressure port 46a communicating with the line pressure circuit 101 and two output ports 46b and 46c are formed in the valve body of the shift control valve 46, and the pressure oil from the supply pressure port 46a outputs one of the output oils. While the pressure oil is selectively supplied to the pressure port, the pressure oil of the other output pressure port is connected to a tank (not shown) via a pressure holding valve 47 having a very low valve opening pressure.

Now, when the target speed ratio of the toroidal type continuously variable transmission 2 is set to the Lo side, in FIG. 5, the rod of the step motor 36 rotates in a direction to contract to a predetermined position.
At this time, the spool 4 of the shift control valve 46
6s is displaced upward in the figure, the supply pressure port 46a communicates with the output pressure port 46c, the oil pressure in the oil chamber 30A connected to the output pressure port 46c rises, and the power is changed according to the displacement of the piston 31. A tilting force is generated to the extent that the roller 20 is offset from the neutral point, and the toroidal-type continuously variable transmission 2
Is shifted to the Lo side.

When the speed ratio becomes Lo corresponding to the target speed ratio, the position of the L link 38 changes by the amount that the actual speed ratio is fed back by the precess cam 35.
One end of the link 37 is displaced, and the spool 46s of the shift control valve 46 is driven to return to the neutral point again. When the command value of the step motor 36 and the actual gear ratio match, the spool 46s returns to the neutral position and shifts. Ends.

When a shift command is issued to the Hi side, the operation is reversed.

In this way, the speed ratio of the toroidal type continuously variable transmission 2 can be controlled with high accuracy in accordance with the command value of the step motor 36 and the actual speed ratio of the mechanical feedback mechanism. The unit speed ratio of the infinitely variable speed ratio transmission can be controlled in accordance with the change of the speed ratio of No. 2.

Here, the switching of the shift control performed by the shift control controller 80 will be described in detail with reference to the flowchart of FIG.

First, in step S1, it is detected which control means is the current piston control. If the first piston control means drives the hydraulic cylinder 30, 1 is substituted for the variable S. On the other hand, when the hydraulic cylinder 30 is driven by the second piston control means, 2 is substituted for the variable S.

Next, at step S2, the vehicle speed sensor 83
After reading the vehicle speed VSP from step S3, the flow proceeds to step S3 to determine which piston control means is performing control. If the first piston control means is S = 1, the flow proceeds to step S4. If it is the second piston control means, the process proceeds to step S5.

In step S4, if the detected vehicle speed VSP exceeds the predetermined value Vu, the flow advances to step S7 to set S = 2 and switch to the second piston control means. The process proceeds to step S6 to hold the current first piston control means.

If the detected vehicle speed VSP is less than the predetermined value Vd in step S5 when the second piston control means is performing control, the process proceeds to step S6 and S = 1.
And switch to the first piston control means.

The predetermined values Vu and Vd are set to Vu> Vd to prevent hunting from occurring at the time of switching of the piston control means, and the predetermined vehicle speed Vd is set, for example, to around the vehicle speed VSP = 0. Is set to

Therefore, when the vehicle is stopped or at a very low speed where the vehicle speed VSP is less than the predetermined value Vd, as shown in FIG. 17, since the power circulation mode is set, the first piston control must be performed near the geared neutral point. By selecting the means, the torque transmitting force of the power roller 20 can be controlled with high accuracy. Particularly, at the geared neutral point, regardless of the manufacturing variation of the hydraulic cylinder 30 and the hydraulic control device, the front and rear of the piston 31 can be controlled. Differential pressure ΔP
Can be accurately controlled to reliably prevent torque from being generated in the direction opposite to the traveling direction intended by the driver, as in the above-described conventional example.
Thus, the operability of the continuously variable transmission having the infinite speed ratio using the above-described method can be greatly improved.

Then, the vehicle speed VSP increases to a predetermined value Vu.
Is exceeded, the control of the hydraulic cylinder 30 is switched to the second piston control means having the mechanical feedback means, so that the tilt angle of the power roller 20 can be controlled quickly and accurately with respect to the target gear ratio. Thus, accurate speed ratio control and torque transmission force control can be achieved at the same time in all shift ranges from near the stopped state of the vehicle speed VSP = 0 to the high speed range.

