JP2003056685A - Variable speed control device for transmission with toroidal continuously variable speed mechanism - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a variable speed control device for a transmission with a toroidal continuously variable speed mechanism for achieving shift ratio control in high follow-up response to a target shift ratio with quick shifting operation in such a condition that sudden shifting operation is required at sudden deceleration or acceleration by using feed forward compensation for an actuator command value without causing hunting in normal control or involving the setting of a difficult target shift ratio. SOLUTION: The variable speed control device for the transmission with the toroidal continuously variable speed mechanism comprises actual tilting angle calculating means for calculating the actual tilting angle of a trunnion, target tilting angle calculating means for calculating the target tilting angle of the trunnion from a target shift ratio, target tilting speed calculating means for calculating the target tilting speed of the toroidal variable speed mechanism from the calculated actual tilting angle and target tilting angle, power roller displacement calculating means for calculating the displacement of a power roller in the axial direction of tilting the trunnion corresponding to the calculated target tilting angle, and actuator command value correcting means for correcting the actuator command value according to the calculated displacement of the power roller.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a toroidal type continuously variable transmission mechanism that operates a gear shift actuator to increase the gear shift speed so that the actual gear ratio can follow a target value when the vehicle acceleration / deceleration is large. Belonging to the technical field of the shift control device of the transmission that has.

[0002]

2. Description of the Related Art In a transmission having a toroidal type continuously variable transmission mechanism, gear ratio control for changing a contact radius ratio between an input / output disk and a power roller is performed by changing a trunnion of the toroidal type continuously variable transmission mechanism in a direction of a tilt axis. And the power roller is tilted.

That is, a drive command based on the number of steps corresponding to a target gear ratio is given to a step motor, which is a gear shift actuator, and a gear shift spool of a shift control valve is displaced through a gear shift link. Due to the displacement of the speed change spool, the hydraulic oil is guided from the shift control valve to one servo piston chamber of the servo piston provided on the trunnion shaft, and the hydraulic oil is discharged from the other servo piston chamber, so that the trunnion is tilted. It is displaced along the axis of rotation. Due to the displacement of the trunnion in the tilt axis direction, the center of rotation of the power roller is offset with respect to the disk rotation center position, and the offset causes the power roller to tilt due to the side slip force generated at the contact portion between the power roller and the input / output disk. Roll over. This tilting motion and offset are transmitted to the speed change link via the precess cam and lever, and the feedback displacement causes the displacement of the speed change spool given by the step motor to be returned. The displacement is returned to the original disk rotation center position, and the power roller stops the tilting operation.

At the time of this gear shift, Japanese Patent Laid-Open No. 2000-23467
As described in Japanese Patent Publication No. 0, vehicle speed (output rotation speed)
And the target input speed from the throttle opening TVO,
Calculate the target gear ratio from the target input speed and output speed,
The number of steps corresponding to the target speed ratio is given to the step motor, which is a speed change actuator, by feedforward control (F / F control).

[0005]

However, in the transmission control device of the transmission having the conventional toroidal type continuously variable transmission mechanism, the step number-gear ratio characteristic with respect to the output torque measured in the fixed gear ratio state is obtained. Based on the gear ratio F
In the gear ratio control that performs the / F control, in the case of sudden braking, the gear shift range from the gear ratio before the gear shift to the gear ratio after the gear shift is large, that is, the tilt angle of the power roller changes greatly. As a result, there is a problem that the position of the shift spool of the shift control valve is moved in the direction of decreasing the shift speed due to mechanical feedback from the recess cam, and the shift is delayed.

As a transmission having a toroidal type continuously variable transmission mechanism, an infinitely variable transmission continuously variable transmission (hereinafter referred to as IVT) which achieves a gear ratio = ∞ (neutral) in a state where a power path is mechanically coupled. ) Will be described as an example.

In the power circulation mode of the IVT, a toroidal type continuously variable transmission with a torque converter (hereinafter, CVT) is used.
Similarly, compared to the direct connection mode of the IVT, the amount of change in the gear ratio necessary for the change in vehicle speed is large. In addition, among them, in the accelerator low opening degree and the accelerator fully closed state, the input rotation speed is controlled low, and in this state,
At the time of sudden braking from the rotation synchronization point (hereinafter, RSP), it is necessary to change the gear ratio of the toroidal continuously variable transmission mechanism from near LOW to just before HIGH at a speed of 20 km / h or less until it stops. No (see Figure 4). In such a situation, if the shift of the toroidal type continuously variable transmission is delayed as described above, the IVT gear ratio does not return to the geared neutral point (GNP) when the vehicle is stopped, and the creep torque exceeds the intended creep torque. Torque is generated,
An engine stall may occur in an IVT in which the power paths are mechanically connected. Also, the higher the speed change speed, the greater the displacement displacement of the power roller. However, especially in a situation where the input rotation is low, a larger displacement displacement of the power roller is required to obtain a large speed change speed. A large deviation occurs in the map of the number of steps-gear ratio (tilt angle) measured in the state.

Of course, considering the acceleration state, the power circulation mode provides a larger driving force than the direct connection mode, so that the vehicle speed changes greatly during acceleration, and therefore the shift speed that must be generated during acceleration is also CVT. In addition, when the accelerator pedal is started, the control must be performed while changing the gear ratio of the toroidal type continuously variable transmission even in the highest acceleration region. The ratio does not move forward despite the vehicle speed, and the target I
In some cases, the engine speed becomes LOW (higher in the toroidal type continuously variable transmission) than the VT gear ratio, becomes higher than the target engine speed, and a feeling of engine idling may occur. By the way, in the case of CVT, when the accelerator is fully opened, the gear ratio is fixed to LOW and is accelerated by the torque increasing action of the torque converter, and the gear ratio is fixed in a large driving force range of 40 km / h or less. .

There is a method of suppressing any of the problems by feedback control (F / B control) of the gear ratio of the toroidal type continuously variable transmission mechanism or IVT input speed, but in this case, the F / B gain is increased. There is no choice but to set, hunting will occur in normal control, and it is difficult to obtain a satisfactory solution.

Therefore, in order to solve such a problem,
Japanese Unexamined Patent Publication No. 11-13876 proposes a technique in which the target gear ratio of the toroidal type continuously variable transmission mechanism is set at a position exceeding an original target value at the time of sudden deceleration, and as a result, the gear shifting speed is increased. However, in this conventional technique, when it is determined that the speed is in rapid deceleration, the gear ratio is uniformly set to be larger than the minimum gear ratio, so that a new target gear ratio on a lower speed side (publication) is disclosed. In setting iG) described in 1), not only many experiments considering the characteristics of the vehicle must be repeated, but also depending on the set value of the speed change ratio, an appropriate speed change corresponding to the deceleration of the vehicle cannot be obtained. Sometimes.

The present invention has been made in view of the above problems, and an object thereof is to feed an actuator command value without causing hunting in normal control and difficult target gear ratio setting. With forward compensation,
In a situation where a sudden gear change operation is required, such as during sudden deceleration or sudden acceleration, a gear shift with a toroidal type continuously variable transmission mechanism that can achieve a gear ratio control with a high follow-up response to a target gear ratio with a fast gear shift speed To provide a shift control device for a machine.

