JP2000179669A - Control device of power train - Google Patents

Abstract

PROBLEM TO BE SOLVED: To stabilize the creep force during the stop in the traveling range without degradation of the durability by controlling the transmission gear ratio of a continuously variable transmission and the tightening force of a friction element to generate the specified creep force when the selection of the traveling range and the stop condition are detected. SOLUTION: A power train is provided with a control unit 300 which controls the transmission gear ratio of a continuously variable transmission and the tightening of a clutch using a hydraulic control circuit, and controls the transmission gear ratio on the whole. The control unit 300 controls transmission gear ratio control means (step motors) 251, 252 and tightening force regulating means (duty solenoids) 271, 272 so as to generate the specified creep force by setting the transmission gear ratio of the continuously variable transmission to be different from the specified value, and setting a friction element in a half- tightened condition when a range detecting means 304 detects the selection of the traveling range, and a stop detecting means 310 detects the stop condition.

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

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

[0002]

2. Description of the Related Art In recent years, a power train using a continuously variable transmission mechanism has been put to practical use as a power train mounted on a vehicle, and an example thereof is disclosed in Japanese Patent Application Laid-Open No. 9-210191. This includes a toroidal-type continuously variable transmission mechanism and a planetary gear mechanism, and a first path through the continuously variable transmission mechanism and the planetary gear mechanism as a power transmission path from the engine side to the drive wheel side; If a second path that passes only through the stepped transmission mechanism is provided, and the first path is set to a power transmission state by selective engagement of friction elements, the low mode is set to a low mode in which forward and reverse at a low gear ratio can be obtained. When the second path is set in the power transmission state, the high speed mode is set so that the high speed ratio can be obtained.

In this type of powertrain, a so-called geared neutral start system is sometimes used. That is, in the case of the above-described configuration, by controlling the speed ratio of the continuously variable transmission mechanism to the predetermined speed ratio in the state where the low mode is selected, it is possible to reduce the output speed to zero while the friction element is engaged. According to this, when the vehicle is stopped in a driving range such as the D range, the vehicle is stopped by setting the gear ratio to the predetermined gear ratio, and the gear ratio is changed from the state to the predetermined gear ratio. , The vehicle starts moving, and during that time, the engagement and release control of the friction element becomes unnecessary.

[0004]

As described above, when the continuously variable transmission mechanism is set in the geared neutral state while the vehicle is stopped in the traveling range, the vehicle is generally provided with an automatic transmission. The creep function disappears,
Problems such as giving the driver a sense of incongruity or making it impossible to obtain smooth start acceleration can be considered. On the other hand, when the vehicle is stopped in the traveling range, the speed ratio of the continuously variable transmission mechanism is set to a speed ratio different from the predetermined speed ratio in which the geared neutral state is set, so that the drive generated according to the difference is set. It is conceivable to use force as creep force.

However, if the creep force is to be generated by setting the speed ratio to a speed ratio different from the predetermined speed ratio as described above, the creep force fluctuates due to a subtle change in the speed ratio and the brake force is changed. When the pedal is depressed and the vehicle is to be kept stationary against this creep force, excessive sliding occurs in the friction portion of the continuously variable transmission mechanism or in the friction element in the engaged state. This would reduce the durability.

Accordingly, the present invention is to realize a power train using the above-described continuously variable transmission mechanism that stabilizes the creep force during a stop in a traveling range and does not reduce the durability. That is the task.

[0007]

In order to solve the above problems, the present invention uses the following means.

First, the invention described in claim 1 of the present application (hereinafter, referred to as a “first invention”) includes a first path passing through a continuously variable transmission mechanism and a gear mechanism, A second path that passes only through the step-variable mechanism, and a friction element that selectively turns these paths into a power transmission state, and the first path is set to a power transmission state by the friction element. In a power train configured to be able to realize a neutral state by controlling the speed ratio of the continuously variable transmission mechanism to a predetermined speed ratio, a range detection unit that detects selection of a range set for the vehicle; A stop ratio detecting unit that detects a stop state of the vehicle, a speed ratio variable unit that changes a speed ratio of the continuously variable transmission mechanism, and a fastening force adjusting unit that adjusts a fastening force of the friction element. When the selection of the traveling range is detected by the detection means and the stop state of the vehicle is detected by the stop detection means, the speed ratio of the continuously variable transmission mechanism is set to a speed ratio different from the predetermined speed ratio and the friction ratio is set. Control means for controlling the speed ratio variable means and the fastening force adjusting means is provided such that a predetermined creep force is generated by controlling the elements to a semi-fastened state.

According to a second aspect of the invention (hereinafter referred to as a "second invention"), in the first aspect, the control means detects the selection of the traveling range by the range detection means and stops the vehicle. When the stop state of the vehicle is detected, the speed ratio variable means is controlled to set the speed ratio of the continuously variable transmission mechanism to a second predetermined speed ratio different from the predetermined speed ratio at which the neutral state is obtained. This second
The creep force is controlled so that the creep force becomes substantially constant under the predetermined gear ratio.

According to a third aspect of the invention (hereinafter referred to as a "third invention"), the control means is the same as the first invention, wherein the selection of the traveling range is detected by the range detection means and the stop detection means is provided. When the stopped state of the vehicle is detected, the fastening force adjusting means sets the fastening force of the friction element to a predetermined fastening force in a semi-fastened state, and the creep force is substantially constant under the predetermined fastening force. The speed ratio of the continuously variable transmission mechanism is controlled by the speed ratio variable means so that the speed ratio becomes as follows.

Further, the invention according to claim 4 (hereinafter referred to as "the invention")
The "fourth invention" is characterized in that, in the third invention, a torque fluctuation detecting means for detecting a fluctuation of the input torque to the power train is provided, and the control means detects the selection of the traveling range by the range detecting means. When the stop state of the vehicle is detected by the stop detecting means, the predetermined engagement force for bringing the friction element into a semi-engaged state is set to be lower as the torque fluctuation detected by the torque fluctuation detecting means is larger. In addition, the speed ratio of the continuously variable transmission mechanism is controlled by the speed ratio variable means so that the speed ratio at which the creep force becomes substantially constant under the predetermined fastening force is obtained.

On the other hand, the invention according to claim 5 (hereinafter referred to as a "fifth invention") is characterized in that, in the first invention, an engine torque detecting means for detecting an engine output torque is provided, and the control means comprises: When the selection of the traveling range is detected by the range detecting means and the stop state of the vehicle is detected by the stop detecting means, the creep force is corrected according to the engine torque detected by the engine torque detecting means. As described above, at least one of the speed ratio of the continuously variable transmission mechanism and the fastening force of the friction element is corrected.

The invention according to claim 6 (hereinafter referred to as "the invention")
The "sixth invention" is the same as the first invention, further comprising a gradient detecting means for detecting a gradient of the traveling road, the control means being configured to detect the selection of the traveling range by the range detecting means, and to provide a stop detecting means. When the stationary state of the vehicle is detected, the speed ratio of the continuously variable transmission mechanism and the engagement of the friction element are adjusted so that the creep force is corrected according to the gradient of the traveling path detected by the gradient detecting means. It is characterized in that at least one of force is corrected.

With the above arrangement, the following effects can be obtained according to the present invention.

First, according to the first aspect of the present invention, when the traveling range is selected and the vehicle is stopped, the speed ratio of the continuously variable transmission is set to the geared neutral state while the friction element is engaged. Is set to a different gear ratio, a creep force is generated to drive the vehicle, and the depression of the brake pedal causes the vehicle to be stopped against this creep force. At this time, since the semi-fastened state is set, the creep force is absorbed by the sliding of the friction element, and the stopped state is maintained without difficulty.

According to the second invention, in the first invention, when controlling the speed ratio of the continuously variable transmission mechanism and the engaging force of the friction element to generate the creep force, the speed ratio is changed to the neutral state. Since the second predetermined speed ratio is set to be different from the obtained predetermined speed ratio, the fastening force of the friction element is controlled such that a substantially constant creep force is obtained under the predetermined speed ratio. The creep force is controlled smoothly and stably.

According to the third aspect, similarly to the second aspect, when controlling the speed ratio of the continuously variable transmission mechanism and the fastening force of the friction element in order to generate the creep force, the fastening force of the friction element is controlled. Is set to a predetermined fastening force that results in a semi-fastened state,
Since the speed ratio of the continuously variable transmission mechanism is controlled so that a substantially constant creep force is obtained under the predetermined fastening force,
This control of the creep force is performed with good responsiveness.

According to a fourth aspect, in the third aspect, when the fastening force of the friction element is set to a predetermined fastening force at which the friction element is in a semi-fastened state, the larger the fluctuation of the input torque to the power train, the more the above-mentioned. Since the predetermined fastening force is set low, this torque fluctuation is effectively absorbed, transmission of engine vibration to the vehicle body is suppressed, and
It is possible to prevent a decrease in durability such as acceleration of wear due to sliding of the friction element in accordance with torque fluctuation under relatively high fastening force.

On the other hand, according to the fifth aspect, when the vehicle is stopped in the traveling range, the speed ratio of the continuously variable transmission mechanism is set to a speed ratio different from the predetermined speed ratio in which the geared neutral state is established, and the friction element is brought into the semi-fastened state. When the required creep force is generated by controlling, the generated creep force is corrected according to the engine torque by correcting at least one of the gear ratio of the continuously variable transmission mechanism or the fastening force of the friction element. Thus, when the engine torque is reduced due to, for example, the operation of an auxiliary machine, the speed ratio of the continuously variable transmission mechanism is corrected in a direction in which the generation of the driving force from the mechanism is increased, or the friction element is fastened. By correcting the force in a direction to increase the force, an appropriate creep force can be maintained.