At the geared neutral point of the continuously variable transmission with infinite transmission ratio and at the vicinity thereof, the first piston control means capable of performing the differential pressure control with high accuracy is selected, so that the vehicle is stopped when the vehicle is stopped. By controlling the differential pressure ΔP across the piston 31 so as not to transmit the torque of the power roller 20, the vehicle can be stopped or the power roller 20 transmits a small torque in the forward direction. By controlling the differential pressure ΔP of the piston 31, creep can be generated in the same manner as in a vehicle equipped with a conventional torque converter, and the creep force can be set to an arbitrary value by controlling the differential pressure. It can be set.

Further, when the vehicle speed exceeds the predetermined vehicle speed Vu,
The second piston control means enables accurate gear ratio control, ensures good gear shift response and stability, and provides the optimum fuel economy point of the power train including the engine (optimum rotation at the required horsepower). Accurate driving force control pursuing the number and torque points) becomes possible.

Further, even when the output ports 46b and 46c are shut off by the servo switching valve 48, the second piston control means always continues the control toward the target gear ratio, thereby providing the first piston control means. At the time of switching from the piston control means to the second piston control means, it is possible to prevent a sudden change in the gear ratio and suppress a switching shock.

Also, during the selection of the first piston control means,
Instead of the target gear ratio, the current actual gear ratio is detected from the input / output shaft rotation speed of the toroidal-type continuously variable transmission 2, and the stepping motor 3 of the second piston control means is set to this actual gear ratio.
6 may be continued to be issued. In this case, when switching from the first piston control means to the second piston control means, the gear ratio does not change, and the switching shock can be prevented. Thus, drivability can be further improved.

FIG. 9 shows a second embodiment in which the first and second piston control means of the first embodiment are switched in accordance with a target gear ratio. This is the same as in the first embodiment.

First, in step S10, it is detected which control means is the current piston control. If the first piston control means drives the hydraulic cylinder 30, 1 is substituted for the variable S. On the other hand, when the hydraulic cylinder 30 is driven by the second piston control means, the variable S
Is substituted for 2.

Next, in step S11, the depression amount of an accelerator pedal (not shown), the unit input shaft rotation speed N
From the map shown in FIG. 10, a target value tNt of the unit input shaft speed is obtained from the target driving force F obtained according to t (= engine speed Ne) and the vehicle speed VSP, and this target input shaft speed tNt is set to the vehicle speed. The target unit speed ratio it is obtained by dividing the VSP by a unit output shaft speed obtained by multiplying the predetermined constant by VSP.
To control the continuously variable transmission 2. Note that the target driving force F is, for example, as disclosed in Japanese Patent Application No. Hei.
739 or the like.

In step S12, it is determined which piston control means is currently performing control.
If it is the first piston control means, the process proceeds to step S13, whereas if it is the second piston control means of S = 2, the process proceeds to step S14.

In step S13, if the target unit speed ratio it exceeds the predetermined value iu, step S13 is executed.
Proceeding to S16, S = 2 is set, and the control is switched to the second piston control means.
Proceed to and hold the current first piston control means.

If the target unit speed ratio it is smaller than the predetermined value id in step S14 when the second piston control means is performing control, the process proceeds to step S15 to set S = 1. While switching to the first piston control means, if not, the process proceeds to step S16 to maintain the current second piston control means.

The predetermined values iu and id are set to iu> id to prevent hunting from occurring when the piston control means is switched, and the predetermined gear ratio id is
For example, the vehicle speed is set near VSP = 0.

Therefore, in a vehicle stopped state or an extremely low speed when the target unit speed ratio it is less than the predetermined value id,
As shown in FIG. 17, since the power circulation mode is set, the first piston control means can always be selected near the geared neutral point to control the torque transmission force of the power roller 20 with high accuracy. In particular, since the piston control means is switched by comparing the target unit speed ratio it with the predetermined value, the gear ratio neutral point in the power circulation mode and the speed ratio control up to the vicinity thereof are controlled by the second.
This can be performed with high accuracy by the piston control means.

FIG. 11 and FIG. 12 show a third embodiment.
The switching of the piston control means according to the vehicle speed VSP of the first embodiment takes into account the load, and the other configuration is the same as that of the first embodiment.