[0012]

In order to achieve the above object, in the invention according to claim 1, an input disk and an output disk which are coaxially opposed to each other, and a power roller sandwiched between the input and output disks. And a trunnion that rotatably supports the power roller and that has a tilt axis of the power roller associated with a shift, a servo piston that strokes the trunnion in the tilt axis direction, and a servo pressure to the servo piston. A toroidal type continuously variable transmission having a speed change actuator; and a target speed ratio calculating means for calculating a target speed ratio of the toroidal type continuously variable transmission mechanism on the basis of an operating state of the vehicle. Of a transmission having a toroidal-type continuously variable transmission mechanism that performs shift control by sending the actuator command value to the shift actuator. In the control device, the actual tilt angle calculating means for calculating the actual tilt angle of the trunnion, the target tilt angle calculating means for calculating the target tilt angle of the trunnion from the target gear ratio, and the calculated actual tilt Target tilt speed calculation means for calculating a target tilt speed of the toroidal type continuously variable transmission mechanism from a tilt angle and a target tilt angle, and a trunnion tilt axis direction of the power roller according to the calculated target tilt angle. And a power roller displacement calculating means for calculating the displacement of the actuator and an actuator command value correcting means for correcting the actuator command value according to the calculated power roller displacement.

According to a second aspect of the present invention, in the shift control device for a transmission having the toroidal type continuously variable transmission mechanism according to the first aspect, the actuator command value correction means is
It is characterized in that it is a means for performing a larger correction as the input speed or output speed of the toroidal type continuously variable transmission mechanism or the sum of the input speed and the output speed becomes smaller.

According to a third aspect of the present invention, in the gear shift control device for a transmission having the toroidal type continuously variable transmission mechanism according to the first or second aspect, the power roller displacement calculating means has a predetermined tilt angle. At the first shift speed obtained from the change in the tilt angle between the input disc and the power roller, or at the second shift speed obtained from the change in the tilt angle between the output disc and the power roller at a predetermined tilt angle, or A means for calculating the power roller displacement by dividing the average shift speed by the input speed of the toroidal type continuously variable transmission mechanism.

According to a fourth aspect of the present invention, in the transmission control device for a transmission having the toroidal type continuously variable transmission mechanism according to the first or second aspect, the power roller displacement calculating means is provided with a predetermined tilt angle. At the first shift speed obtained from the change in the tilt angle between the input disc and the power roller, or at the second shift speed obtained from the change in the tilt angle between the output disc and the power roller at a predetermined tilt angle, or A means for calculating the power roller displacement by dividing the average shift speed by the output rotation speed of the toroidal type continuously variable transmission mechanism.

According to a fifth aspect of the present invention, in the gear shift control device for a transmission having the toroidal type continuously variable transmission mechanism according to the first or second aspect, the power roller displacement calculating means has a predetermined tilt angle. At the first shift speed obtained from the change in the tilt angle between the input disc and the power roller, or at the second shift speed obtained from the change in the tilt angle between the output disc and the power roller at a predetermined tilt angle, or The power roller displacement is calculated by dividing the average shift speed by the sum of the input rotation speed and the output rotation speed of the toroidal continuously variable transmission mechanism.

According to a sixth aspect of the invention, in the shift control device for a transmission having the toroidal type continuously variable transmission mechanism according to the third, fourth or fifth aspect, the power roller displacement calculating means is a predetermined one. Of the tilt angle between the input disc and the power roller at a tilt angle of 1 or a second shift speed determined from the tilt angle change between the output disc and the power roller at a predetermined tilt angle. Shift speed, or the average shift speed of these,
The input roller speed or output wheel speed of the toroidal type continuously variable transmission mechanism, or a value calculated by dividing by the sum of the input wheel speed and the output wheel speed is multiplied by a variable according to the tilt angle, and the power roller It is a means for calculating displacement.

According to a seventh aspect of the present invention, in the gear shift control device of the transmission having the toroidal type continuously variable transmission mechanism according to any one of the first to sixth aspects, the actuator command value correcting means includes 1 shift speed, or
If the second speed change speed or the average speed change speed thereof is equal to or lower than a predetermined speed, the means does not perform correction.

According to an eighth aspect of the present invention, in the transmission control device for a transmission having the toroidal type continuously variable transmission mechanism according to the seventh aspect, the actuator command value correction means is:
Input speed of the toroidal type continuously variable transmission, or
It is characterized in that the higher the output speed is, the wider the area where the correction is not applied is set.

According to a ninth aspect of the present invention, in the shift control device of the transmission having the toroidal type continuously variable transmission mechanism according to the first aspect, the transmission having the toroidal type continuously variable transmission mechanism has a unit input shaft. A toroidal type continuously variable transmission and a constant transmission mechanism, respectively, a sun gear connected to the output shaft of the toroidal type continuously variable transmission, a carrier connected to the output shaft of the constant transmission mechanism, and a ring gear connected to the unit output shaft. Consisting of a planetary gear mechanism, a power circulation mode clutch provided between the output of the constant speed change mechanism and the carrier, and a direct connection mode provided between the output of the toroidal type continuously variable transmission and the unit output shaft. A continuously variable transmission having an infinite transmission ratio having a clutch, and the actuator command value correction means is a direct connection mode when the power circulation mode clutch is engaged. Means for correcting the gear ratio of the infinitely variable transmission continuously variable transmission by issuing a command to the speed change actuator of the toroidal type continuously variable transmission in the power circulation mode with the clutch released and only when the brake is operated. Is characterized in that.

According to a tenth aspect of the present invention, in the transmission control device for a transmission having the toroidal type continuously variable transmission mechanism according to the ninth aspect, the target tilt speed calculation means is an infinitely variable transmission continuously variable transmission. Is a means for calculating the target tilting speed using the deceleration of the output shaft rotation.

[0022]

In the invention according to claim 1, the target gear ratio calculating means calculates and calculates the target gear ratio of the toroidal type continuously variable transmission mechanism based on the operating state of the vehicle during traveling. The gear shift control is performed by sending an actuator command value corresponding to the target gear ratio to the gear shift actuator of the toroidal type continuously variable transmission mechanism. On the other hand, the actual tilt angle calculation means calculates the actual tilt angle of the trunnion, the target tilt angle calculation means calculates the target tilt angle of the trunnion from the target gear ratio, and the target tilt speed calculation means: The target tilt speed of the toroidal type continuously variable transmission is calculated from the calculated actual tilt angle and the target tilt angle, and the power roller displacement calculation means calculates the trunnion tilt of the power roller according to the calculated target tilt angle. The displacement in the rolling axis direction is calculated, and the actuator command value correction means corrects the actuator command value according to the calculated power roller displacement. In other words, the required speed change speed of the toroidal type continuously variable transmission mechanism is estimated from the tilt speed of the trunnion that has a corresponding relationship with the speed change speed, and the tilt speed is increased in accordance with the target tilt speed. The actuator command value is corrected. Here, the tilting speed of the trunnion increases as the offset amount of the rotation center of the power roller with respect to the disk rotation center position increases. Therefore, in a situation where a sudden gear shift operation is required such as during sudden deceleration or sudden acceleration, the tilting speed (= gear shifting speed) is increased by increasing the displacement amount of the power roller in the trunnion tilting axis direction. . Therefore, hunting does not occur in normal control as in the case of increasing the feedback control gain, and it is difficult to set the target gear ratio as when setting the target gear ratio itself beyond the original target value. The feed-forward compensation of the actuator command value that does not cause the speed change ratio control with high follow-up responsiveness to the target speed ratio with a fast speed change in situations where a sudden speed change operation such as during sudden deceleration or sudden acceleration is required. Can be achieved.

According to the second aspect of the present invention, in the actuator command value correcting means, the input speed or output speed of the toroidal continuously variable transmission mechanism, or the sum of these input speed and output speed. The smaller the value is, the larger the correction is made.Therefore, an appropriate amount of correction can be obtained according to the input rotation speed of the toroidal type continuously variable transmission mechanism. Hunting that occurs too much can be effectively suppressed. That is, for example, when stopping from a low speed state by sudden braking, or when accelerating to start by depressing the accelerator from a stop, there is a demand for higher follow-up response to the target gear ratio than during deceleration or acceleration in the high speed range. is there.