Similarly, according to the sixth aspect, when the vehicle is stopped in the traveling range, the speed ratio of the continuously variable transmission mechanism is set to a speed ratio different from the predetermined speed ratio in which the geared neutral state is established, and the friction element is partially engaged. By controlling at least one of the speed ratio of the continuously variable transmission mechanism and the fastening force of the friction element when the required creep force is generated, the generated creep force is adjusted according to the gradient of the road surface. Since the correction is performed, for example, when the creep force is insufficient due to a large gradient, the speed ratio of the continuously variable transmission mechanism is corrected in a direction in which the generation of the driving force from the mechanism is increased, or the frictional element is corrected. It is possible to maintain an appropriate creep force, for example, by correcting in a direction to increase the fastening force of.

[0021]

DESCRIPTION OF THE PREFERRED EMBODIMENTS Hereinafter, a power train according to an embodiment of the present invention will be described.

As shown in FIGS. 1 and 2, a power train 10 according to the present embodiment includes an input shaft 11 connected to an output shaft 2 of an engine 1 via a torsion damper 3, A hollow primary shaft 12 loosely fitted to the outside, and these shafts 11,
And a secondary shaft 13 arranged in parallel with the shaft 12, and these shafts 11 to 13 are all arranged to extend in the lateral direction of the vehicle.

The input shaft 11 and the primary shaft 1 in the power train 10
The first and second toroidal-type continuously variable transmission mechanisms 20 and 30 and a loading cam mechanism 40 that applies an axial load to these to enable power transmission are arranged on the second axis. At the same time, a planetary gear mechanism 50, a low clutch 60 and a high clutch 70 are disposed on the axis of the secondary shaft 13. And the axis of the input shaft 11 and the primary shaft 12,
A low mode gear train 80 and a high mode gear train 90 are interposed between the secondary shaft 13 and the axis.

The first and second continuously variable transmission mechanisms 20 and 30 have substantially the same configuration, and each of the input and output disks 21 and 31 and the output disk 2 has a toroidal surface.
2, 32, and between each of these opposed toroidal surfaces,
Two power rollers 23 and 33 for transmitting power between the two disks 21 and 22 and between the disks 31 and 32 are interposed.

The first continuously variable transmission mechanism 20 arranged farther from the engine 1 has an input disk 21 on the opposite side to the engine and an output disk 22 on the engine side.
The second continuously variable transmission mechanism 30 disposed closer to the engine 1 has an input disk 31 on the engine side and an output disk 32 on the opposite side to the engine. The input disks 21 and 31 are respectively coupled to both ends of the primary shaft 12, and the output disks 22 and 32 are integrated and rotatably supported by an intermediate portion of the primary shaft 12.

A first gear 81 constituting the low mode gear train 80 is connected to an end of the input shaft 11 on the side opposite to the engine, and the first gear 81 and the input disk 21 of the first continuously variable transmission mechanism 20 are connected to each other. The loading cam mechanism 40 is interposed between the output disk 2 and the output disk 2 integrated with the first and second continuously variable transmission mechanisms 20 and 30.
A first gear 91 constituting the high mode gear train 90 is provided on the outer periphery of the gears 2 and 32.

On the other hand, a second gear 82 constituting the low mode gear train 80 is rotatably supported at an end of the secondary shaft 13 on the side opposite to the engine.
3 and is connected to the first gear 81 via
The planetary gear mechanism 50 is disposed at an intermediate portion of the secondary shaft 13. And, the planetary gear mechanism 50
A low clutch 60 for connecting or disconnecting the pinion carrier 51 and the second gear 82 of the low mode gear train 80 is interposed between the pinion carrier 51 and the second gear 82 of the low mode gear train 80.

On the engine side of the planetary gear mechanism 50, a second gear 92 constituting a high mode gear train 90 is rotatably supported, and the first and second continuously variable transmission mechanisms 2 are provided.
The second gear 92 and the sun gear 52 of the planetary gear mechanism 50 are engaged with a first gear 91 provided on the outer periphery of the output disks 22 and 32 at 0 and 30. The internal gear 53 of the mechanism 50 is connected to the secondary shaft 13.
On the engine side of the planetary gear mechanism 50, a high clutch 70 for connecting or disconnecting the second gear 92 of the high mode gear train 90 and the secondary shaft 13 is provided.

Further, a differential device 5 is connected to an end of the secondary shaft 13 on the engine side via an output gear train 4 including first and second gears 4a and 4b and an idle gear 4c. Drive shafts 6a and 6b extending left and right from the differential device 5 are connected to left and right drive wheels (not shown).

An oil pump 100 is arranged at an end of the input shaft 11 on the side opposite to the engine, and is driven by the input shaft 11 via the first gear 81 of the low mode gear train 80.

Next, the first and second continuously variable transmission mechanisms 20,
The configuration of 30 will be described in more detail by taking the first continuously variable transmission mechanism 20 as an example.

As shown in FIG. 3, a pair of power rollers 2
The trunnions 2 and 3 are connected to the input and output discs 21 and 22 via shafts 24 and 24 extending substantially in the radial direction.
5, 25, respectively, and the input and output disks 21,
180 ° on the circumference of the opposing toroidal surfaces of 22
It is arranged on the opposite side in an almost horizontal posture and vertically parallel,
The two discs 2 are located at two positions 180 ° opposite each other on the circumferential surface.
They are in contact with the toroidal surfaces 1 and 22, respectively.

The trunnions 25, 25 are supported between left and right support members 26, 26 attached to the case 101 of the power train 10, and the two disks 2
Rotation about the axis X, X in the horizontal direction orthogonal to the shafts 24, 24 of the power rollers 23, 23 in the tangential direction of the power rollers 23, 23 and linear reciprocating motion in the direction of the axis X, X are enabled. ing. And these trunnions 25, 25
A rod 2 extending to one side along the axis X, X
A gear change control unit 110 for tilting the power rollers 23, 23 via these rods 27, 27 and trunnions 25, 25 is attached to the side surface of the case 101. ing.

The transmission control unit 110 has a hydraulic control unit 111 and a trunnion drive unit 112. The trunnion drive unit 112 is attached to the upper and lower trunnions 25, 25 in opposition to the rods 27, 27, respectively. The speed increasing and deceleration pistons 113 and 114 are arranged, and the opposing pistons 113 and 114 form speed increasing and deceleration hydraulic chambers 115 and 116, respectively.

In the trunnion 25 located above, the speed-increasing hydraulic chamber 115 is located on the power roller 23 side.
The deceleration hydraulic chamber 116 is disposed on the side opposite to the power roller 23, and for the trunnion 25 located below, the speed increasing hydraulic chamber 115 is positioned on the side opposite to the power roller 23.
Hydraulic chambers 116 for reduction are arranged on the power roller 23 side.

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

Here, the specific operation of the speed ratio control for the first continuously variable transmission mechanism 20 will be described. First, the upper and lower trunnions 25, 25 are controlled by the hydraulic control unit 111 shown in FIG.
Is increased relative to the deceleration hydraulic pressure PL supplied to the deceleration hydraulic chambers 116, 116 from a predetermined balanced state. And the upper trunnion 25 on the drawing,
To the right, the lower trunnions 25 will each move horizontally to the left.

At this time, the output disk 2 shown in FIG.
2 is rotating in the x direction, the upper power roller 23 moves to the right to move the output disk 22
, And receives an upward force from the input disk 21 which is on the near side of the drawing and is rotating in the anti-x direction. The lower power roller 23 receives an upward force from the output disk 22 and a downward force from the input disk 21 by moving to the left.

As a result, the upper and lower power rollers 23, 23
In both cases, the contact position with the input disk 21 is tilted so as to move outward in the radial direction, and the contact position with the output disk 22 is tilted so as to move inward in the radial direction. ).

Conversely, the upper and lower trunnions 2
The deceleration hydraulic pressure PL supplied to the 5, 25 deceleration hydraulic chambers 116, 116 is more relative to the speed increasing hydraulic pressure PH supplied to the speed increasing hydraulic chambers 115, 115 than a predetermined balanced state. When the upper trunnion 25 moves to the left side in the drawing and the lower trunnion 25 moves to the right side in the drawing, the upper power roller 23 applies an upward force from the output disk 22 to the input disk 21.
From the lower power roller 23
Receives a downward force from the output disk 22 and an upward force from the input disk 21. As a result, both the upper and lower power rollers 23, 23 are tilted so that the contact position with the input disk 21 moves inward in the radial direction and the contact position with the output disk 22 moves outward in the radial direction. The gear ratio of No. 20 increases (deceleration).

The supply control of the speed increasing and decelerating oil pressures PH and PL by the oil pressure control unit 111 will be described in detail in the explanation of the oil pressure control circuit described later.

The structure and operation of the first continuously variable transmission mechanism 20 are the same as those of the second continuously variable transmission mechanism 30. As shown in FIGS. 1 and 2, first and second continuously variable transmission mechanisms 2 are provided at both ends of a hollow primary shaft 12 loosely fitted on an input shaft 11.
The input disks 21 and 31 of 0 and 30 are spline-fitted, so that the input disks 21 and 31 always rotate in the same way.
Since the output disks 22 and 32 of both transmission mechanisms 20 and 30 are integrated, the rotational speeds on the output side of both transmission mechanisms 20 and 30 are always the same. Therefore, the gear ratio control of the first and second continuously variable transmission mechanisms 20 and 30 by the hydraulic control of the power rollers 23 and 33 as described above is also performed so that the gear ratio is always kept the same.