Steps S20 and S21 in FIG.
The detection of the target driving force F in step S22 is the same as steps S1 and S2 in step S22 of the second embodiment.
Perform in the same manner as in 11.

At step S23, the vehicle speed VSP
From the map shown in FIG. 12, it is determined from the map shown in FIG.

In FIG. 12, a region I is a region where the first piston control means performs the speed change by controlling the pressure difference of the hydraulic cylinder 30, and a region III is the second piston control means which performs the speed change by controlling the tilt angle of the power roller 20. The region II between the line u and the line d in the figure is a transition region having hysteresis, and switching from the region I to the region III or the reverse direction is performed according to the current piston control means.

Therefore, in step S23, the vehicle speed V
Based on SP and the target driving force F, if the operation region is in the region I, the process proceeds to step S25, and the first piston control means is selected. If the operation region is in the region III, the process proceeds to step S26, and the second piston control device is selected. I do. If it is in the area II, the process proceeds to step S24. If the first piston control means is currently selected, the process proceeds to step S25 to maintain the current control state. If the control means has been selected, the process proceeds to step S26 to maintain the current control state.

Therefore, even if the vehicle speed VSP increases from 0 to the low speed side, the gear ratio is set to be low (geared neutral side) in the low speed range in the case of the high load, so that the high load and the low speed In the area, controlling the torque transmitting force by the first piston control means can improve the control accuracy. When the vehicle speed VSP increases beyond the u line, the processing moves to the area III and the second piston control means , The gear ratio control can be performed with high accuracy by controlling the tilt angle of the power roller 20, and the accuracy of gear shift control can be ensured regardless of the load.

FIG. 14 shows a fourth embodiment in which the servo switching valve 48 of the first embodiment is replaced by a mode switching valve 70 which also switches between the power circulation mode clutch 9 and the direct connection mode clutch 10. The other configuration is the same as that of the first embodiment.

The mode switching valve 70 opens the oil chambers 30A and 30B according to the signal pressure of the servo switching solenoid 56.
At the same time as connecting to one of the oil passages 41, 42 of the first piston control means and one of the output ports 46b, 46c of the second piston control means, pressurized oil is supplied to one of the power circulation mode clutch 9 and the direct connection mode clutch 10. When the hydraulic cylinder 30 is controlled by the first piston control means, the power circulation mode clutch 9 is engaged, and the hydraulic cylinder 30 is engaged. When 30 is controlled by the second piston control means, the direct connection mode clutch 10 is engaged.

That is, since the geared neutral point exists only in the power circulation mode, the differential pressure control by the first piston control means in the power circulation mode enables the shift control in a very low speed range of a very low speed or a high load from a stopped state. Is performed with high accuracy, and in the region in which the second piston control means performs the shift control, the gear ratio control is performed by the tilt angle control in the direct connection mode, similarly to the normal toroidal type continuously variable transmission.

Therefore, by switching the power transmission mode and switching the piston control means at the same time,
Control contents can be simplified, and in the direct connection mode, shift control of the conventional toroidal-type continuously variable transmission can be performed, so that tuning and the like can be easily performed.

Further, since an actuator or the like for switching the power transmission mode is not required, the manufacturing cost can be reduced by reducing the number of parts.

FIG. 14 shows a fifth embodiment, in which the tilt angle φ of the toroidal type continuously variable transmission is detected or estimated, and when the tilt angle exceeds a predetermined tilt angle range, the second piston control means is activated. By switching and feeding back the actual gear ratio, the trunnion 23 is prevented from colliding with the tilt stopper. Note that the tilt stopper limits the tilt angle so that the trunnion 23 does not exceed the tilt range that is mechanically limited.
The structure is the same as that of No.-7949.

First, at step S31, it is detected which control means is the current piston control. If the first piston control means drives the hydraulic cylinder 30, 1 is substituted for the variable S. On the other hand, when the hydraulic cylinder 30 is driven by the second piston control means, the variable S
Is substituted for 2.

Next, in step S2, it is determined whether or not the current piston control means is the second piston control means. If so, the flow advances to step S35 to maintain the current control state. If the first piston control means is selected, the process proceeds to step S32 to detect or estimate the tilt angle φ of the power roller 20.