In the invention according to claim 3, in the power roller displacement calculating means, the first shift speed obtained from the change of the tilt angle between the input disk and the power roller at a predetermined tilt angle, or the predetermined shift speed. Displacement of the power roller by dividing the second speed change speed obtained from the change in the tilt angle between the output disc and the power roller at the tilt angle or the average speed change speed thereof by the input speed of the toroidal type continuously variable transmission mechanism. Therefore, the power roller displacement can be easily calculated, and the actuator command value is corrected according to the value of the power roller displacement to maintain the response to the target gear ratio even during a sudden gear shift. can do.

In the invention according to claim 4, in the power roller displacement calculating means, the first shift speed obtained from the change of the tilt angle between the input disc and the power roller at a predetermined tilt angle, or the predetermined shift speed. The power roller displacement is obtained by dividing the second speed change speed obtained from the change in the tilt angle between the output disc and the power roller at the tilt angle or the average speed change speed thereof by the output speed of the toroidal type continuously variable transmission mechanism. Therefore, the power roller displacement can be easily calculated, and the actuator command value is corrected according to the value of the power roller displacement to maintain the response to the target gear ratio even during a sudden gear shift. can do.

In the invention according to claim 5, in the power roller displacement calculating means, the first shift speed obtained from the change of the tilt angle between the input disk and the power roller at a predetermined tilt angle, or the predetermined shift speed. The second speed change speed obtained from the change in the tilt angle between the output disc and the power roller at the tilt angle or the average speed thereof is divided by the sum of the input speed and the output speed of the toroidal continuously variable transmission mechanism. Since the power roller displacement is calculated by doing so, the power roller displacement can be easily calculated, and by correcting the actuator command value according to the value of the power roller displacement, the target gear ratio can be achieved even during a sudden gear shift. It is possible to maintain the following responsiveness of.

According to the invention of claim 6, in the power roller displacement calculating means, the shift speed is the input rotation speed of the toroidal type continuously variable transmission mechanism, or the shift speed is the output rotation speed of the toroidal type continuously variable transmission mechanism. Alternatively, the power roller displacement is calculated by multiplying a value calculated by dividing the speed change speed by the sum of the input speed and the output speed of the toroidal continuously variable transmission mechanism by a variable according to the tilt angle. Therefore, compared with the invention according to claims 3, 4, and 5, the power roller displacement can be calculated more accurately, and the actuator command value is corrected according to the value of the power roller displacement, so that the target gear ratio can be achieved even during a sudden gear shift. The tracking response can be maintained.

In the invention according to claim 7, in the actuator command value correction means, if the first speed change speed, the second speed change speed, or the average speed change speed thereof is equal to or lower than a predetermined speed, the correction is performed. Since it is not added, the region where the speed of change of the gear ratio is not so much required is the control that emphasizes gear shift stability, and it is possible to prevent this correction control from acting sensitively due to rotational speed measurement accuracy and control delay. In a situation where a large speed change speed is required, the response of the speed change control can be surely improved.

According to the eighth aspect of the invention, in the actuator command value correcting means, the higher the input rotational speed or the output rotational speed of the toroidal continuously variable transmission mechanism, the wider the region where correction is not performed is set. Therefore, it is possible to effectively prevent the control overshoot that tends to occur in the high rotation speed range, and obtain the necessary control response in the low rotation speed range.

In the invention according to claim 9, in the actuator command value correcting means, the power circulation mode clutch is engaged and the direct coupling mode clutch is released in the power circulation mode, and only when the brake is operated, the toroidal type clutch is not used. The command to the gear shift actuator of the stepped speed change mechanism corrects the gear ratio of the infinitely variable continuously variable transmission to the neutral side, so that this control logic can be operated only in the situation where this control is most necessary. In addition to preventing engine stall due to sudden braking, it is possible to reduce the addition of a shift control controller in other operating areas, and it is possible to reduce the number of man-hours required for logic matching and confirmation. Further, in other driving regions, the shift control stability is not affected, so that it is possible to make a stronger correction, and it is possible to further improve the engine stall resistance during sudden braking.

In the tenth aspect of the present invention, the target tilting speed calculating means calculates the target tilting speed by using the deceleration of the output shaft rotation of the continuously variable transmission having an infinite gear ratio.
That is, in the situation where the brake that decelerates the output shaft rotation is stepped on, it is necessary to control the input shaft rotation of the continuously variable transmission with an infinite transmission ratio to be substantially constant. Therefore, by using the deceleration of the output shaft rotation of the continuously variable transmission with an infinite transmission ratio for the calculation of the target tilting speed, it is possible to perform control in accordance with the required shift speed with a very simple logic. Further, since the control can be performed based on the signal of the rotation sensor with high accuracy such as ABS, it is possible to secure a considerable control accuracy without increasing the cost.

[0032]

BEST MODE FOR CARRYING OUT THE INVENTION Hereinafter, an embodiment for realizing a shift control device for a transmission having a toroidal type continuously variable transmission mechanism according to the present invention will be described as a first embodiment corresponding to claims 1 to 6, and A description will be given based on a second embodiment corresponding to claims 7 and 8, and a third embodiment corresponding to claims 9 and 10.

(First Embodiment) First, the structure will be described. FIG. 1 is an overall configuration diagram showing an infinitely variable transmission continuously variable transmission (an example of a transmission having a toroidal type continuously variable transmission mechanism) according to the first embodiment, in which an IVT input shaft 1 connected to an engine and drive wheels are connected. A double-cavity toroidal continuously variable transmission 4 (CVU) composed of a half toroidal, a constant transmission 5, and a single planetary planetary gear mechanism 6 are arranged between the IVT output shaft 2 to be connected. Has been done.
A differential 3 is interposed between the IVT output shaft 2 and the drive wheels.

The toroidal type continuously variable transmission mechanism 4 includes two sets of input disks 41, 41 connected to the input shaft 1 and two sets of output disks 4 connected to the output chain mechanism 42.
3, 43 and two power rollers 44, 44 sandwiched between the input / output disks 41, 43. The power rollers 44, 44 are supported by trunnions 45, 45 (see FIG. 2), and the trunnions 45, 45 are tilted to the power rollers 44, 44 by slightly displacing the trunnions 45, 45 in the tilt axis direction. A CVU gear ratio determined by the tilt angles of the power rollers 44, 44 is changed steplessly by applying a force and tilting until the tilt angle corresponds to the control command.

The constant speed change mechanism 5 meshes with the input gear 51 fixed to the IVT input shaft 1 and the input gear 51.
The output gear 52 is rotatably supported on the T output shaft 2 and includes a reduction gear mechanism.

The planetary gear mechanism 6 comprises a sun gear 61 rotatably supported on the IVT output shaft 2, a pinion carrier 62 supporting a pinion, and a ring gear 63 fixed to the IVT output shaft 2. There is.

The pinion carrier 6 of the planetary gear mechanism 6
2 and the output gear 52 of the constant speed change mechanism 5, a power circulation mode clutch 8 is arranged, and a direct connection mode clutch 9 is provided between the output chain mechanism 42 of the toroidal continuously variable speed change mechanism 4 and the IVT output shaft 2. It is arranged. The engagement force of the direct coupling mode clutch 9 is applied to the first hydraulic servo 91 by the first
The hydraulic pressure supplied from the solenoid valve 92 determines the engagement force of the power circulation mode clutch 8 based on the hydraulic pressure supplied from the second solenoid valve 94 to the second hydraulic servo 93.

The CVU input rotation sensor 10 is arranged at a position close to the teeth of the input gear 51 of the constant speed change mechanism 5, and the CVU input position at the position close to the output chain mechanism 42 of the toroidal type continuously variable speed change mechanism 4. The output rotation sensor 11 is arranged, and the output rotation sensor 12 (vehicle speed sensor) is arranged at a position close to the tooth portion of the sensor gear 13 provided on the IVT output shaft 2.