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

As shown in FIG. 4, the hydraulic control circuit 20
0, a regulator valve 202 that adjusts the pressure of hydraulic oil discharged from the oil pump 100 to a predetermined line pressure and outputs the adjusted line pressure to a main line 201;
A relief valve 204 for generating a predetermined relief pressure using the line pressure supplied from 01 as a source pressure and outputting the relief pressure to a relief pressure line 203, a D range, an R range, an N range, and a P range by a switching operation of a driver. And a manual valve 205 that allows the user to select one.

Of these valves, the manual valve 205 connects the main line 201 to the first and second output lines 206 and 207 in the D range, and to the first and third output lines 206 and 208 in the R range. In addition to the communication, the line pressure is cut off in the N range and the P range.

The regulator valve 202 and the relief valve 204 are provided with a line pressure control linear solenoid valve 209 and a relief pressure control linear solenoid valve 210, respectively. The linear solenoid valves 209 and 210 each generate a control pressure based on the constant pressure generated by the reducing valve 211. .

When these control pressures are supplied to the control ports 202a, 204a of the regulator valve 202 and the relief valve 204, the line pressure and the relief pressure are reduced by the linear solenoid valve 20.
9 and 210, respectively.

Further, the constant pressure generated by the reducing valve 211 is also guided to an on / off solenoid valve 213 for operating the fail-safe valve 212. The on / off solenoid valve 213 is normally turned on to reduce the above-mentioned constant pressure to the fail-safe valve 212.
Is supplied to the control port 212a of the
While the spool 12 is moved to the right, it is turned off at the time of fail-safe, and the constant pressure is drained from the control port 212a of the fail-safe valve 212, thereby moving the spool of the valve 212 to the left. I have.

In the hydraulic control circuit 200, a speed increasing hydraulic pressure PH and a decelerating hydraulic pressure PL are generated for shifting control based on the line pressure and the relief pressure at the time of forward movement and reverse movement, respectively. Forward three-layer valve 2
20 and retraction three-layer valve 230 and these three-layer valves 22
0, 230 is selectively operated.

In the shift valve 240, the position of the spool is determined by whether or not a line pressure is supplied as a control pressure to a control port 240a at one end. When the line pressure is not supplied, the spool moves to the right side. Position, the main line 201 is communicated with a line pressure supply line 241 communicating with the forward three-layer valve 220, and when line pressure is supplied, the spool is located on the left side and the main line 201 is retracted. It operates to communicate with a line pressure supply line 242 that communicates with the three-layer valve 230.

Here, the line pressure is supplied to the control port 240a of the shift valve 240 because the spool of the fail-safe valve 212 is normally located on the right side and the manual valve 205 is located in the R range. At this time, the line pressure is supplied from the manual valve 205 to the control port 240a of the shift valve 240 via the third output line 208 and the face safe valve 212. Then, even when the spool of the fail-safe valve 212 is located on the right side during normal times, when the manual valve 205 is located in the D range, the line pressure is not supplied to the control port 240a of the shift valve 240. . When the fail-safe control is executed, as described above, the on / off solenoid valve 213 is switched to the fail-safe valve 212.
When the operating pressure is drained from the control port 212a, the spoon of the valve 212 moves to the left, and the connection between the shift valve 240 and the third output line 208 is shut off. However, the line pressure is not supplied to the control port 240a of the shift valve 240.

Also, the three-way valve 2 for forward and backward movement
20 and 230 have the same configuration, and bores 221 and 23 are provided.
1 and the spools 223 and 233 are respectively fitted to the sleeves 222 and 232 so as to be movable in the axial direction. The valve body 111 of the hydraulic control unit 111 in the shift control unit 110 shown
a.

Further, line pressure ports 224, 234 to which line pressure supply lines 241, 242 led from the shift valve 240 are connected are provided at the center of these three-layer valves 220, 230. , The first and second relief pressure ports 225, 22 to which the relief pressure line 203 is branched and connected respectively.
6,235,236 are provided. Further, the line pressure ports 224 and 234 and the first relief pressure port 2
25, 235 between the pressure increasing ports 227, 237
However, between the line pressure ports 224 and 234 and the second relief pressure ports 226 and 236,
8, 238 are provided respectively.

Then, the forward and backward three-layer valve 22
Lines 243 and 244 respectively derived from the pressure-increasing pressure ports 227 and 237 of the three-layer valves 220 and 230 for forward and backward movement.
Lines 245 and 246 respectively led from the line 38 are connected to the shift valve 240. When the spool of the shift valve 240 is located on the right side, the speed increasing pressure port 227 and the deceleration of the forward three-layer valve 220 are reduced. Pressure port 2
Lines 243 and 245 derived from 28 are connected to a speed increasing line 247 and a deceleration line 248, respectively.
Hydraulic chambers 115 for increasing speed and hydraulic chamber 1 for deceleration
16 and 116, respectively.

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

The operation of these three-layer valves 220 and 230 will be described with reference to FIG. In FIG. 5,
The directions of the three-layer valves 220 and 230 are opposite to those in FIG.

As shown in the drawing, when the positional relationship between the sleeve 222 and the spool 223 is in the neutral position, for example, the sleeve 222 of the three-way forward valve 220 is relatively moved in the drawing.
Moving to the left, the communication between the line pressure port 224 and the speed-up pressure port 227 and the second relief pressure port 22
6 and the communication between the deceleration pressure port 228 increase,
Conversely, when the sleeve 222 moves relatively to the right, the degree of communication between the line pressure port 224 and the deceleration pressure port 228 and the degree of communication between the first relief pressure port 225 and the speed increasing pressure port 227 increase. Therefore, in the former case, the speed increasing hydraulic pressure PH rises and the deceleration hydraulic pressure PL increases.
In the latter case, the deceleration hydraulic pressure PL increases and the speed-increasing hydraulic pressure PH decreases.

The above operation is effected by the retreat three-layer valve 230.
The same applies to these three-layer valves 220 and 230.
Step motors 251 and 252 for operating the sleeves 222 and 232 of the three-way valves 220 and 230 are connected to the sleeves 222 and 232 of the forward and backward three-layer valves 220 and 230 via link members 253 and 254, respectively. I have.

The step motors 251, 251
A cam mechanism 260 is provided that moves the spools 223 and 233 in the axial direction against the spring force of the springs 229 and 239 in accordance with the movement of the sleeves 222 and 232 by the 52.

As shown in FIGS. 5 and 6, one end surface of the cam mechanism 260 is a spiral cam surface 261a, and a rod of the trunnion 35 located above the second continuously variable transmission mechanism 30 is provided. 37, and one end of spools 223, 233 of the three-way valves 220, 230 for forward and backward movement, which are arranged in a direction orthogonal to the ones of the spool body 223 and the valve body 11 of the hydraulic control unit 111.
A shaft 262 rotatably supported by the shaft 1a; a driven lever 263 attached to one end of the shaft 262 and having a swinging end abutting against the cam surface 261a of the precess cam 261; And the swing end is moved forward and backward by the three-layer valve 220, 2
30 are provided with drive levers 264 and 265 for forward and backward movement engaged with cuts 223 a and 233 a provided at one end of the spools 223 and 233.

When the upper power roller 33 in the second continuously variable transmission mechanism 30 is tilted by the control of the speed-increasing hydraulic pressure PH and the deceleration hydraulic pressure PL, the upper trunnion 35 and the upper trunnion 35 When the rod 37 rotates integrally around the axis X, the precess cam 261 also rotates integrally with these, and the cam surface 26
The driven lever 263 whose swinging end abuts on 1a swings by a predetermined amount, and the forward and backward drive levers 264 and 265 also swing by the same angle via the shaft 262,
As a result, the spools 223, 233 of the forward and backward three-layer valves 220, 230 move in the axial direction by an amount corresponding to the swing angle.

Therefore, these spools 223, 2
33 is the position of the power roller 3 of the second continuously variable transmission mechanism 30.
3 (and the power roller 23 of the first continuously variable transmission mechanism 20)
, In other words, the gear ratio of these transmission mechanisms 20 and 30.

Here, the control operation of the speed ratio of the continuously variable transmission mechanisms 20 and 30 (hereinafter referred to as “toroidal ratio”) is described as follows.
A description will be given of a case where the vehicle moves forward.

First, the control pressures of the regulator valve 202 and the relief valve 204 are generated by the line pressure control linear solenoid valve 209 and the relief pressure control linear solenoid valve 210 in the hydraulic control circuit 200, and the control pressure corresponding to the control pressure is generated. A line pressure and a relief pressure are generated.

Of these oil pressures, the line pressure changes from the main line 201 to the shift valve 240 and the line 24.
1 is supplied to the line pressure port 224 of the three-layer valve 220. The relief pressure is supplied to the first and second relief pressure ports 225 and 22 of the three-layer valve 220 through the line 203.
6. Then, based on the line pressure and the relief pressure, the step-up motor 251 controls the three-layer valve 220 to control the speed increasing hydraulic chamber 11 of the speed change control unit 110.
5, 115 and the pressure difference ΔP (= PH−PL) between the speed-up hydraulic pressure PH and the deceleration hydraulic pressure PL supplied to the deceleration hydraulic chambers 116, 116, respectively.

In this differential pressure control, as shown in FIG. 6, the trunnion 25 or the power roller 23 is held at a predetermined neutral position against the traction force T acting on the trunnion 25 of the continuously variable transmission mechanisms 20, 30. At the same time, the trunnion 25 and the power roller 23 are moved from the neutral position to the axis X.
This is performed in order to change the toroidal ratio by tilting the power roller 23 by moving it along the direction.