The detection of the tilt angle φ is not shown,
The displacement angle may be determined by a displacement sensor such as a rotary encoder provided at the upper end of the trunnion 23, and the tilt angle φ is estimated by the ratio between the unit input shaft rotation speed Nt and the continuously variable transmission output shaft rotation speed No, that is, , Can be estimated from the actual speed ratio of the continuously variable transmission 2.

Then, in a step S33, it is determined whether or not the tilt angle φ is within a predetermined range φd <φ <φu. If the tilt angle φ is within the predetermined range, the process proceeds to step S34 to maintain the first piston control means, while if the tilt angle φ exceeds the predetermined range, the process proceeds to step S35, The control is switched to the second piston control means.

Therefore, if the input torque suddenly changes when the hydraulic cylinder 30 is controlled by the first piston control means, for example, when the vehicle jumps, the hydraulic response of the first piston control means is delayed. For example, the piston chamber 30 controlled according to the input torque
If the balance between the hydraulic pressures of A and 30B is lost and the recovery takes time, the power roller 20 applies the differential pressure ΔP
As a result, tilting may be continued, and there is a possibility of jumping over a predetermined tilting angle range.

Therefore, when the control is performed by the first piston control means and the control exceeds the predetermined range of the tilt angle, the control is switched to the second piston control means. Since the angle φ can be fed back, the angle does not exceed the predetermined tilt angle range, and the trunnion 23 can be prevented from colliding with the tilt stopper.

FIGS. 15 and 16 show a sixth embodiment.
In addition to the switching of the piston control means according to the tilt angle φ of the fifth embodiment, the upper limit value φu and the lower limit value φd for regulating the tilt angle φ are variably controlled by the tilt speed dφ / dt. The rest is the same as the fifth embodiment.

In step S32 ', the tilt speed dφ / dt is obtained from the detected or estimated tilt angle φ, and in the next step S40, the upper limit value φu is calculated from the tilt speed dφ / dt.
And the lower limit φd.

Based on the map shown in FIG. 16, the upper limit value φu is set smaller as the tilt speed dφ / dt increases toward the positive side (right side in the figure), and the lower limit value φd is set as the tilt speed dφ / d
As t decreases to the negative side (left side in the figure), it is set to be larger,
As the absolute value of the tilting speed dφ / dt increases, the usable tilting range is reduced, and even when the power roller 20 suddenly tilts, the response delay of the second piston control unit is taken into account. Therefore, it is possible to reliably prevent the trunnion 23 from colliding with the tilt stopper.

[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 conceptual diagram showing a relationship among an I-link, an L-link, a shift control valve, and a step motor.

FIG. 7 shows the solenoid output pressure, control pressure Pc and line pressure P
The figure which shows the relationship of L.

FIG. 8 is a flowchart illustrating an example of control performed by a shift control controller, and illustrates a main routine.

FIG. 9 is a flowchart illustrating another example of the control performed by the shift control controller according to the second embodiment.

FIG. 10 is a map showing a relationship between a vehicle speed VSP using the engine speed Ne as a parameter and a target driving force F.

FIG. 11 is a flowchart illustrating another example of the control performed by the shift control controller according to the third embodiment.

FIG. 12 is also a map of an area determined according to a vehicle speed VSP and a target driving force F.

FIG. 13 is a schematic circuit diagram illustrating a configuration of a hydraulic control device according to a fourth embodiment.

FIG. 14 is a flowchart illustrating another example of the control performed by the shift control controller according to the fifth embodiment.

FIG. 15 is a flowchart illustrating another example of control performed by the shift control controller according to the sixth embodiment.

FIG. 16 is a map of the upper limit and the lower limit of the tilt angle according to the tilt angle and the tilt angular velocity.