FIG. 2 is a mechanical configuration diagram of a hydraulic system for managing CVU shifting. Power roller 44 is trunnion 45
Supported from the back. Gear shifting in the toroidal type continuously variable transmission mechanism 4 is performed by displacing the trunnion 45 vertically from an equilibrium point, and this displacement causes the power roller 4 to move.
4 and the input / output disks 41 and 43 have different rotational direction vectors, and the power roller 44 tilts.

The trunnion 45 is connected to the servo piston 31 of the hydraulic servo 30, and the hydraulic servo 30
Oil in the Hi side cylinder 30a and the Low side cylinder 30
It is displaced by the differential pressure of the oil in b. The hydraulic pressure of the Hi side cylinder 30a and the hydraulic pressure of the Low side cylinder 30b are controlled by the shift control valve 46.

The shift control valve 46 allows the oil supplied from the line pressure port 46L to flow through the Hi side port 4 when the speed change spool 46S in the valve is displaced.
The pressure difference in the hydraulic servo 30 is changed by causing the oil to flow to one of the 6Hi or Low side port 46Low and causing the oil to flow from the other Low side port 46Low or Hi side port 46Hi to the drain port 46L.

A precess cam 35 is attached to one of the trunnions 45, and the precess cam 3 is provided.
No. 5 has a groove. The groove of the recess cam 35 is in contact with one end of the L-shaped link 38, and one end of the L-shaped link 38 is freely supported by one end of the speed change link 37. Therefore, the displacement and tilt angle of the trunnion 45 are fed back to the speed change link 37. The other end of the speed change link 37 is connected to a step motor 36 (speed change actuator), and the speed change spool 46S of the shift control valve 46 is freely supported on the speed change link 37.
Therefore, the displacement amount of the step motor 36 and the recess cam 3
The displacement of the speed change spool 46S is determined by the feedback amount of 5.

FIG. 3 shows an IVT electronic control system. The IVT controller 20 outputs a control command based on input information and the trunnion 4 supporting the power roller 44.
5, a step motor 36 as a gear shift control actuator, a low clutch solenoid 21 which is a control actuator for engaging and releasing the power circulation mode clutch 8, and a high actuator which is a control actuator for engaging and releasing the direct coupling mode clutch 9. Clutch solenoid 22
Is equipped with.

The IVT controller 20 includes a range signal from an inhibitor switch (not shown) and the CVU.
CVU input rotation signal from the input rotation sensor 10, CVU output rotation signal from the CVU output rotation sensor 11, and output shaft rotation (vehicle speed) from the output rotation sensor 12.
Signal, throttle opening signal from a throttle opening sensor not shown, and idle S from an idle switch not shown
W signal and brake SW from the brake switch (not shown)
The signal, the oil temperature sensor signal from the oil temperature sensor 24, other signals, and the information on the target idle speed and the idle control flag from the engine controller 25 are input.

In the IVT controller 20, the power circulation mode in which the power circulation mode clutch 8 is disengaged and the direct connection mode clutch 9 is disengaged to transmit power including GNP having an infinite IVT gear ratio, and the direct connection mode clutch 9 is engaged. And a direct connection mode in which power is transmitted according to the output of the toroidal type continuously variable transmission mechanism 4 by releasing the power circulation mode clutch 8, a rotation synchronization point (RS
P; also referred to as a mode switching point) and the like, clutch switching control is performed, and based on the calculation processing result, a control command is issued to the low clutch solenoid 21 of the second solenoid valve 94 and the high clutch solenoid 22 of the first solenoid valve 92. Is output. Further, the target IVT speed ratio is determined by the target IVT input speed and the detected output speed (actual IVT output speed), and the speed ratio of the toroidal type continuously variable transmission mechanism 4 is controlled so as to realize the target IVT speed ratio. The step motor 36 is controlled based on the calculation processing result.
A control command is output to.

Next, the IVT speed ratio characteristic shown in FIG. 4 (the vertical axis represents IVT speed ratio = output shaft rotational speed / input shaft rotational speed and the horizontal axis represents CVU gear ratio) will be described. In the power circulation mode, the pinion carrier 6 of the planetary gear mechanism 6
2 rotates at a constant rotation speed in which the IVT input rotation speed is reduced by the constant transmission mechanism 5 by engaging the power circulation mode clutch 8, and the sun gear 61 changes its rotation speed according to the change of the CVU gear ratio, and the rotation of the ring gear 63. Is the output shaft rotation of an infinitely variable transmission. Therefore, in the power circulation mode, as the CVU gear ratio changes from LOW to HIGH, the output shaft rotation of the IVT becomes forward → neutral → reverse, and as the CVU gear ratio changes from HIGH to LOW, The rotation of the output shaft of the IVT is backward → neutral → forward. Further, in the direct coupling mode, the pinion carrier 62 of the planetary gear mechanism 6 can freely rotate by releasing the power circulation mode clutch 8, and the ring gear 63 is coupled with the sun gear 61 by fastening the direct coupling mode clutch 9 to have the same rotation speed. To rotate. Therefore, in the direct connection mode, as the CVU gear ratio changes from LOW to HIGH, the IVT gear ratio (1 / IVT speed ratio) changes from forward LOW to HIG.
It becomes H. From the above, the relationship between the IVT speed ratio and the CVU gear ratio is as shown in FIG.

Next, the gear ratio control of the toroidal type continuously variable transmission 4 (CVU), which is a component of the IVT, will be described. The CVU gear ratio gear ratio control is performed by feedback control (F / B control). However, in order to secure the stability and responsiveness of the control, it is desirable that the physically determined shift control actuator command value is calculated by the feedforward (F / F), and the work of the feedback control is made as small as possible. This makes it possible to secure the required responsiveness and stability even with a low F / B control gain. Hereinafter, in the present embodiment, a method of calculating the step motor command value of the directional flow rate control method in the toroidal type continuously variable transmission mechanism 4 using the recess cam 35, the speed change links 37, 38 and the step motor 36 will be described.

FIG. 5 shows a toroidal type continuously variable transmission 4 (CV
FIG. 6 is a diagram showing a control block of the gear ratio control of FIG.
In the figure, 50 is a basic step number Sb, a CVU input torque correction step number Sts, and a shift speed correction step number S, which will be described later.
An FF step number calculation block for calculating the FF step number Sff by the sum with v, and an FB step number calculation block for calculating the FB step number Sfb based on the target input rotation deviation. Then, the FF control amount calculation block 50
Has the following three control blocks. The first is the basic step number calculation block 51, which is the target CV.
The basic step number Sb during steady no-load due to the U gear ratio ict
Is a block for calculating. The second is the CVU input torque correction step number calculation block 52, which is necessary for correcting a steady shift of the gear ratio (referred to as torque shift) when the input torque acts on the CVU 4, Sts. Is a block for calculating. Three
The third is a shift speed correction step number calculation block 53, which is a characteristic configuration of the present invention, and a block for calculating a shift speed correction step number Sv for compensating for a transient change in the gear ratio change rate (shift speed). Is.

Next, the operation will be described.

[Gear Ratio Control Processing] FIG. 6 shows I of the first embodiment.
Each step will be described below with a flowchart showing the flow of the gear ratio control process performed by the VT controller 20. This flow has a fixed control cycle (for example, 10
msec), the number of pulses required to operate the number of steps operated in this control cycle is calculated, and the step motor 36 is driven.