That is, in the transmission mechanisms 20 and 30, when the power rollers 23 and 33 are driven in the directions b and b by the rotation of the input disks 21 and 31 in the directions a and a, the power rollers 23 and 33 and A force acts on the supporting trunnions 25 and 35 to drag them in the same direction as the rotation directions b and b of the input disks 21 and 31. The rotation of the power rollers 23, 33 in the directions a, a causes the output disks 22, 32 to move in the direction c (x in FIG. 3).
Direction), a force in the opposite direction to the rotation direction c of the output disks 22, 32 acts on the power rollers 23, 33 or the trunnions 25, 35 as a reaction force. As a result, the traction forces T, T in the illustrated directions are applied to the power rollers 23, 33 and the trunnions 25, 35.
Will work.

Therefore, in order to hold the power roller 23 at the neutral position against the traction forces T, T, the speed increasing and deceleration hydraulic chambers 115, 116 provided in the respective trunnions 25, 35 have a difference. The pressure-increasing hydraulic pressure PH and the deceleration hydraulic pressure PL are supplied such that the pressure ΔP has a magnitude that balances the traction force T.

Then, for example, the toroidal ratio is reduced (increased speed) from this state, and the sleeve 222 of the forward three-layer valve 220 is driven by the step motor 251.
Is moved to the left in FIG. 5 (to the right in FIG. 4), the degree of communication between the line pressure port 224 and the speed-up pressure port 227 of the three-layer valve 220, and the second relief pressure port 2
When the degree of communication between the pressure increasing port 26 and the deceleration pressure port 228 increases, the speed increasing hydraulic pressure PH supplied to the speed increasing hydraulic chambers 115 from the speed increasing pressure line 247 shown in FIG.
Is increased, and the deceleration hydraulic pressure PL supplied from the deceleration pressure line 248 to the deceleration hydraulic chambers 116, 116 is reduced to increase the differential pressure ΔP. As a result, the differential pressure ΔP is increased.
P overcomes the traction force T and trunnion 2
5, 35 or the power rollers 23, 33 move in the directions d, d shown in FIG.

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

Further, the cam mechanism 26 is tilted by the tilting of the power roller 33 in the second continuously variable transmission mechanism 30 as described above.
0 in the same direction (e shown in FIG. 5).
Direction), the driven lever 263, the shaft 262, and the drive lever 264 of the cam mechanism 260 rotate in the direction f shown in FIG.

As a result, the spool 223 of the three-layer valve 220
Moves in the g direction, that is, the left direction in FIG. 5 by the spring force of the spring 229. This direction is the direction in which the sleeve 222 is moved by the step motor 251. Therefore, as described above, Once the increased degree of communication between the line pressure port 224 and the speed increasing pressure port 227 and the degree of communication between the second relief pressure port 226 and the deceleration pressure port 228 return to the initial neutral state.

As a result, the differential pressure ΔP is balanced with the traction force T again, and the above-described speed change operation is completed, and the speed ratio of the continuously variable transmission mechanisms 20, 30, ie, the toroidal ratio is changed by a predetermined amount. Will be fixed above.

In this case, the speed change operation ends when the spool 223 moves to a position where the spool 223 moves to a predetermined neutral position in relation to the sleeve 222. 222 is moved to the position where the cam mechanism 26 is moved.
0 and the power rollers 23 and 33 and the trunnion 2
The position of the sleeve 222 corresponds to the tilt angle of the power rollers 23 and 33 and the trunnions 25 and 35 because the position is associated with the tilt angle of the trunnions 25 and 35. As a result, the control amount of the step motor 251 becomes the first, the second
This corresponds to the speed ratio of the continuously variable transmission mechanisms 20 and 30, and the pulse control of the step motor 251 controls the toroidal ratio.

The above operation is performed by the step motor 251.
The same operation is performed when the sleeve 222 of the three-layer valve 220 is moved to the opposite side.
Gear ratio increases (deceleration).

Here, the characteristics of the change of the toroidal ratio with respect to the number of pulses of the control signal output to the step motors 251 and 252 are as shown in FIG. 7, for example, and the speed ratio becomes smaller as the number of pulses increases. (High side).

On the other hand, as shown in FIG. 4, the hydraulic control circuit 200 has two duty ratios for controlling the low clutch 60 and the high clutch 70 in addition to the above-described structure for controlling the gear ratio. Solenoid valves 271, 2
The first output line 206 led from the manual valve 205 is connected to the low clutch duty solenoid valve 271 and the second output line 20
7 is a high clutch duty solenoid valve 272
Connected to each other.

Then, the line pressure from the first output line 206 is adjusted by the low clutch duty valve 271 to generate the engagement pressure of the low clutch 60 (hereinafter, referred to as “low clutch pressure”), which is a Felsafe pressure. By being supplied to the hydraulic chamber of the low clutch 60 via the valve 273 and the low clutch line 274,
The low clutch 60 is fastened with a fastening force corresponding to the magnitude. Further, by operating the high clutch duty solenoid valve 272, the line pressure from the second output line 207 is adjusted to generate the engagement pressure of the high clutch 70 (hereinafter referred to as “high clutch pressure”), which is By being supplied to the hydraulic chamber of the high clutch 70 via the line 275, the high clutch 70 is engaged with an engagement force according to the size.

In this case, the duty solenoid valves 271 and 272 do not output the clutch pressure when the duty ratio of the control signal is 100% (fully closed), and reduce the line pressure supplied when the duty ratio of the control signal is 0%. Output as it is as clutch pressure (fully open). At an intermediate duty ratio, a clutch pressure corresponding to the value is generated.

The low clutch line 274 and the high clutch line 275 are provided with accumulators 276 and 277, respectively, so that the supply of the engagement pressure to the low clutch 60 and the high clutch 70 is performed gently. The generation of a shock when the clutches 60 and 70 are engaged is suppressed.

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

The power train 10 according to this embodiment
Has the above-described mechanical configuration and the hydraulic control circuit 200, and uses the hydraulic control circuit 200 to
A control unit that controls the speed ratio (hereinafter, referred to as “unit ratio”) of the entire power train 10 by performing the speed ratio control of the second continuously variable transmission mechanisms 20 and 30 and the engagement control of the clutches 60 and 70. 300
Is provided.

As shown in FIG. 8, the control unit 300 includes a vehicle speed sensor 301 for detecting the vehicle speed of the vehicle, an engine speed sensor 302 for detecting the engine speed, and a throttle opening for detecting the throttle opening. Sensor 303, selected range sensor 304 for detecting the range selected by the driver, hydraulic control circuit 200
A line pressure sensor 305 for detecting the line pressure, an oil temperature sensor 306 for detecting the temperature of the hydraulic oil, and an input speed for detecting the input speed and the output speed of the first and second continuously variable transmission mechanisms 20 and 30, respectively. A sensor 307 and an output speed sensor 308, an idle switch 309 for detecting non-depression of an accelerator pedal, a brake switch 310 for detecting depression of a brake pedal, a steering angle sensor 311 for detecting a steering angle of the steering wheel, and a gradient of a road surface of the vehicle. Is input from a gradient sensor 312 or the like for detecting the

Then, these sensors and switches 301
312, the linear solenoid valves 209 and 210 for line pressure control and relief pressure control, the on / off solenoid valve 213 for fail safe, the low clutch 60 and the high clutch 70, duty solenoid valves 271 and 272, first and second on-off valves 283 and 284 for lubrication control, and step motors 251 and 251 for the forward three-layer valve 220 and the reverse three-layer valve 230.
252 and the like.

Here, the control unit 300
The control of the unit ratio according to will be briefly described.

First, in the N range, both the low clutch 60 and the high clutch 70 are disengaged. Therefore, the power transmitted from the input shaft 11 side to the secondary shaft side is not transmitted to the planetary gear mechanism 50 and the secondary shaft 13, and therefore, the power is not output from the differential device 5 to the drive wheels. Absent.

At this time, in the planetary gear mechanism 50,
The sun gear 52 is driven by the power from the high mode gear train 90, but the power from the low mode gear train 80 is only transmitted to the rotating member 60a on the input side of the low clutch 60, and is not transmitted to the pinion carrier 51, Also,
Since the internal gear 53 connected to the secondary shaft 13 is fixed, the pinion carrier 51
Is in a state where the sun gear rotates in a no-load state in conjunction with the rotation of the sun gear.

In this state, the pinion carrier 5 is set by setting the toroidal ratio to a predetermined value.
1 can be controlled to a speed at which the engagement of the low clutch 60 and the rotation speed of the output-side rotating members 60a and 60b become equal. In other words, by controlling the toroidal ratio to the predetermined value, the low speed can be controlled. Even if the clutch 60 is connected, the rotation of the internal gear 53 or the secondary shaft 13 can be reduced to zero, and a so-called geared neutral state can be obtained.

In the power train 10, as described later, in the N range in which the low clutch 60 and the high clutch 70 are disengaged, the toroidal ratio is a value at which the geared neutral state is obtained (hereinafter referred to as "GN ratio"). ) Is not controlled, but is controlled to a different value.