FIG. 17 is a graph showing a conventional example and showing a relationship between a speed ratio of a continuously variable transmission and a unit speed 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 Cylinders 30A, 30B Oil chamber 31 Piston 35 Precess cam 36 Step motor 37 I-link 38 L-link 40 + Torque control valve 41, 42 Oil passage 45-Torque control valve 46 Shift control valve 47 Holding pressure valve 48 Servo switching valve 50 + Torque solenoid 55 -Torque solenoid 56 Servo switching solenoid 70 Mode switching valve 80 Shift control controller 81 Input shaft speed sensor 82 Output shaft speed sensor 83 Vehicle speed sensor 101 Line pressure circuit

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁶ Identification code FI F16H 59:70 63:06

Claims (12)

[Claims] 1. A toroidal-type continuously variable transmission that continuously changes the 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. An infinitely variable speed ratio transmission in which the output shafts of a step 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 connection mode clutch, and the toroidal type transmission through a trunnion. Speed change of an infinitely variable speed continuously variable transmission, comprising: a hydraulic cylinder that drives a power roller of a step transmission; and a speed change control unit that controls the hydraulic cylinder so as to have a target speed ratio according to an operation state of a vehicle. In the control device, the shift control means includes: first piston control means for controlling pressures of first and second oil chambers defined by pistons of the hydraulic cylinder A shift control valve for controlling a hydraulic pressure applied to the hydraulic cylinder; second piston control means for feeding back a tilt angle or an actual gear ratio of a power roller to a shift control valve; and the first and second piston control means. And a control switching means for selectively switching the speed change ratio of the continuously variable transmission. 2. The continuously variable transmission according to claim 1, wherein the control switching means selectively switches the first and second piston control means in accordance with an operation state. Transmission control device. 3. The speed change control means detects at least a vehicle speed as an operating state, and the control switching means selects the first piston control means when the vehicle speed is equal to or lower than a predetermined vehicle speed, and selects the second piston control means when the vehicle speed exceeds a predetermined vehicle speed. The speed change control device for an infinitely variable speed ratio transmission according to claim 1, wherein the piston control means is selected. 4. The shift control means includes means for detecting or estimating the tilt angle of the power roller, and the control switching means sets the tilt angle when the first piston control means is selected. 2. The speed change control device for an infinitely variable speed ratio transmission according to claim 1, wherein the control is switched to the second piston control means when the speed exceeds or is about to exceed a predetermined range. 5. The control switching means selects the second piston control means when the target gear ratio exceeds a predetermined value, and selects the first piston control means when the target gear ratio is equal to or less than a predetermined value. The speed change control device for a continuously variable transmission with an infinite speed ratio according to claim 1. 6. The speed change control means includes means for detecting or setting a target driving force or load, and the control switching means changes the predetermined vehicle speed to a high speed side in accordance with an increase in the target driving force or load. The shift control device for a continuously variable transmission with an infinite speed ratio according to claim 3, wherein: 7. The power transmission mode switching means for selectively switching between the power circulation mode and the direct connection mode by selectively engaging the power circulation mode clutch and the direct connection mode clutch, wherein the control switching is performed. 2. The infinitely variable speed ratio transmission according to claim 1, wherein the means selects the first piston control means in the power circulation mode while selecting the second piston control means in the direct connection mode. Gear shift control device. 8. The apparatus according to claim 7, wherein the shift control means includes a mode switching valve for synchronously switching between a power circulation mode and a direct connection mode and switching between the first and second piston control means. 3. The shift control device for an infinitely variable speed ratio transmission according to claim 1. 9. The second piston control means continues control so as to always follow a target gear ratio or an actual gear ratio even when the control switching means selects the first piston control means. The shift control device for a continuously variable transmission with an infinite speed ratio according to claim 1, wherein: 10. The control apparatus according to claim 1, wherein the first piston control means continues to control so as to always follow the target gear ratio even when the control switching means selects the second piston control means. The speed change control device for an infinitely variable speed ratio transmission according to claim 1. 11. The shift control means includes means for detecting a tilt angular velocity of a power roller, and the control switching means sets a range of a tilt angle for switching from the first piston control means to the second piston control means. The speed change control device for an infinitely variable speed ratio transmission according to claim 4, wherein the speed is reduced according to the magnitude of the absolute value of the tilt angular velocity. 12. The shift control means includes means for detecting a tilt angular velocity of a power roller, and the control switching means includes an upper limit of a tilt angle range for switching from the first piston control means to the second piston control means. The value is decreased as the tilt angular velocity increases in the positive direction, while the lower limit of the range of the tilt angle is increased as the tilt angular velocity decreases in the negative direction. Item 5. A shift control device for an infinitely variable speed ratio continuously variable transmission according to item 4.