First, FIG. 6a is a diagram showing the overall flow of the gear ratio control. First, in step (S1), each state quantity is detected. Next, in step (S2),
The FF step number Sff for the F / F control of the step number is calculated (FIG. 6b). Next, in step (S3),
The FB step number Sfb for the F / B control of the step number is calculated (FIG. 6c). Next, in step (S4),
FF calculated in step (S1) and step (S2)
The step number command value Sd is calculated by adding the step number Sff and the FB step number Sfb. Then, in step (S5), the step number command value Sd obtained in step (S4) is commanded to the step motor 36, and the step motor 36 is driven.

Next, FIG. 6b is a diagram showing a flow of calculating the FF step number Sff performed in step (S2) of the overall flow chart 6a. First, in step (S21),
Based on the target CVU gear ratio ict, the steady basic no-load step number Sb is calculated (FIG. 6d). Next, step (S
22), when the input torque acts on the CVU 4, the CVU required to correct the steady shift of the gear ratio.
The input torque correction step number Sts is calculated (Fig. 6).
e). Next, in step (S23), the shift speed correction step number Sv for compensating for the transitional speed ratio change rate (shift speed) difference is calculated (FIG. 6f). Then, in step (S24), the FF step number Sff for F / F control is obtained by the equation Sff = Sb + Sts + Sv.

Next, FIG. 6d is a flow chart for calculating the number of FF steps. The number of basic steps performed in step (S21) of FIG. 6b.
It is a figure which shows the calculation flow of Sb. First, step (S2
In 11), the arrival target input rotation Nio for the vehicle speed Vsp of FIG. 7 is map-searched for the arrival target input rotation Nio from the throttle opening TVO and the vehicle speed Vsp. Next, in step (S212), the reaching target input rotation Nio is filtered by the time constant to generate the target input rotation Nit. Next, in step (S213), the target IVT gear ratio et = Nit / No (No: IVT output rotation) is calculated. Next, at step (S214), the target CVU gear ratio ict is changed from the target IVT gear ratio et to IV of FIG.
The T gear ratio-CVU gear ratio map is searched (target gear ratio calculating means). This is the gear ratio of each gear, CVU4
It is theoretically calculated from the gear ratio, and it is possible to obtain it by calculation instead of map search.
Then, in step (S214), the basic step number Sb for the target CVU gear ratio ict is retrieved from the map of FIG. This map is created from the characteristics of the CVU gear ratio with respect to the number of steps under no load.

Next, FIG. 6e is a diagram showing a calculation flow of the CVU input torque correction step number Sts performed in the step (S22) of the FF step number calculation flow. First, in step (S221), the throttle opening TV
O is filtered to find TVOt. This is performed by using a time constant to perform a filtering process in order to correct the fact that the actual engine torque is delayed with respect to the change in the throttle opening. Next, in step (S222),
Throttle opening filtered value TVOt and IVT input rotation Ni
From (= engine rotation), the engine torque Te is map-searched from the engine torque map shown in FIG. Next, in step (S233), the IVT gear ratio e is calculated by e = No / Ni. Ni is the IVT input speed and No is the IVT output speed. Next, step (S
224), engine torque Te and IVT gear ratio e
To CVU input torque Tci are retrieved using the map of FIG. Since Tci≈Te in the direct connection mode, Tci = Te in the region where the gear ratio e is in the direct connection mode. Of course, the map may be searched only in the power circulation mode by determining the mode. In the power circulation mode, the CVU input torque Tci is the IVT speed ratio (IVT gear ratio)
Depends on This can be calculated from the gear ratio of each part gear, the CVU gear ratio, and the transmission efficiency of each part. Since these calculations are too heavy for the controller to perform, they are searched by the map of FIG. 11 prepared in advance. Next, in step (S225), CV
From the U gear ratio ic and the CVU input torque Tci, the CVU input torque correction step number Sts is searched using the map of FIG.

Next, FIG. 6f is a diagram showing a calculation flow of the shift speed correction step number Sv performed in the step (S23) of the FF step number calculation flow in FIG. 6b. First, in step (S231), the current tilt angle φ is searched from the tilt angle map for the CVU gear ratio ic in FIG. 13 (tilt angle calculation means). Next, in step (S232), the target tilt angle φt with respect to the target CVU gear ratio ict calculated in step (S214) is also searched by the map of FIG. 13 (target tilt angle calculation means). Next, in step (S232), the required tilt speed dφ / dt is calculated from the difference between the target tilt angle φt and the current tilt angle φ (target tilt speed calculation means). This calculation is equivalent to the differential calculation performed in a fixed control cycle, so the required tilt speed d
φ / dt represents the changing speed of the tilt angle required for gear shifting. In the drawings, dφ / dt is represented by a symbol with a dot above φ. Next, in step (S234), C
The sum Na of VU input / output rotations is calculated by the equation Na = Ni + Nco. Note that Ni is IVT input rotation, that is, CVU input rotation, and Nco is CVU output rotation. CVU output rotation N
co is calculated in consideration of the CVU output rotation from the CVU output rotation sensor 11 and the reduction ratio of the output chain mechanism 42. Next, in step (S235), an approximate value of the power roller displacement Ya is obtained by the equation Ya = Gf · (dφ / dt) / Na (power roller displacement calculation means). The displacement of the power roller 44 is approximately proportional to the tilt angle change speed (required tilt speed dφ / dt) and inversely proportional to the sum Na of the input and output rotations (described later in detail). Therefore, some coefficient Gf is (d
By multiplying φ / dt) / Na, it is possible to obtain it to some extent. To be precise, the tilt angle position of the power roller 44 is also
Since Gf is affected, it is possible to accurately obtain the power roller displacement Ya by making Gf a function of the tilt angle, but it is possible to sufficiently secure the effect of this control even with this method ( Power roller displacement calculation means according to claim 6. In addition, Gf can be obtained both theoretically and experimentally. Next, in step (S236), the shift speed correction step number Sv is set to the power roller displacement Ya.
It is calculated by multiplying by the gain Gy (actuator command value correction means). Here, the gain Gy is obtained by converting the power roller displacement Ya into the displacement of the step motor 36 according to the link ratio of the speed change link 37, and further converting it into the number of steps. Therefore, this gain Gy is converted into the speed change link design parameter and the number of steps of the step motor 36. Therefore, this gain Gy is a fixed value determined by the relationship between the speed change link design parameter, the number of steps of the step motor 36, and the displacement, and the gain does not need to be adapted.

Next, FIG. 6c is a diagram showing a flow of calculating the FB step number Sfb performed in step (S3) of the overall flow chart 6a. First, in step (S31),
The deviation eNi between the target input speed Nit and the measured input speed Ni is obtained. Next, in step (S32), the F / B step amount SP for proportional control is obtained by the product of the proportional constant KP and the deviation eNi. Next, in step (S33),
The F / B step number SI for the integral is obtained by adding the product of the integral constant KI and the deviation eNi to the step number SI for the integral integrated value up to the previous time. The number of steps SI for the integration up to the previous time SI
Does not show in the flow that SI is reset to SI = 0 at the start of the overall control. Then, in step (S34), the FB step number Sfb is calculated by adding the proportional SP and the integral SI, respectively.

[Relationship between Power Roller Displacement and Tilt Speed] First, the speed change operation of the CVU 4 is performed by the power roller 4
The rotation center of No. 4 is offset with respect to the center of rotation of the disk, and this offset causes the power roller 44 to tilt by the side slip force generated at the contact portion between the power roller 44 and the input / output disks 41, 43. However, when the offset amount is increased, the side slip force is increased and the shift speed is increased. That is, according to the present invention, the displacement of the power roller 44 in the tilt axis direction (Y
This is an invention in which the target gear ratio is followed by increasing the displacement = offset amount. Hereinafter, the relationship between the power roller displacement (hereinafter, y displacement) and the tilting speed (dφ / dt) will be described.