When starting from the N range, D
When the driving range is switched to a driving range such as a range, the toroidal ratio is controlled so as to approach the GN ratio, and the low clutch 60 is engaged when the GN ratio or a value close to the GN ratio is reached. At this time, the unit ratio becomes infinite, as indicated by reference numerals a and b in FIG. 9, but if the number of control signal pulses for the step motors 251 and 252 is reduced from this state, the toroidal ratio increases. The internal gear 53 of the planetary gear mechanism 50 starts rotating in the forward direction by changing to the direction (low side) and decreasing the input rotation speed to the sun gear 52. This realizes the forward low mode characteristic L in which the unit ratio decreases as the number of pulses decreases.

In the forward low mode, the unit ratio gradually decreases from the start, and when a predetermined switching ratio is reached, as shown by reference numeral c in FIG. 70, the power from the input shaft 11 is transmitted to the secondary shaft 13 via the first and second continuously variable transmission mechanisms 20, 30, the high mode gear train 90, and the high clutch 70.
Will be transmitted to In this state, assuming that the gear ratio of the high mode gear train 90 is 1, the unit ratio is equal to the toroidal ratio, and the high mode characteristic H identical to the characteristic of the toroidal ratio shown in FIG. 7 is realized. Become.

It should be noted that the sun gear 5 is changed from the above-described geared neutral state by increasing the number of control signal pulses for the step motors 251 and 252 to decrease the toroidal ratio (high side).
2 is increased, the planetary gear mechanism 50
The internal gear 51 starts to rotate in the reverse direction, and a low mode characteristic R in the R range in which the absolute value of the unit ratio decreases as the number of pulses increases is obtained.

Next, of the above-described shift control, the control operation from switching from the N range to the driving range such as the D range until the vehicle starts moving will be described in detail.

This control is performed in accordance with the flowcharts of FIGS. 10 and 11 and the time chart of FIG. 12. First, in step S1, each sensor and switch 3 shown in FIG.
01 to 312, the toroidal ratio R at present, the throttle opening TVO of the engine, the range selected by the driver, the hydraulic control circuit 20
At 0, various state quantities such as the line pressure PL, the hydraulic oil temperature TMP, the road surface gradient α, the steering angle β, and the ON / OFF state of the idle switch 309 and the brake switch 310 are detected.

Next, at step S2, the line pressure PL
Is greater than or equal to a predetermined value PL1, that is, whether the line pressure PL is adjusted to an appropriate value. And PL ≧ PL
If 1, and the following control can be performed correctly, then it is determined in step S3 whether the selected range is the N range. Control when the line pressure PL is less than the predetermined value PL1 will be described later.

If the selected range is the N range, the target value R0 of the toroidal ratio R is determined in step S4.
Is set to the target ratio Ra for the predetermined N range, and in step S5, the duty ratio D output to the duty solenoid valve 271 for the low clutch 60 is set to a predetermined value Da%.

As shown in FIGS. 12 and 13, the target ratio Ra in the N range is a predetermined amount lower (lower gear ratio side) than the GN ratio Rgn, and
It is set to a value on the high side (high gear ratio side). The duty ratio Da% output to the duty solenoid valve 271 is 100%.
1 is in a fully closed state.

FIG. 13 shows the relationship between the toroidal ratio R and the output torque of the power train 10 (hereinafter referred to as “unit torque”) in the low mode. The toroidal ratio R is shifted from the GN ratio Rgn to the low side (from the GN ratio Rgn). When changing to the forward side, the unit torque once increases from 0 and then gradually decreases.

Next, in step S6, the above-mentioned step S
The toroidal ratio (hereinafter, also referred to as “actual ratio”) R detected in step 1 is compared with the target ratio R0 set in step S4, and the target ratio R0 of the actual ratio R is compared.
Is calculated, and in step S7,
A gain G set in advance is set according to the input torque to the power train 10 and an operation mode such as forward, reverse, low mode or high mode, and step S8.
Then, based on the gain G and the deviation ΔR, the number of pulses N to be output to the step motors 251 and 252 is determined from a map set in advance as shown in FIG.

Here, in the above map, as the absolute value of (deviation ΔR × gain G) increases, the number of pulses N to be output increases.
When the deviation ΔR is positive (when the actual ratio R is a value lower than the target ratio R0), the number N of output pulses is increased to increase the actual ratio R.
Is corrected to the high side, and conversely, when the deviation ΔR is negative (when the actual ratio R is a value higher than the target ratio R0),
It is set so that the actual ratio R is corrected to the low side by decreasing the number N of output pulses.

Then, in step S9, the signal of the pulse number N thus obtained is output to the stepping motor 251,
252. As a result, the step motors 251 and 252 or the sleeves 222 and 232 of the three-layer valves 220 and 230 are driven so that the deviation ΔR is eliminated, and the steps S6 to S9 are repeatedly executed. Ratio R is the target ratio R0,
That is, N set to a predetermined amount low side from the GN ratio Rgn
The target ratio Ra of the range is set by feedback control.

In step S10, the duty ratio D = Da (10
By outputting the 0)% signal, the low clutch pressure is set to 0, and the low clutch 60 is released.

Further, in step S11, an idle speed control valve (hereinafter referred to as "ISC
Control).

As shown in FIG. 15, the ISC valve 400 adjusts the opening degree of a bypass passage 403 that bypasses a throttle valve 402 provided in an intake passage 401 of the engine. The engine output is controlled as necessary by adjusting the intake air amount in the idle state. In the N range, predetermined control is performed according to the engine load and the like.

As described above, in the N range,
When the low clutch 60 (and the high clutch 70) is released, the toroidal ratio R is set to a predetermined target ratio Ra on the lower side of the GN ratio Rgn, but by setting the toroidal ratio R, the N range is set. , The occurrence of noise such as rattling noise in the power train 10 is suppressed.

Here, the noise will be described. In the N range where the low clutch 60 and the high clutch 70 are released, the rotation of the input shaft 11 causes the gears 81 to 83 of the low mode gear train 80 to the input side of the low clutch 60 to rotate. Of the high mode gear train 90 by the output rotation of the first and second continuously variable transmission mechanisms 20 and 30.
The sun gear 52 and the pinion carrier 51 of the planetary gear mechanism 50, and the rotation member 60b on the output side of the low clutch 60
(See FIG. 1), and rotation from both these systems is met by the low clutch 60.

In this case, if the toroidal ratio R is set to the GN ratio Rgn, the low clutch 60
Since the rotating members 60a and 60b on the input and output sides rotate at a constant speed, they do not apply a load to each other, and both rotating systems rotate without any load. For this reason, noise due to rattling such as rattling noise is generated at a gear meshing portion, a spline engaging portion, and the like in both rotation systems.

On the other hand, when the toroidal ratio R is set to a ratio Ra different from the GN ratio Rgn in the N range as described above, the relative rotation between the rotating members 60a and 60b on the input and output sides of the low clutch 60 is increased. For this reason, the rotating systems from both sides exert a load on each other due to, for example, resistance due to the viscosity of the hydraulic oil between the input and output friction plates. As a result, the rattling at the meshing portion of the gears of the two systems is absorbed, and the generation of noise such as rattling noise is reduced.

On the other hand, when the selected range is switched from the N range to another range by the driver's operation, the process proceeds from step S3 to step S1 in the flowchart of FIG.
Step 2 is executed to determine whether or not the selected range is a driving range such as the D range. When the driving range is selected, it is further determined in step S13 whether the idle switch 309 is ON or OFF. Then, when the idle switch 309 is OFF, that is, when the accelerator pedal is depressed and the vehicle is not in the idle state and the vehicle shifts to the acceleration state, the start acceleration control in step S14 is executed according to a separately set program.

On the other hand, when the idle switch 309 is
If N, that is, if it has been switched to the travel range, but the accelerator pedal has not yet been depressed, it is determined in step S15 whether or not the current control cycle is the cycle immediately after switching to the travel range, In the case of the cycle immediately after the switching, the first timer T1 is set in step S16. Then, in step S17, it is determined whether or not the measurement time t1 of the first timer T1 has exceeded the first predetermined time Ta. Until the measurement time t1 exceeds the first predetermined time Ta, the duty ratio D of the low clutch duty solenoid valve 271 is set to 0 in step S18. %, And in the aforementioned step S10,
The signal having the duty ratio of 0% is output to the duty solenoid valve 271.

As a result, as indicated by reference numeral f in FIG. 12, the first predetermined time T
Until a elapses, the low clutch 60 is precharged with hydraulic oil. That is, the duty solenoid valve 271 is fully opened, and the hydraulic chamber of the low clutch 60 is rapidly filled with the working oil. As a result, the control of the low clutch pressure thereafter is performed with good responsiveness. Here, when the first predetermined time Ta elapses, the duty ratio D is temporarily returned to the duty ratio Da% (100%) before switching to the traveling range.

If the first predetermined time Ta has elapsed, in step S19, the measurement time t1 of the first timer T1 is calculated.
Is determined to have exceeded a second predetermined time Tb (> Ta), and step S20 is executed until it exceeds the second predetermined time Tb (> Ta). That is, after the completion of the pre-charging of the hydraulic oil, the duty ratio D of the duty solenoid valve 271 is changed from Da% (100%) to Db by taking a time (Tb-Ta) until the second predetermined time Tb is reached. % Is set so as to decrease linearly (refer to the reference numeral in FIG. 12). Then, in step S10, the signal of the duty ratio D is output to the duty solenoid valve 271.

As a result, the low clutch pressure is increased from 0 to the predetermined value Pb over the second predetermined time Tb from the time of switching to the travel range, as indicated by the reference character C in FIG. In this case, the predetermined value Pb is set to a hydraulic pressure at which the low clutch 60 enters the half-clutch state. As described above, since the hydraulic oil is precharged when switching to the travel range, the low clutch pressure is increased to the predetermined value Pb without delay.

On the other hand, in the control of the toroidal ratio R, which is performed in parallel with the low clutch pressure control, in step S21, the second predetermined time Tb is set after the first timer T1 is switched to the travel range. Over
The target ratio R0 is set so as to change from the value Ra for the N range to a predetermined value Rb equal to or close to the GN ratio Rgn. Then, the aforementioned step S6
From S9 to S9, the actual ratio R becomes equal to the target ratio R0 (= R
The feedback control by the step motor 251 is executed so as to become b).

As a result, the toroidal ratio R is
As indicated by the symbol in FIG. 2, the value Ra in the N range
Over a predetermined time Tb, the value changes to a predetermined value Rb that is equal to or close to the GN ratio Rgn. Then, when the second predetermined time Tb has elapsed, the low clutch pressure is increased to the predetermined value Pb as described above, and the low clutch 60 is engaged in a half-clutch state.

That is, the low clutch 60 is set so that the power train 10 is in a geared neutral state or a state close to the geared neutral state, and the rotational speed transmitted from the input shaft 11 to the input side rotary member 60a via the low mode gear train 80 is determined. In this state, the rotational speed transmitted through the continuously variable transmission mechanisms 20 and 30, the high mode gear train 90, and the planetary gear mechanism 50 to the output side rotation member 60b is substantially equalized. As a result, the occurrence of shock when the low clutch 60 is engaged is suppressed.

If it is determined in step S19 that the measurement time t1 of the first timer T1 has exceeded the second predetermined time Tb, the process proceeds to step S2 in the flowchart of FIG.
Step 2 is executed to determine whether or not the time of the first timer T1 has just expired. If so, the second timer T2 is set in step S23.

Then, in step S24, it is determined whether or not the measured time t2 of the second timer T2 has exceeded the third predetermined time Tc. Until the measured time t2 exceeds the third predetermined time Tc, the target time Tc is multiplied by the time Tc in step S25. The ratio R0 is set so as to change from a predetermined value Rb equal to or close to the above-described GN ratio Rgn to a low-side predetermined value Rc. Also,
In step S26, the duty ratio D of the duty solenoid valve 271 is set to the predetermined value Db%.

Therefore, by the feedback control in steps S6 to S9 described above, the toroidal ratio R takes the third predetermined time Tc from the predetermined value Rb to the predetermined value Rc on the low side over a third predetermined time Tc, as shown by a symbol C in FIG. And during this time, the low clutch pressure P is maintained at the predetermined pressure Pb that brings the low clutch 60 into the half-clutch state.

Further, the measurement time t of the second timer T2
If 2 exceeds the third predetermined time Tc, step S24
From step S27 to the brake switch 310
Is determined to be ON or OFF. And this switch 3
When 10 is ON and the vehicle is still in a stopped state, the stopped creep force control in step S28 and thereafter is executed.

That is, first, in step S28, the target ratio R0 is held at the value Rc set as described above, and in step S29, the engine torque is estimated based on the throttle opening TVO and the like.
Based on (Rc) and the estimated engine torque, a target value Pc of the low clutch pressure P is obtained from a preset map as shown in FIG. Then, in step S31, a predetermined correlation function [D = f (P)] between the low clutch pressure P and the duty ratio D of the duty solenoid valve 271.
Is used to set the duty ratio D at which the target low clutch pressure Pc is obtained.

Here, the map shown in FIG. 16 shows the relationship between the toroidal ratio R and the low clutch pressure P for realizing a constant creep force, and the toroidal ratio R
Acts in the direction in which the unit torque in the forward direction decreases as the distance from the GN ratio Rgn toward the forward side (low side) decreases (see FIG. 13), whereas the lower the low clutch pressure P, the lower the The engagement force of the clutch 60 increases and acts in a direction to increase the unit torque. Therefore, in the map of FIG. 16 in which the creep force is kept constant, the low clutch pressure P is set to increase as the toroidal ratio R decreases. In this map, the target creep force is corrected according to the engine torque estimated in step S29 and the road surface gradient α detected in step S1.

Then, as indicated by arrows X and X, the low clutch pressure P corresponding to the target ratio Rc set in step S28 is set to the reference characteristic (1) or the reference characteristic (1) for keeping the creep force constant. Correction characteristics (2), (3) for correcting constant creep force according to engine torque and gradient α
To obtain the low clutch pressure P.

Thereafter, according to steps S6 to S9 in FIG. 10, feedback control of the step motor 251 is performed so that the target ratio R0 (= Rc) is obtained, and at step S10, the map and the function are converted as described above. By outputting a signal of the duty ratio D obtained using the duty solenoid valve 271, the low clutch pressure P controlled to the target clutch pressure Pc for obtaining a constant creep force is supplied to the low clutch 60.
Thereby, the creep force is maintained at the optimum value during the period shown in FIG.

At this time, the low clutch pressure P (P
Since the low clutch 60 is set in the range where the low clutch 60 is in the half clutch state, the driving force in the forward direction generated by setting the toroidal ratio R to the ratio Rc different from the GN ratio Rgn as described above. That is, the creep force is absorbed by the slip of the low clutch 60, and the stationary state can be maintained with a constant creep force regardless of the generation of the driving force.

Then, by appropriately controlling the low clutch pressure Pc or the engaging force of the low clutch 60 in that case, the creep transmitted to the driver as the strength of the depressing force of the brake pedal for maintaining the stopped state. The magnitude of the force can be set appropriately, a good stopping feeling can be obtained, and especially, the toroidal ratio R
Is controlled to be constant, and the creep force is adjusted by the engagement force of the low clutch 60, so that the adjustment of the creep force can be performed smoothly and stably.

Here, as described above, in the map of FIG. 16 showing the relationship between the toroidal ratio R and the low clutch pressure P for keeping the creep force constant, the map is based on the road surface gradient α and the engine torque. Thus, the strength of the creep force that is kept constant is corrected. In other words, if the relationship between the toroidal ratio R and the low clutch pressure P is uniquely set in order to always keep the creep force constant, the creep force is increased when the engine torque is increased due to idle-up or the like in a cold state. It becomes too strong, and the slope α of the road surface results in insufficient creep force on an uphill slope and excessive creep on a downhill slope.

Therefore, when the engine torque is high or on a downhill, the correction is made so as to reduce the target constant creep force by using the correction characteristic (2) in FIG. When the distance is small or on an uphill, the correction is performed so as to increase the target constant creep force by using the correction characteristic (3) in FIG. As a result, the creep force is always set appropriately so as to meet the driver's requirements and feeling.

Next, as described above, after switching to the travel range, the vehicle goes through the stopped state by depressing the brake pedal, and then releases the brake pedal when starting, so that the brake switch 310 is turned off.
Steps S27 to S32 and S33 of the flowchart of FIG.
In the control cycle immediately after turning OFF, the third timer T
Set 3

Then, in step S34, this timer T
It is determined whether or not the measured time t3 of Step 3 exceeds the fourth predetermined time Td. Until the measured time t3 exceeds the fourth predetermined time Td, the duty ratio D of the duty solenoid valve 271 for the low clutch is set to the value during the above creep force control in Step S35. D [= f (P
c)] is set to a smaller predetermined value Dd, the number of pulses N output to the step motor 251 is set to 0 in step S36, and the position of the sleeve 222 in the three-layer valve 220, that is, the toroidal ratio R is set to the brake switch 31.
The value is maintained at the value Rc during creep force control before 0 becomes OFF.

In other words, during this predetermined time Td, FIG.
The feedback control of the toroidal ratio R is stopped to fix the ratio R to the value Rc immediately before the brake pedal is released, and the low clutch pressure P is increased by a predetermined amount as shown by the symbol And low clutch 6
Increase the fastening force of 0. Then, in step S37, IS
The duty ratio Disc output to the C valve 400 is increased by a predetermined amount ΔD1, and the opening of the ISC valve 400 is increased.

As described above, the ISC valve 400 adjusts the opening degree of the bypass passage 403 that bypasses the throttle valve 402 provided in the intake passage 401 of the engine, so that the throttle valve 402 is in the fully closed idle state. To adjust the engine torque at
The opening degree increases as the input duty ratio Disc increases, and the engine torque during idling increases.

As described above, the increase in the engagement force of the low clutch 60 and the increase in the engine torque provide a required acceleration force at the time of starting, and during that time, the feedback control of the toroidal ratio R is prohibited. Therefore, a smooth and powerful start acceleration can be obtained.

Thereafter, the measurement time t of the third timer T3
If 3 exceeds the fourth predetermined time Td, then step S38
The measured time t3 is the fifth predetermined time Te (Te>T>
Determine if d) has been exceeded. Until the fifth predetermined time Te is exceeded, in step S39, FIG.
2, the time between the two (Te-Td)
To linearly change the duty ratio D from the predetermined value Dd to a smaller predetermined value De.

As a result, the low clutch pressure further increases, and the low clutch 60 is almost completely engaged.
In this case, in order to reduce the shock when the low clutch 60 is completely engaged, the brake switch 31
The fourth predetermined time Td until the clutch 60 is shifted to the fully engaged state after the OFF of 0 is set as follows.

That is, as described above, when the engaging force of the low clutch 60 is increased with the target toroidal ratio R0 fixed at the predetermined value Rc at the time of starting, the continuously variable transmission mechanism 20,
Due to the bending of each part due to an increase in the transmission torque inside the motor 30, there is a phenomenon that the toroidal ratio R changes to the GN ratio Rgn as shown by reference numeral C in FIG. The rotation speeds of the input and output side rotation members 60a and 60b approach, and the relative rotation decreases.