Tilt angle φ, y displacement between input disk 41 and power roller 44, input rotation speed ωi, tilt speed d
φ / dt has a relationship represented by the equation (1) described later. In the equation, θ is the half-vertical angle of the power roller, R is the radius of the power roller, and k (= e / r0) is the cavity center offset. The tilt angle φ, the y displacement, the output rotation speed ωo, and the tilt speed dφ / dt between the output disk 43 and the power roller 44 have a relationship expressed by the equation (2) described later. When this is represented by the input rotation speed ωi, there is a relationship represented by the equation (2) ′ described later. FIG. 15 is a graph showing the speed change (tilt) speed per 1 mm of y displacement according to these equations. In FIG. 15, the downward-sloping graph represents the equation (1), and the downward-sloping graph represents the equation (2) ′. Of these, the input speed Ni
The equations (1) and (2) 'when 1000 rpm is as shown in FIG. Therefore, when the relational expressions of the tilt angle φ, the y displacement, the input rotation speed ωi, and the tilt speed dφ / dt expressed in the expressions (1) and (2) ′ are approximately shown, y = Gi · ( dφ / dt) / ωi, and this Gi
The calculation is simplified by obtaining the value for each input rotation speed. However, as can be seen from FIG. 16, there is a large variation in the shift (tilt) speed depending on the magnitude of the tilt angle φ.
Equation (2) 'is the average of the equation, and the tilt angle φ is, for example,
Consider as 62.5 deg fixed. From this idea, Gi is y displacement = 1, tilting speed dφ / dt≈110, input rotation speed ωi = 1000r
Obtained as pm. For each input rotation (for example, 500rp
m, 1000 rpm, 1500 rpm ...), the y displacement can be easily calculated (power roller displacement calculating means in claim 3).

Similarly, the tilt angle φ, the y displacement, the output rotation speed ωo, and the tilt speed dφ / dt between the output disk 43 and the power roller 44 are expressed by the equation (3) described later. There is. When the expression (1) is expressed by the output rotation speed ωo, the expression (4) described later is obtained.
There is a relationship represented by. When the shift (tilt) speed per 1 mm of y displacement is represented in the graph according to these equations, the inclination is opposite to that in FIG. 15, although not shown. Of these, output speed
The equations (3) and (4) when the No is 1000 rpm are as shown in FIG. Therefore, when the relational expressions of the tilt angle φ, the y displacement, the output rotation speed ωo, and the tilt speed dφ / dt expressed by the equations (3) and (4) are approximately shown, y = Go · (dφ / dt) / ωo and this Go
The calculation is simplified by calculating the value for each output speed. However, as can be seen from FIG. 17, there is a large variation in the shifting (tilting) speed depending on the size of the tilt angle φ.
Equation (4) is the average, and the tilt angle φ is, for example, 6
Consider as fixed at 2.5deg. From this concept, Go has y displacement = 1, tilt speed dφ / dt ≈ 110, output speed ωo = 1000 rpm.
Is obtained as. For each output rotation (for example, 500 rpm,
1000 rpm, 1500 rpm ...), the y displacement can be easily calculated (power roller displacement calculating means in claim 4).

As described above, a method of calculating Gi and Go for each of the input rotation speed and the output rotation speed from FIGS. 16 and 17 may be used, but since there is still variation depending on the tilt angle φ, the input rotation speed and the output When Ga is calculated for each rotation speed that is the sum of the rotation speeds, the variation due to the size of the tilt angle φ is reduced, and the control accuracy is improved. For example, when the sum of input and output speeds is 2000 rpm, Ga is almost constant regardless of tilt angle φ, with y displacement = 1, speed (tilt) speed dφ / dt ≈ 110, and sum of input and output speeds = 2000 rpm. (See FIG. 18). The relational expression of the tilt angle φ, y displacement between the input disk 41 and the power roller 44, the sum ωa of the input and output rotational speeds, and the tilting speed dφ / dt is the expression (5) described later, and the output Tilt angle φ between disk 43 and power roller 44, y displacement, sum of input / output rotation speed ω
The relational expression of “a” and the tilting speed dφ / dt is the equation (6) described later. Therefore, when the relational expression of the tilt angle φ, the y displacement, the sum ωa of the input and output rotation speeds, and the tilt speed dφ / dt expressed by the equations (5) and (6) is approximately shown, y = Ga・ (Dφ / dt) / ωa,
The calculation is simplified by obtaining this Ga for each sum of input and output rotation speeds. From the above, in step (S235), the power roller 4 is calculated according to the equation of Ya = Gf · (dφ / dt) / Na.
The Y displacement of 4 can be obtained, which is more accurate than the calculation by only the input rotational speed or the output rotational speed (power roller displacement calculating means of claim 5).

[0061]

[Shifting Action During Sudden Braking] FIG. 19 is a time chart showing a comparison of the shifting ratio control actions of the first embodiment and the conventional example in the case of stopping by braking. First, in the case of the conventional example, when the brake operation is performed at the time t0 and a sudden shift request is issued, the shift speed of the toroidal type continuously variable transmission mechanism 4 is delayed with respect to this sudden shift request, and the target CV is delayed.
The U gear ratio cannot be followed, the engine stalls at time t1 due to a large creep torque, and then the gear cannot be changed. On the other hand, in the first embodiment, when a sudden shift request is issued, the shift speed correction step number Sv (the hatched portion of the STEP number characteristic in FIG. 19) is added to the feedforward step number Sff for the step motor 36, and By ensuring a certain offset amount with respect to the roller 44, the speed change speed of the toroidal type continuously variable transmission mechanism 4 is increased, the target CVU speed change ratio is followed, and the IVT speed ratio (IVT speed change ratio is changed without engine stall). ratio)
The gear shift ends at time t2 when G becomes GNP.

Next, the effect will be described.

In the speed change control device for an infinitely variable speed ratio continuously variable transmission according to the first embodiment, even when the CVU 4 is controlled at a very high speed by performing the speed ratio control as described above. A high gear shift followability can be obtained. Very high deceleration occurs especially during sudden braking,
Moreover, the range in which the CVU 4 is shifted is large with respect to changes in the vehicle speed. The target of the IVT input rotation speed (= CVU input rotation speed) at that time is a low value, and a very high shift speed and a large power roller offset are required to achieve the shift speed in accordance with the high deceleration. . By this control, the shift of the shift control valve opening corresponding to the power roller offset can be corrected, which is very effective in preventing the engine stall. Further, since the shift speed is compensated by the F / F, the F / B can be increased in order to improve the target followability.
It is not necessary to increase the control gain, and highly responsive shift control can be realized without impairing control stability. In addition, in the first embodiment, most of the F / F control parameters can be designed based on the theory and experimental values, so that there is an advantage that the number of man-hours for adaptation in the vehicle can be extremely reduced.

(Second Embodiment) In the second embodiment, if the tilting speed dφ / dt of the power roller 44 is equal to or lower than a predetermined speed, a dead zone region in which the shift speed correction is not applied is set, and the toroidal type continuously variable In this example, the higher the input / output speed of the transmission mechanism 4 is, the wider the dead zone in which the speed change correction is not applied is set.

First, the structure will be described. FIG. 20 shows a shift speed correction step number calculation block 53 'of the second embodiment. Only the shift speed correction step number calculation block is different from the first embodiment, and other configurations are the same as those of the first embodiment (see FIG. 1).
To FIG. 5), illustration and description thereof will be omitted.

Next, the operation will be described.

FIG. 21 is a flowchart for calculating the number of FF steps.
It is a figure which shows the calculation flow of the shift speed correction | amendment step number Sv of 2nd Example performed by step (S23) of this.

First, from step (S231a) to step (S234a), the step (S23) of the first embodiment is performed.
This is common to 1) to step (S234).