Therefore, the fourth predetermined time Td is set from the time when the brake switch 310 is turned off to the time when the toroidal ratio R becomes closest to the GN ratio Rgn due to the above-mentioned phenomenon caused by the increase in the low clutch pressure P, After that time, the low clutch 60 is shifted to the fully engaged state. As a result, the clutch 60 is completely engaged in a state where the relative rotation of the on / off rotation members 60a and 60b is small, and the engagement shock is reduced.

Here, the phenomenon in which the toroidal ratio R changes to the GN ratio Rgn with an increase in the transmission torque will be described by taking the first continuously variable transmission mechanism 20 as an example.

In the low mode, the torque from the engine is applied to the input shaft 11 and the low mode gear train 8.
0, the torque as a reaction force generated by the planetary gear mechanism 50 on the secondary shaft 13 is transmitted to the secondary shaft 13 via the high mode gear train 90. Output disk 22,3
2 is recirculated in the direction indicated by the arrow c in FIG. 6 to generate a so-called circulating torque, which acts on the trunnions 25 and 35 as traction force in the direction indicated by the arrow d opposite to the traction force T described above. Become.

At this time, hydraulic pressure is supplied to the hydraulic chambers 115 and 116 connected to the trunnions 25 and 35 so as to be able to oppose such traction force.
13 and 114 are held at predetermined positions, and the power rollers 23 and 33 are
Rods 27, 37, trunnions 25,
35, deflection or deformation of the roller support shafts 24, 34, etc., or rattling at these connecting portions, etc.,
The power rollers 23 and 33 are displaced by being dragged by the traction force in the direction of the arrow d, so that the toroidal ratio R changes to the high side, that is, the direction approaching the GN ratio Rgn.

By shifting the low clutch 60 to the fully engaged state with the toroidal ratio R approaching the GN ratio in this manner, the occurrence of the engagement shock is suppressed as described above. .

As described above, the brake switch 31
The duty ratio D is reduced from the predetermined value Dd to the predetermined value De from the time when the fourth predetermined time Td elapses to the time when the fifth predetermined time Te elapses after the OFF of 0 to increase the low clutch pressure. Thus, the low clutch 60 is controlled to be completely engaged. At this time, the toroidal ratio R takes the time (Te-Td) as shown by the symbol S in FIG. From the ratio Rc in the middle to greater than the ratio Rc (low side)
In step S40, the ratio is changed to the predetermined ratio Re.
Then, a pulse signal is output to the step motor 251 at a constant frequency. In step S41, the above-mentioned IS
Duty ratio Disc of signal output to C valve 400
Is further increased by a predetermined amount ΔD2.

As a result, the low clutch 60 is completely engaged to generate a large driving force, and the unit ratio is quickly changed to the high side, and the engine torque is further increased during that time. Become,
The vehicle shifts smoothly and quickly from the starting state to the creep running state. In that case, during this time (T
Since the control of the toroidal ratio R in e-Td) is performed by the feedforward control that outputs a pulse signal at a constant frequency as described above, there is no delay despite the relatively large change amount of the toroidal ratio R. It will shift to the predetermined value Re.

The increase ΔD2 of the duty ratio Disc with respect to the ISC valve 400 is equal to the time (Te−T
This is set in accordance with the amount of change (Re-Rc) of the toroidal ratio R to the low side during d). Therefore,
For example, when the amount of change in the toroidal ratio R (Re-Rc) is large and the unit ratio changes to the high side relatively quickly, the engine torque is insufficient and the required acceleration force cannot be obtained, or conversely, the toroidal ratio The amount of change in the ratio R (R
The amount of increase in the engine torque with respect to e-Rc) does not become excessive, and an unnecessarily large acceleration force is not generated.

In order to precisely control the low clutch pressure P during the time Td and (Te-Td) when the low clutch 60 is shifted from the half-clutch state to the fully engaged state, the duty solenoid valve 27
The duty ratios Dd to De output to 1 are corrected by the hydraulic oil temperature TMP.

This correction is made according to the map shown in FIG. 17, and the duty ratio D is made larger as the oil temperature TMP of the hydraulic oil becomes lower. This is the oil temperature TMP
Is low, the duty ratio D is the same as when it is high.
To compensate for the tendency that the working pressure (low clutch pressure) actually generated increases.
Regardless of the level of the oil temperature TMP, almost the same low clutch pressure P is always obtained under the same conditions.

As described above, when the vehicle starts by releasing the brake pedal, creep running is performed until the driver depresses the accelerator pedal. At this time, the creep vehicle speed V is controlled. Is performed. Next, the control of the creep vehicle speed will be described.

This control is performed at the step S38 in the third
The measurement time t3 of the timer T3 is started when it is determined that the fifth predetermined time Te has elapsed since the brake switch 310 was turned off.
Then, by setting the duty ratio D of the duty solenoid valve 271 to 0%, the low clutch pressure P is set to the maximum value, and in step S43, the target creep vehicle speed V0 is set.

The target creep vehicle speed V0 is set according to the road surface gradient α and the steering angle β. First, the gradient α is determined based on the characteristics shown in the map of FIG. Is set.

In this map, the road surface gradient α is plus,
That is, in the case of an uphill, the target creep vehicle speed V0 is set to be low in accordance with the gradient α, and in the case of a downhill, that is, in the case of a downhill, the target creep vehicle speed V0 is set to be high according to the gradient α. Have been. Therefore, in the case of an uphill, the vehicle speed is lower than that in the case of a flat road, and in the case of a downhill, the vehicle speed becomes faster than that of a flat road, and the degree thereof corresponds to the magnitude of the gradient α. A creep vehicle speed V suitable for the driver's feeling during creep running on a road can be obtained.

The characteristic set in this map is particularly based on the characteristic of the creep vehicle speed on a slope road of a vehicle (hereinafter referred to as an "AT vehicle") equipped with a conventional automatic transmission with a torque converter indicated by a chain line. The amount of change in the vehicle speed V with respect to the gradient is suppressed, thereby preventing the creep vehicle speed from increasing more than necessary on a steep downhill or the creep vehicle speed from decreasing significantly on a steep uphill. And an appropriate creep vehicle speed V is obtained according to the gradient α.

The setting of the target creep vehicle speed V0 with respect to the steering wheel angle β is performed based on the characteristics shown in the map of FIG.

In this map, the larger the absolute value of the steering angle β in the plus side and the minus side, that is, in either direction, the larger the target creep vehicle speed V
0 is set to be low. Therefore, at the time of turning during creep running, the vehicle speed becomes lower as the vehicle turns sharply, and good turning traveling performance is obtained.

In this case, the characteristic of the target creep vehicle speed V0 with respect to the steering wheel angle β is represented by A shown by a chain line.
The creep vehicle speed V is set to be lower than the characteristics of the T-vehicle, whereby the load of the brake operation during turning is reduced, and good turning performance during creep running is realized.

When the target creep vehicle speed V0 is set as described above, in step S44, the target value of the toroidal ratio R for realizing the target creep vehicle speed V0 is set using a predetermined function. . In particular,
The target ratio R0 is determined from the target creep vehicle speed V0, the input rotation speed (engine rotation speed) Rev at that time, and vehicle specifications such as the gear ratio (final gear ratio) of the differential device 5 and the effective radius of the tire. Feedback control is performed in steps S6 to S9 so as to achieve the target ratio R0.

In the above-described control of the creep vehicle speed V, control of the engine torque by the above-described ISC valve 400 is performed as necessary in order to achieve the target creep vehicle speed V0.

As described above, the control up to the depression of the accelerator pedal is performed through the switching from the N range to the travel range.
Is performed when it is determined in step S2 that the line pressure PL is equal to or higher than the predetermined value PL1. On the contrary,
When PL <PL1, that is, when the required line pressure PL is not obtained due to high viscosity of the hydraulic oil immediately after the start of the engine or the like, in step S45, the pulse motors 251 and 251 for controlling the three-layer valve are set. The number of pulses N output to the motor 252 is set to 0, and the operation of these motors 251 and 252 is prohibited.

That is, when the line pressure PL is insufficient, the step motors 251 and 252 or the three-layer valves 220 and 23
When the feedback control of the toroidal ratio R by 0 is performed, a signal for causing the sleeves 222 and 232 of the three-layer valves 220 and 230 to stroke beyond the movable range is output in order to converge the actual ratio R to the target ratio R0. This may cause troubles such as damaging the sleeves 222, 232 and the like.

Accordingly, in such a case, the three-layer valve 22
The control of 0,230 is prohibited, and the toroidal ratio is fixed at the value at that time. In this case, the control of the duty ratio D is also stopped, and the engaged state of the low clutch 60 is fixed at the state at that time. This control suspension operation is performed not only immediately after the start of the engine but also in any stage until the accelerator pedal is depressed, when PL <PL1 for some reason.

In the above control, in the control of the creep force during the stoppage after the switch to the travel range and until the brake switch is turned off in steps S28 to S31 in FIG. 11, first, the target of the toroidal ratio R is set. The value R0 is set to a predetermined value Rc, and the engagement force of the low clutch 60 at which the target creep force is obtained under the toroidal ratio Rc is controlled. May be set to a value at which a required half-clutch state can be obtained, and the toroidal ratio R may be controlled so as to obtain a target creep force under the engagement force.

In this case, steps S28 to S31 in the flowchart are changed as shown in FIG.