Next, in step (S235a),
The approximate value of the power roller displacement Ya is Ya = Gf · (dφ / dt) /
Tilt speed dφ /
From the sum Na of dt and input / output rotation, shift speed correction step number Sv
Are searched based on the map of FIG. In the map shown in FIG. 22, the shift speed correction step number Sv is plotted with respect to the tilt speed dφ / dt for each sum Na of input and output rotations,
It is obtained by searching the map from each parameter.

In this map, as the tilting speed dφ / dt increases, the absolute value of the shift speed correction step number Sv | Sv
As the sum | of input and output rotations becomes smaller as well as | becomes larger, the shift speed correction step number Sv becomes larger. And, to some extent, the tilting speed dφ /
A slight dead zone is provided so that the shift speed correction step number Sv does not occur unless dt occurs. The dead zone is set to increase as the sum Na of input / output rotations increases.

Next, the effect will be described.

In the speed change control device of the continuously variable transmission having an infinite transmission ratio of the second embodiment, as described above, the power roller 44 is used.
If the tilting speed dφ / dt is less than or equal to a predetermined speed, a dead zone region is set in which the speed change speed is not corrected, and the speed change speed is corrected as the input / output speed of the toroidal type continuously variable transmission mechanism 4 increases. Since the dead zone that is not added is set wide, it prevents the compensation amount from becoming too sensitive due to the calculation error of the required tilt (shift) speed, etc. to prevent the stability from deteriorating. It is possible to secure a sufficient compensation amount and secure high responsiveness. Further, by increasing the dead zone in a state where the CVU input / output rotation speed is high in which the shift speed is easily obtained, it is possible to optimize stability and responsiveness.

(Third Embodiment) The third embodiment is an example in which the shift speed is corrected by using the deceleration of the output shaft rotation of the IVT only in the power circulation mode and during the brake operation.

23, a shift speed correction step number calculation block 53 "of the third embodiment is shown in FIG. 23. Only this shift speed correction step number calculation block 53 differs from that of the first embodiment. The configuration of the first embodiment (Fig. 1
To FIG. 5), illustration and description thereof will be omitted.

Next, the operation will be described.

FIG. 24a is a flow chart for calculating the number of FF steps.
It is a figure which shows the calculation flow of the shift speed correction | amendment step number Sv of 3rd Example performed by the step (S23) of b. First, in step (S231b), it is determined whether the brake SW is ON and is in the power circulation mode. If the brake is OFF or the direct connection mode, the process proceeds to step (S233b), and the shift speed correction step number Sv
However, Sv = 0 is set. When the condition of step (S231b) is satisfied, the process proceeds to step (S233b), and the deceleration dNo / dt which is the differential value of the IVT output rotation speed No.
The shift speed correction step number Sv is calculated in accordance with (indicated by No in the drawing with a dot). That is, FIG.
In step (S233b1), the IVT measured last time
The deceleration dNo / dt is detected from the difference between the output rotation speed and the IVT output rotation speed measured this time, and the step (S233b1)
Then, the shift speed correction step number Sv is retrieved from the map of FIG. 25 for the deceleration dNo / dt.

When the brake is turned on in the power circulation mode, this control is most necessary as described in the problems of the prior art. By applying this control only in this situation, the engine stall can be prevented with a very simple logic. In this situation, the target input rotation is almost constant from idle to near the fuel cut recovery rotation, and the deceleration dN of the IVT output shaft 2 is dN.
By measuring o / dt, the speed change speed of the toroidal-type continuously variable transmission item 4 can be estimated to some extent. Further, since the input rotation does not change significantly, the shift speed with respect to the displacement of the power roller does not change significantly, so that the parameters used for map search can be reduced.

Next, the effect will be described.

In the transmission control device for an infinitely variable transmission having a continuously variable transmission according to the third embodiment, the deceleration dNo of the output shaft rotation of the IVT is only in the power circulation mode and when the brake is operated as described above. Since the shift speed is corrected using / dt, it is possible to reduce not only the controller calculation load when the brake is OFF, but also the calculation load of the IVT controller 20 when the brake is ON as compared with the other embodiments. Is. Further, it is possible to use the signal of the ABS wheel speed sensor for the deceleration of the output shaft rotation. In this case, a considerably high accuracy can be expected in the deceleration calculation. It also becomes possible to perform control.

(Other Embodiments) The gear shift control device of the transmission having the toroidal type continuously variable transmission mechanism of the present invention has been described above based on the first to third embodiments. However, the specific configuration will be described. The present invention is not limited to these embodiments, and design changes and additions are allowed without departing from the gist of the invention according to each claim of the claims.

For example, in the first to third embodiments, the application example to the continuously variable transmission having an infinite gear ratio is shown. However, the toroidal type continuously variable transmission in which the rotation of the output disk is directly used as the transmission output rotation of the transmission. It can also be applied to machines.

[Brief description of drawings]

FIG. 1 is an overall configuration diagram showing an infinitely variable transmission continuously variable transmission to which a shift control device of a first embodiment is applied.

FIG. 2 is a diagram showing a toroidal type continuously variable transmission mechanism of an infinitely variable transmission continuously variable transmission to which the gear shift control device of the first embodiment is applied.

FIG. 3 is a diagram illustrating a gear ratio control system in a gear shift control device for an infinite gear ratio continuously variable transmission according to a first embodiment.

FIG. 4 is a diagram showing an IVT speed ratio characteristic by a shift control of the infinitely variable transmission continuously variable transmission according to the first embodiment.

FIG. 5 is a shift control block diagram of an infinitely variable transmission continuously variable transmission according to the first embodiment.

FIG. 6 is a flowchart showing a flow of a shift control process performed by the IVT controller of the first embodiment.

FIG. 7 is a map for obtaining a reaching target input rotation according to a vehicle speed and a throttle opening.

FIG. 8 is a map for obtaining a target CVU gear ratio from a target IVT speed ratio.

FIG. 9 is a map for obtaining a basic step number from a target CVU gear ratio.

FIG. 10 is a map for obtaining engine torque from IVT input rotation and throttle opening.

FIG. 11: C from engine torque based on IVT speed ratio
It is a map which calculates | requires VU input torque.

FIG. 12 is a map for obtaining the CVU input torque correction step number from the CVU gear ratio on the basis of CVU input torque.

FIG. 13 is a map showing a relationship between a CVU gear ratio and a CVU power roller tilt angle.

FIG. 14 is an explanatory diagram of parameters used in a relational expression between the displacement of the power roller and the tilting speed.

FIG. 15 is a graph showing a relational expression between the input disc and the power roller and a relational expression between the output disc and the power roller when the input rotation is variously changed.

FIG. 16 is a first speed change speed when the input rotation is constant,
It is a graph which shows a 2nd speed change speed and an average speed change speed.

FIG. 17 is a first speed change speed when the output rotation is constant,
It is a graph which shows a 2nd speed change speed and an average speed change speed.

FIG. 18 is a graph showing the first speed change speed, the second speed change speed, and the average speed change speed when the sum of input and output rotations is constant.

FIG. 19 is an engine speed characteristic and CV according to the control of the conventional control and the control of the first embodiment when stopping by brake operation.
7 is a time chart showing a U gear ratio characteristic, a step number characteristic, and a power roller y displacement characteristic.

FIG. 20 is a diagram showing a shift speed correction step number calculation block of a second embodiment.

FIG. 21 is a diagram showing a calculation flow of a shift speed correction step number according to the second embodiment.

FIG. 22 is a map for obtaining the shift speed correction step number from the sum of the tilt speed and the input / output rotation speed.

FIG. 23 is a diagram showing a shift speed correction step number calculation block of a third embodiment.

FIG. 24 is a diagram showing a calculation flow of a shift speed correction step number in the third embodiment.

FIG. 25 is a map for obtaining the shift speed correction step number from the deceleration of the IVT output shaft.

[Explanation of symbols]

1 IVT input shaft 2 IVT output shaft 3 differential 4 Toroidal continuously variable transmission (CVU) 41 Input Disc 42 output chain mechanism 43 Output disc 44 power roller 45 trunnion 5 constant speed change mechanism 6 Planetary gear mechanism 61 Sun Gear 62 pinion carrier 63 ring gear 8 Power circulation mode clutch 9 Direct connection mode clutch 10 CVU input rotation sensor 11 CVU output rotation sensor 13 sensor gears 14 Output rotation sensor (vehicle speed sensor) 20 IVT controller 30 hydraulic servo 31 Servo piston 35 Precessum 36 step motor (shift actuator) 37 speed change link 38 L-shaped link 46 shift control valve 46S variable speed spool

─────────────────────────────────────────────────── ─── Continuation of the front page (51) Int.Cl. ⁷ Identification code FI theme code (reference) F16H 63:06 F16H 63:06 F term (reference) 3J062 AA18 AB06 AB35 AC03 BA29 BA40 CG35 CG52 CG82 3J552 MA09 NA01 NB01 PA20 PA56 SA32 SA44 TA06 VA24W VA24Y VA32W VA37W VA74Z VA77Z VB01W VC03 VD02W VD11W

Claims (10)

[Claims] 1. An input disk and an output disk which are coaxially opposed to each other, a power roller sandwiched between the input and output disks, a power roller rotatably supporting the power roller, and a power roller associated with a shift. A toroidal continuously variable transmission having a trunnion having a tilt axis, a servo piston that strokes the trunnion in the tilt axis direction, and a speed change actuator that generates a servo pressure to the servo piston; A toroidal type continuously variable transmission having a target gear ratio calculating means for calculating a target gear ratio of the toroidal type continuously variable transmission mechanism, and performing gear shift control by sending an actuator command value according to the calculated target gear ratio to the gear shift actuator. In a shift control device for a transmission having a gear shift mechanism, an actual tilting for calculating an actual tilting angle of the trunnion is provided. Degree calculation means, a target tilt angle calculation means for calculating a target tilt angle of the trunnion from the target gear ratio, and a target of the toroidal continuously variable transmission mechanism from the calculated actual tilt angle and target tilt angle. Target tilt speed calculation means for calculating tilt speed, power roller displacement calculation means for calculating displacement of the power roller in the trunnion tilt axis direction according to the calculated target tilt angle, and calculated power roller A shift control device for a transmission having a toroidal type continuously variable transmission mechanism, comprising: an actuator command value correcting means for correcting the actuator command value in accordance with a displacement. 2. The shift control device for a transmission having a toroidal type continuously variable transmission mechanism according to claim 1, wherein the actuator command value correction means is an input rotation speed or an output of the toroidal type continuously variable transmission mechanism. Number of revolutions,
Alternatively, the shift control device for a transmission having a toroidal type continuously variable transmission mechanism is characterized in that it is means for performing a larger correction as the sum of the input rotation speed and the output rotation speed is smaller. 3. The shift control device for a transmission having the toroidal type continuously variable transmission mechanism according to claim 1, wherein the power roller displacement calculation means is configured to operate the input disk and the power at a predetermined tilt angle. A first shift speed obtained from a change in tilt angle between the rollers, a second shift speed obtained from a change in tilt angle between the output disc and the power roller at a predetermined tilt angle, or an average shift speed thereof. Is a means for calculating the power roller displacement by dividing by the input rotation speed of the toroidal type continuously variable transmission mechanism, and a transmission control device for a transmission having a toroidal type continuously variable transmission mechanism. 4. A gear shift control device for a transmission having a toroidal type continuously variable transmission mechanism according to claim 1, wherein the power roller displacement calculation means is configured such that the power disc displacement calculation means and the input disc and the power at a predetermined tilt angle. A first shift speed obtained from a change in tilt angle between the rollers, a second shift speed obtained from a change in tilt angle between the output disc and the power roller at a predetermined tilt angle, or an average shift speed thereof. Is a means for calculating the power roller displacement by dividing by the output rotation speed of the toroidal type continuously variable transmission mechanism, and a transmission control device for a transmission having a toroidal type continuously variable transmission mechanism. 5. The transmission control device for a transmission having the toroidal type continuously variable transmission mechanism according to claim 1, wherein the power roller displacement calculation means is configured to operate the input disk and the power at a predetermined tilt angle. A first shift speed obtained from a change in tilt angle between the rollers, a second shift speed obtained from a change in tilt angle between the output disc and the power roller at a predetermined tilt angle, or an average shift speed thereof. Is a means for calculating the power roller displacement by dividing by the sum of the input rotation speed and the output rotation speed of the toroidal type continuously variable transmission mechanism, and the shift control of the transmission having the toroidal type continuously variable transmission mechanism. apparatus. 6. A transmission control device for a transmission having a toroidal type continuously variable transmission mechanism according to claim 1, wherein the power roller displacement calculation means is configured to control the input disk and the power at a predetermined tilt angle. A first shift speed obtained from a change in tilt angle between the rollers, a second shift speed obtained from a change in tilt angle between the output disc and the power roller at a predetermined tilt angle, or an average shift speed thereof. Is the input speed of the toroidal type continuously variable transmission,
Alternatively, it is a means for calculating the power roller displacement by multiplying the output rotation speed or a value calculated by dividing the sum of the input rotation speed and the output rotation speed by a variable according to the tilt angle. A shift control device for a transmission having a toroidal type continuously variable transmission mechanism. 7. A shift control device for a transmission having a toroidal-type continuously variable transmission mechanism according to claim 1, wherein the actuator command value correction means is the first shift speed, or A shift control device for a transmission having a toroidal type continuously variable transmission mechanism, characterized in that it is means for not making a correction if the second shift speed or an average shift speed thereof is equal to or lower than a predetermined speed. 8. The shift control device for a transmission having a toroidal type continuously variable transmission mechanism according to claim 7, wherein the actuator command value correction means is an input rotation speed or an output of the toroidal type continuously variable transmission mechanism. A shift control device for a transmission having a toroidal-type continuously variable transmission mechanism, characterized in that it is means for setting a region where correction is not performed wider as the rotational speed is higher. 9. A transmission control device for a transmission having a toroidal type continuously variable transmission mechanism according to claim 1, wherein the transmission having the toroidal type continuously variable transmission mechanism is a toroidal type connected to a unit input shaft, respectively. A planetary gear mechanism including a continuously variable transmission mechanism and a constant transmission mechanism, a sun gear connected to the output shaft of the toroidal type continuously variable transmission mechanism, a carrier connected to the output shaft of the constant transmission mechanism, and a ring gear connected to the unit output shaft. A gearshift having a power circulation mode clutch provided between the output of the constant speed change mechanism and the carrier, and a direct connection mode clutch provided between the output of the toroidal type continuously variable transmission and the unit output shaft The infinite ratio continuously variable transmission, wherein the actuator command value correction means has the power circulation mode clutch engaged and the direct connection mode clutch released. In the force circulation mode and only when the brake is operated, a command to the speed change actuator of the toroidal type continuously variable transmission is means for correcting the speed ratio of the infinitely variable transmission continuously variable transmission to the neutral side. Gear shift control device for a transmission having a toroidal type continuously variable transmission mechanism. 10. The transmission control device for a transmission having a toroidal-type continuously variable transmission according to claim 9, wherein the target tilt speed calculation means reduces the output shaft rotation of the continuously variable transmission with an infinite transmission ratio. A shift control device for a transmission having a toroidal type continuously variable transmission mechanism, characterized in that the shift control device is a means for calculating a target tilting speed using the speed.