That is, in step S24, the second timer T2
When the measured time t2 exceeds the third predetermined time Tc and it is determined in step S27 that the brake switch 310 is ON, first in step S28 ', the target value P0 of the low clutch pressure P is reduced by half. In step S29 ', a predetermined correlation function [D = f between the low clutch pressure P and the duty ratio D of the duty solenoid valve 271 is set.
(P)], the duty ratio D for obtaining the target low clutch pressure Pc is set.

Next, in step S30 ', the engine torque is estimated based on the throttle opening TVO and the like. In step S31', based on the estimated engine torque and the target low clutch pressure Pc, as shown in FIG. The target value R0 of the toroidal ratio R is determined from a map set in advance. Then, step S in the flowchart of FIG.
In steps 6 to S9, feedback control of the toroidal ratio R is performed so as to achieve the target ratio R0, and a signal of the duty ratio D is output to the duty solenoid valve 271 in step S10.

Thus, similar to the case shown in FIG.
The creep force during stop will be controlled properly,
If the low clutch pressure P is fixed and the toroidal ratio R is controlled as in the control in FIG. 20, the response of the creep force control will be improved.

In this case, especially when the torque fluctuation of the engine is large, if the target value P0 of the low clutch pressure P is set low according to the magnitude, for example, a value Pc 'shown in FIG. The burden on the clutch 60 is reduced,
The clutch 60 by absorbing large torque fluctuations
In addition to preventing the deterioration of the durability, the torque fluctuation is effectively absorbed by the clutch 60, and transmission of engine vibration to the vehicle body is suppressed.

The map in FIG. 21 used in the control for fixing the low clutch pressure P is the same as the map in FIG. 16 used in the control for fixing the toroidal ratio R, and is indicated by arrows Y and Y. Thus, the only difference is that the toroidal ratio P is determined based on the low clutch pressure P. Then, the correction of the target creep force based on the engine torque and the gradient α is similarly performed.

As described above, when the torque fluctuation is large, the low clutch pressure P is set to a value lower than the normal value P.
When it is set to c ', the target force of the toroidal ratio R is increased to obtain the same creep force as shown by arrows Y' and Y 'in order to increase the driving force generated from the continuously variable transmission mechanisms 20 and 30. The value R0 'is controlled to the GN ratio Rgn side.

[0168]

As described above, according to the present invention, in the power train of a vehicle using a continuously variable transmission of a toroidal type or the like, the speed of the continuously variable transmission can be changed when the traveling range is selected. While setting the gear ratio to a gear ratio different from the predetermined gear ratio that is in the geared neutral state,
Since the friction element is controlled to be in the semi-fastened state, this creep force is absorbed by the sliding of the friction element when the vehicle is stopped by depressing the brake pedal while generating the required creep force. become. Accordingly, a stable creep force can be obtained, and when the vehicle is to be stopped in a state in which the friction element is completely fastened against the creep force, excessive sliding in the friction element or the continuously variable transmission mechanism is required. Therefore, the durability of these members or the powertrain as a whole is improved.

In particular, according to the second invention, the above-described creep force control is performed more smoothly and stably.
According to the invention, the creep force control is performed with good responsiveness.

Further, according to the fourth aspect, the fastening force for bringing the friction element into the semi-fastened state is set to be lower as the variation of the input torque to the power train is larger, so that this torque variation can be effectively reduced. It can absorb and prevent the transmission of engine vibrations to the vehicle body, and reduce the durability of the friction element, such as acceleration of wear, caused by sliding according to torque fluctuation under relatively high fastening force. Will be prevented.

Further, according to the fifth invention, the creep force when the vehicle is stopped in the driving range is corrected according to the engine torque, and according to the sixth invention, the creep force is corrected according to the gradient of the road surface. As a result, in any case, an appropriate creep force can always be obtained regardless of the magnitude of the engine torque or the gradient of the road surface.

[Brief description of the drawings]

FIG. 1 is a skeleton diagram showing a mechanical configuration of a toroidal-type continuously variable transmission according to an embodiment of the present invention.

FIG. 2 is a plan view showing a specific structure of a main part of the transmission in an expanded state.

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

FIG. 4 is a circuit diagram of hydraulic control of the transmission.

FIG. 5 is a partial cross-sectional view of the periphery of a shift control three-layer valve as viewed from a direction B in FIG.

FIG. 6 is a partial cross-sectional view of the vicinity of a speed change control mechanism as viewed from a direction C in FIG. 3;

FIG. 7 is a characteristic diagram showing a relationship between the number of pulses of a step motor and a toroidal ratio.

FIG. 8 is a block diagram showing a control system of the entire power train.

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

FIG. 10 is a flowchart showing a first half of a control operation from a stop state in an N range to a start through switching to a travel range.

FIG. 11 is a flowchart showing a latter half of the control operation.

FIG. 12 is a time chart of the control operation.

FIG. 13 is a characteristic diagram showing a relationship between a toroidal ratio and a unit torque.

FIG. 14 is a map showing a relationship between a deviation for feedback control and the number of pulses of a step motor.

FIG. 15 is a schematic explanatory view of an ISC valve.

FIG. 16 is a map showing a relationship between a toroidal ratio and a low clutch pressure for obtaining a constant creep force.

FIG. 17 is a map showing a relationship between a duty ratio of a duty solenoid valve and a low clutch pressure.

FIG. 18 is a map showing a relationship between a gradient of a traveling road and a target creep vehicle speed.

FIG. 19 is a map showing a relationship between a steering wheel angle and a target creep vehicle speed.

FIG. 20 is a flowchart illustrating another control example of creep force control.

21 is a map showing a relationship between a toroidal ratio and a low clutch pressure for obtaining a constant creep force used in the control of FIG. 21.

[Explanation of symbols]

DESCRIPTION OF SYMBOLS 1 Engine 10 Powertrain 20, 30 Continuously variable transmission mechanism 50 Gear mechanism (planetary gear mechanism) 60, 70 Friction element (low clutch, high clutch) 200 Hydraulic control circuit 251, 252 Gear ratio control means (step motor) 271, 272 Friction element control means (duty solenoid valve) 300 Control means (control unit)

──────────────────────────────────────────────────続 き Continued on the front page (51) Int.Cl. ⁷ Identification code FI Theme coat ゛ (Reference) F16H 63:06

Claims (6)

[Claims] 1. A vehicle has a first path passing through a continuously variable transmission mechanism and a gear mechanism, and a second path passing only through the continuously variable transmission mechanism. A configuration in which a neutral state can be realized by controlling the speed ratio of the continuously variable transmission mechanism to a predetermined speed ratio in a state where the first path is in a power transmission state by the friction element. A power train control device, wherein the range detection means for detecting selection of a range set for the vehicle, the stop detection means for detecting a stop state of the vehicle, and the speed ratio of the continuously variable transmission mechanism. A change ratio changing means for changing, and a fastening force adjusting means for adjusting the fastening force of the friction element are provided, and the selection of the traveling range is detected by the range detecting means,
When the stop state of the vehicle is detected by the stop detecting means, the speed ratio of the continuously variable transmission mechanism is set to a speed ratio different from the predetermined speed ratio, and the friction element is controlled to a half-engaged state. A control device for a power train, further comprising control means for controlling the speed ratio variable means and the engagement force adjusting means so as to generate the creep force. 2. The control means controls the variable gear ratio means when the selection of the traveling range is detected by the range detection means and the stopped state of the vehicle is detected by the stop detection means, and the stepless speed change means controls the continuously variable transmission. The speed ratio of the mechanism is set to a second predetermined speed ratio different from the predetermined speed ratio at which the neutral state is obtained, and the engagement force is such that the creep force becomes substantially constant under the second predetermined speed ratio. The control apparatus for a power train according to claim 1, wherein the control means controls a fastening force adjusting means of the friction element. 3. The control means controls the engagement force of the friction element by the engagement force adjusting means when the selection of the travel range is detected by the range detection means and the stop state of the vehicle is detected by the stop detection means. The gear ratio of the continuously variable transmission mechanism is set by the gear ratio variable means so that the gear ratio is set to a predetermined engagement force in a semi-engaged state and the gear ratio becomes substantially constant under the predetermined engagement force. The control device for a power train according to claim 1, wherein: 4. A vehicle according to claim 1, further comprising: a torque fluctuation detecting means for detecting a fluctuation of an input torque to the power train; a control means detecting a selection of a traveling range by a range detecting means; When the stopped state of the vehicle is detected, the larger the torque fluctuation detected by the torque fluctuation detecting means, the lower the predetermined engagement force for setting the friction element to the semi-engaged state, and
4. The power train according to claim 3, wherein the speed ratio of the continuously variable transmission mechanism is controlled by the speed ratio variable means so that the speed ratio at which the creep force becomes substantially constant under the predetermined fastening force. Control device. 5. An engine torque detecting means for detecting an output torque of the engine is provided, and the control means detects the selection of the traveling range by the range detecting means and detects the stop state of the vehicle by the stop detecting means. When it is detected, at least one of the speed ratio of the continuously variable transmission mechanism and the engaging force of the friction element is corrected so that the creep force is corrected according to the engine torque detected by the engine torque detecting means. The power train control device according to claim 1, wherein: 6. A vehicle comprising: a gradient detecting means for detecting a gradient of a traveling road; a control means detecting a selection of a traveling range by a range detecting means and detecting a stop state of the vehicle by a stop detecting means; Is corrected, at least one of the speed ratio of the continuously variable transmission mechanism and the engaging force of the friction element is corrected so that the creep force is corrected according to the gradient of the traveling path detected by the gradient detecting means. The power train control device according to claim 1, wherein: