JPH06109113A - Speed change ratio control device of continuously variable transmission - Google Patents

Abstract

PURPOSE:To provide a speed change ratio control device of a continuously variable transmission capable of sufficiently improve the acceleration feeling and increasing the speed change responsiveness at the initial acceleration condition. CONSTITUTION:The stationary target speed of a primary pulley 31 is computed by a control unit CU in the normal condition, and the primary speed is feedback controlled so as to follow this stationary target speed. On the other hand, in the transitional condition (acceleration and deceleration), the throttle opening after the time period of delay in response of the control system is expected based on the throttle opening and the rate of change by means of the control unit CU, and the transitional target speed of the primary pulley 31 is computed by the annealing treatment based on this expected value of the throttle opening. The primary speed is feedforward controlled followed by this transitional target speed, increasing the speed change responsiveness in the transitional condition.

Description

Detailed Description of the Invention

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a gear ratio control device for a continuously variable transmission.

[0002]

2. Description of the Related Art Generally, an automobile is provided with a transmission that changes the output torque of an engine in accordance with the driving state. , 4 forward and 1 reverse) are often used. However, in the multi-stage transmission, since it is not possible to set a gear ratio other than several preset gear ratios, there is a problem that the gear ratio most suitable for the operating state of the vehicle cannot be obtained. There is a problem that a gear shift shock occurs when changing gears.

Therefore, in recent years, there has been proposed a continuously variable transmission capable of setting a gear ratio to an arbitrary value within a predetermined range,
Specifically, for example, a belt type continuously variable transmission (hereinafter, referred to as CVT) has been proposed. In such a CVT, usually, a primary pulley (driving pulley) and a secondary pulley (driven pulley) whose pulley diameters can be changed respectively, and a belt wound around both pulleys are provided, and for example, a hydraulic piston is used. By changing the pulley diameters of both pulleys, it is possible to set a gear ratio of an arbitrary value.

In such a CVT, the gear ratio is basically controlled according to the vehicle speed and the throttle opening (engine load). Specifically, for example, based on the secondary pulley rotation speed (corresponding to the vehicle speed) and the throttle opening, a gear ratio most suitable for such operating conditions or a physical quantity indirectly representing the gear ratio, for example, the primary pulley rotation speed, The gear ratio control is performed by calculating and using this as a control target value, and performing feedback control such that the actual gear ratio or the above physical quantity follows the control target value.

By the way, in a CVT in which the above gear ratio control is performed with the primary pulley rotation speed as the control target, for example, during sudden acceleration, the throttle opening sharply rises in a substantially stepwise manner within a short time. The primary pulley rotation speed target value (control target value) also suddenly rises in a substantially stepwise manner. In this case, since a large deviation occurs between the primary pulley rotation speed target value and the actual primary pulley rotation speed, the primary pulley rotation speed is rapidly increased to eliminate the deviation. However, when the primary pulley rotation speed is rapidly increased in this way, there is a problem that a phenomenon in which the primary pulley rotation speed greatly exceeds the target value, that is, an overshoot occurs. Also, in this case, the engine speed rapidly increases as the primary pulley speed rapidly increases.However, since the engine has a considerable rotational inertia, a significant part of the engine torque is due to this rotational inertia. It is consumed to increase the number of vehicles, which causes a slow increase in the driving force of the automobile, and there is a problem that good acceleration performance or sense of acceleration cannot be obtained.

Therefore, at the time of predetermined acceleration, a transient control target value having a smaller change amount than the control target value at the normal time (for example, the primary pulley rotation speed target value) is set to suppress a rapid increase in the control amount at the time of acceleration. CVT is designed to prevent the occurrence of overshoot, reduce the engine torque consumed to increase the engine speed by suppressing the rapid increase of the engine speed, and enhance the acceleration or feeling of acceleration of the vehicle. (For example, Japanese Patent Publication No. 3-27)
791 publication). In such a conventional CVT,
Normally, the transient control target value is set, for example, by performing a smoothing process on the normal control target value, and the shift operation is started from the time when the transient control target value reaches a predetermined noise threshold value. It is like this.

[0007]

However, the conventional CVT in which the transient control target value is set by performing the smoothing process on the control target value for the normal time as described above.
Then, in the initial stage of acceleration, the start of the rise of the transient target value is delayed with respect to the start of the increase of the throttle opening, so that there is a problem that the shift responsiveness in the initial stage of acceleration deteriorates.

The present invention has been made in order to solve the above-mentioned problems of the prior art, in which a normal control target value is set during normal operation, while a transient control target value is set during predetermined acceleration, and an acceleration feeling is set. It is an object of the present invention to provide a gear ratio control device for a continuously variable transmission that can sufficiently improve the gear ratio and enhance the gear shift response in the initial stage of acceleration.

[0009]

In order to achieve the above object, as shown in FIG. 1, the first invention is a control target for setting a control target value of a predetermined physical quantity indicating a shift characteristic based on an engine load. A value setting means A, a shift control means B for controlling the physical quantity so as to follow the control target value, an acceleration state detecting means C for detecting whether or not the engine is in a predetermined acceleration state, and the acceleration state. When the detection means C detects that the engine is in the predetermined acceleration state, the transient control target for setting a transient control target value having a smaller variation than the control target value set by the control target value setting means A. Continuously variable transmission E provided with value setting means D
In the gear ratio control device, the engine load predicting means F for calculating an engine load predictive value after a predetermined time based on the change rate of the engine load is provided, and the transient control target value setting means D is the engine load predicting means. Provided is a gear ratio control device for a continuously variable transmission, wherein a transient control target value is set based on an engine load predicted value calculated by F.

A second aspect of the present invention is the gear ratio control device for a continuously variable transmission E according to the first aspect, in which the engine load predicting means F is delayed by a time corresponding to the response delay of the transient control target value setting means D. The present invention provides a gear ratio control device for a continuously variable transmission, characterized in that it calculates an engine load prediction value.

A third aspect of the present invention is the gear ratio control device for a continuously variable transmission E according to the first or second aspect, wherein the transient control target value setting means D predicts the engine load only within a predetermined period after the start of acceleration. Provided is a transmission ratio control device for a continuously variable transmission, characterized in that a transient control target value is set based on the value.

A fourth aspect of the present invention is the gear ratio control device for a continuously variable transmission E according to any one of the first to third aspects, wherein the transient control target value setting means D controls the transient control target value and the control. A transmission ratio control device for a continuously variable transmission, characterized in that a next transient control target value is set based on a deviation from the control target value set by the target value setting means A. provide.

[0013]

EXAMPLES Examples of the present invention will be specifically described below.
As shown in FIG. 2, the power train PT for automobiles is
4-cylinder engine CE including first to fourth cylinders # 1 to # 4
And a hydraulically actuated transmission CT. Here, the engine CE outputs the engine torque to the crankshaft 1
Output is made to the transmission CT side via the (engine output shaft). Further, the transmission CT shifts the torque of the transmission input shaft 2 that rotates integrally with the engine output shaft 1 according to the operating state, and reverses the rotation direction when the reverse range is selected to change the transmission output shaft 2. It is designed to output to 3. The transmission output shaft 3
Is transmitted to the drive wheels (not shown) via the reduction gear mechanism 4 and the differential device 5.

The transmission CT includes a torque converter 7 that shifts the torque of the transmission input shaft 2 via hydraulic oil and outputs the torque to the turbine shaft 6, and rotation of the turbine shaft 6 when the reverse range is selected. A forward / reverse switching mechanism 9 that reverses and transmits to the intermediate shaft 8 and a belt type continuously variable transmission 10 that continuously changes the torque of the intermediate shaft 8 and outputs it to the transmission output shaft 3 (hereinafter, referred to as CV
(T10) is provided.

The torque converter 7 includes a pump cover 11
A pump 12 connected to a transmission input shaft 2 via a turbine, a turbine 14 connected to a turbine shaft 6 via a connecting member 13 and driven to rotate by hydraulic fluid discharged from the pump 12, and a turbine 14 to a pump 12 And a stator 15 that rectifies the hydraulic oil that flows back to the pump 12 in a direction that assists the rotation of the pump 12. The transmission has a gear ratio corresponding to the speed ratio of the pump 12 and the turbine 14 (turbine rotation speed / pump rotation speed). The torque of the input shaft 2 is changed. Here, the stator 15 is a one-way clutch 16
It is fixed to the transmission case 25 (fixed portion) via the.

Further, the torque converter 7 is provided with a lockup clutch 17 for directly connecting (locking up) the transmission input shaft 2 and the turbine shaft 6 in a predetermined operating region in order to improve fuel efficiency. The lockup clutch 17 is locked up (on) when hydraulic pressure is applied to the rear oil chamber 17r from a hydraulic mechanism FS, which will be described later, and released when hydraulic pressure is applied to the front oil chamber 17f. (Off). An oil pump 19 that is rotatably driven by the pump 12 (pump shell 49) via the connecting shaft 18 is disposed slightly behind (on the left side in FIG. 2) the torque converter 7.

The forward / reverse switching mechanism 9 is a planetary gear system. The forward / backward switching mechanism 9 includes a ring gear 21 connected to the turbine shaft 6 via a torque input member 20 and a sun gear connected to the intermediate shaft 8. 22
And a plurality of pinion gears 23 that mesh with the ring gear 21 and the sun gear 22, and rotate these pinion gears 23.
A carrier 24 that supports (rotatably) is provided. A forward clutch 26 is provided between the torque input member 20 and the carrier 24, and the reverse brake 2 is provided between the carrier 24 and the transmission case 25.
7 is provided. Here, the forward clutch 26
The reverse brake 27 is turned on (fastened) when hydraulic pressure is supplied from a hydraulic mechanism FS, which will be described later,
It is designed to be turned off (released) when the hydraulic pressure is released.

In the forward / reverse switching mechanism 9, when both the forward clutch 26 and the reverse brake 27 are turned off, the neutral state is established, and no torque is transmitted from the turbine shaft 6 to the intermediate shaft 8. When only the forward clutch 26 is turned on, the ring gear 21 and the carrier 24 cannot be differentiated from each other, so that the forward / reverse switching mechanism 9 is directly connected and the intermediate shaft 8 is integrated with the turbine shaft 6 in the same direction. It rotates and the drive wheels are driven in the forward direction.

When only the reverse brake 27 is turned on, the carrier 24 is fixed to the transmission case 25, so that the ring gear 21, the pinion gear 23 and the sun gear 2 are provided.
2 and 2 function as a fixed gear train that meshes in this order. At this time, the sun gear 22 rotates in the opposite direction to the ring gear 21, so that the intermediate shaft 8 becomes the turbine shaft 6.
And the drive wheels are driven in the reverse direction.
In this case, the gear is changed at a gear ratio determined by the number of teeth of the ring gear 21 and the number of teeth of the sun gear 22. The forward clutch 26 and the reverse brake 27 are not both turned on.

The CVT 10 includes a primary pulley 31 (driving pulley) that rotates integrally with the intermediate shaft 8, a secondary pulley 32 (driven pulley) that rotates integrally with the transmission output shaft 3, and a primary pulley 31 and a secondary pulley 32.
And a V-belt 33 that transmits torque between and. Note that, hereinafter, for convenience, the engine side (right side in FIG. 2) when viewed in the axial direction of the intermediate shaft 8 is referred to as “front” or “front”, and the opposite side is referred to as “rear” or “rear”.

The primary pulley 31 is the intermediate shaft 8
And a first movable conical plate 35 which is arranged on the rear side of the first fixed conical plate 34 so as to be opposed thereto and is movable in the front-rear direction. ing. A primary oil chamber 36 that controls the position of the first movable conical plate 35 in the front-rear direction is provided.
Here, when hydraulic pressure is applied to the primary oil chamber 36, hydraulic oil is supplied into the primary oil chamber 36, the first movable conical plate 35 moves to the front side, and the holding position of the V belt 33 changes to the outer peripheral side, The effective pulley diameter of the primary pulley 31 becomes large. Conversely, when the hydraulic pressure is released, the hydraulic oil in the primary oil chamber 36 is drained and the primary pulley 31
The effective pulley diameter becomes smaller. That is, the effective pulley diameter of the primary pulley 31 can be freely changed by supplying or discharging hydraulic pressure or hydraulic oil to or from the primary oil chamber 36.

The secondary pulley 32 has basically the same structure as that of the primary pulley 31, and includes a second fixed conical plate 37 fixed to the transmission output shaft 3 and a front side of the second fixed conical plate 37. The second which is arranged so as to face this
It is composed of a movable conical plate 38. A secondary oil chamber 39 is provided to control the position of the second movable conical plate 38 in the front-rear direction.

In the CVT 10, the hydraulic mechanism F
A hydraulic pressure (hereinafter, referred to as a primary hydraulic pressure) corresponding to the gear ratio to be set is applied from S to the primary oil chamber 36. On the other hand, the secondary oil chamber 39 basically has V
A hydraulic pressure sufficient to maintain the tension of the belt 33, that is, a minimum hydraulic pressure capable of transmitting the driving force without causing belt slip (hereinafter, referred to as secondary hydraulic pressure) is applied. That is, in CVT10,
The gear ratio is determined by the primary hydraulic pressure, and the belt tension is determined by the secondary hydraulic pressure. As will be described later, since the line pressure of the hydraulic mechanism FS is introduced into the secondary oil chamber 39, the secondary hydraulic pressure is substantially synonymous with the line pressure.

Specifically, when the primary hydraulic pressure rises, the effective pulley diameter of the primary pulley 31 increases accordingly. For this reason, the tension of the V-belt 33 tries to increase, but the secondary hydraulic pressure is adjusted so as not to increase this tension.
(Line pressure) is adjusted, and the effective pulley diameter of the secondary pulley 32 is reduced. In this way, the primary pulley 31
While the effective pulley diameter of the secondary pulley 3 increases
Since the effective pulley diameter of No. 2 becomes small, the gear ratio of the CVT 10 changes to the speed increasing side (OD side). On the other hand, when the primary oil pressure decreases, the gear ratio of the CVT 10 changes to the deceleration side (LOW side) contrary to the above case.

The hydraulic mechanism F is connected to the transmission CT.
S is provided, and this hydraulic mechanism FS responds to a signal from the control unit CU in accordance with the operating state, according to the operating state, the front oil chamber 17f and the rear oil chamber 17r of the lockup clutch 17, the forward clutch 26 of the forward / reverse switching mechanism 9 and the reverse clutch. The brake 27, the primary oil chamber 36 and the secondary oil chamber 39 of the CVT 10 are supplied with and discharged with operating oil or control oil pressure to perform a predetermined gear shift operation. Here, the line pressure (secondary pressure) of the hydraulic mechanism FS is also controlled by the control unit CU.

The hydraulic mechanism FS will be described below. As shown in FIG. 3, the hydraulic mechanism FS is supplied with operating oil (original pressure) from the oil pump 19. The hydraulic mechanism FS includes a line pressure adjusting valve 41, a pressure reducing valve 42, a gear ratio control valve 43, a gear ratio fixed valve 44,
A hydraulic pressure correction valve 45, a clutch valve 46, a manual valve 47, a relief valve 48, a lockup valve 49, etc. are provided. Here, the gear ratio control valve 4
3 is controlled by the first duty solenoid 51,
The fixed gear ratio valve 44 is controlled by the first on / off solenoid 52, the hydraulic pressure correction valve 45 is controlled by the second duty solenoid 53, and the clutch valve 46.
Is controlled by the clutch duty solenoid 54, and the lockup valve 49 is controlled by the second on / off solenoid 55.

In the hydraulic mechanism FS, the hydraulic oil discharged from the oil pump 19 is first adjusted to a predetermined line pressure by the line pressure adjusting valve 41, and the line L1
The oil is supplied to the secondary oil chamber 39 through the (hydraulic passage) and to the clutch valve 46 through the line L2. The clutch valve 46 adjusts the hydraulic pressure in the line L2 to a predetermined pressure by the clutch duty solenoid 54, and then supplies the adjusted hydraulic pressure to the manual valve 47 and the lockup valve 49 via the line L3. Has become. Pressure reducing valve 42
Reduces the line pressure supplied to the secondary oil chamber 39 to form pilot pressure for the hydraulic pressure correction valve 45, the gear ratio control valve 43, the gear ratio fixed valve 44, and the clutch valve 46.

The pilot pressure for controlling the line pressure is adjusted by controlling the duty ratio of the second duty solenoid 53. That is, the hydraulic pressure controlled by the second duty solenoid 53 is introduced into the pilot chamber of the hydraulic pressure correction valve 45, the hydraulic pressure correction valve 45 is opened / closed according to this hydraulic pressure, and the line L4 in the line L4 formed according to this open / closed state is opened. Hydraulic pressure is line pressure adjustment valve 4
It is introduced as a pilot pressure of 1 to obtain a desired line pressure. The hydraulic pressure correction valve 45
Alternatively, the line pressure adjusting valve 41 may be directly controlled by a duty solenoid or the like.

The gear ratio control valve 43 is controlled by the first duty solenoid 51, and the gear ratio control valve 4
The hydraulic pressure in the line L6 formed by 3 is supplied to the primary oil chamber 36 via the gear ratio fixed valve 44. The fixed gear ratio valve 44 is controlled by the first on / off solenoid 52. When the first on / off solenoid 52 is in the on state, the line L7 connected to the primary oil chamber 36 communicates with the line L6, while in the off state. The communication is cut off. In other words, by turning off the first solenoid 52, the hydraulic pressure applied to the primary oil chamber 36 is fixed to the current value irrespective of the operation of the gear ratio control valve 43, thereby fixing the gear ratio. It has become.

The gear ratio control valve 43 is controlled by the first duty solenoid 51, and when the first duty solenoid 51 is in the ON state, the hydraulic pressure in the primary oil chamber 36 is in order of line L7, line L6 and line L8. And drained through the relief ball 58,
No oil pressure is applied to the primary oil chamber 36. On the other hand, the first
When the duty solenoid 51 is off,
While the line L8 (drain path) is closed, the transmission ratio control valve 43 is opened at an opening ratio according to the duty ratio of the first duty solenoid 51, and the line pressure is changed to the orifice 59.
And is introduced into the primary oil chamber 36 via the line L6. Since the orifice 59 is provided, the oil chamber in the primary oil chamber 36 does not suddenly rise.

The clutch valve 46 is controlled by the clutch duty solenoid 54, and the line pressure adjusted by the clutch duty solenoid 54 is
It is supplied to the manual valve 47 and the lockup control valve 49 via the line L3. This adjusted line pressure is applied to the forward clutch 2 via the line L3, the manual valve 47 and the line L10 in the forward drive state.
6, while the hydraulic pressure in the reverse brake 27 is released through the line L12. On the other hand, in the reverse drive state, the line pressure is set to the line L3 and the line L1 only when the lockup valve 49 is in the non-lockup state.
3 is supplied to the reverse brake 27 via the line L12.

The lockup valve 49 is controlled by the second on / off solenoid 55, and at the time of lockup,
The line L16 connected to the front hydraulic pressure 17f communicates with the relief valve 48 via the relief line L15. On the other hand, when the lockup is released, the line L17 connected to the rear oil chamber 17r communicates with the relief valve 48 via the relief line L15.

Next, the control mechanism of the transmission CT will be described. As shown in FIG. 4, the control mechanism of the transmission CT includes:
Control unit C consisting of a microcomputer
U is provided. Then, the control unit CU includes a shift position signal (P, R, N, D, 2, 1) detected by the shift position sensor 62, a rotation speed of the primary pulley 31 detected by the primary rotation speed sensor 63 ( Hereinafter, this is referred to as the primary rotation speed), the rotation speed of the secondary pulley 32 detected by the secondary rotation speed sensor 64 (hereinafter, referred to as the secondary rotation speed), the throttle opening detected by the throttle opening sensor 65, the engine The engine speed detected by the speed sensor 66, the turbine speed detected by the turbine speed sensor 67, the oil temperature detected by the oil temperature sensor 68, the oil pressure detected by the oil pressure sensor 69, etc. are input as control information. It is supposed to be done.

The control unit CU includes a control target value setting means, a shift control means,
A comprehensive control device for a transmission CT including an acceleration state detection means, a transient control target value setting means, and an engine load prediction means, which is predetermined for each of the solenoids 51 to 55 based on the above various control information. The control signal is output and predetermined control is performed, but only the shift control related to the gist of the present application will be described below.

Hereinafter, the control method of the shift control by the control unit CU will be described according to the flowcharts shown in FIGS. 5 and 6 and with reference to FIGS. 2 to 4 as appropriate. First, the main routine shown in FIG. 5 will be described. It should be noted that this main routine is performed every predetermined time (for example, 2
It is repeatedly executed (every 0 ms). When control is started, in step # 1, shift position Range (select range), throttle opening TVO, primary rotation speed Np, secondary rotation speed Ns, engine rotation speed Ne, turbine rotation speed Nt, oil temperature THO, hydraulic pressure. Various signals such as Poil are read as control information. Next, in step # 2, the steady target rotation speed Nps (i) and the transient target rotation speed Npt (i) of the primary pulley 31 are calculated. These calculations are shown in FIG. 6 as will be described later. It is adapted to be carried out in a target speed calculation subroutine shown in the flowchart.

Next, at step # 3, it is determined whether or not the absolute value | Nps (i) -Np | of the deviation of the actual primary rotation speed Np from the steady target rotation speed Nps (i) is not less than the predetermined value C1. It is compared and judged. Here, C1 is CVT1 when the absolute value of the deviation | Nps (i) −Np | is smaller than this.
The boundary value is set so that 0 is considered to be in a substantially steady state and there is no problem.

Then, in step # 3, | Nps (i) -Np |
When it is determined that <C1 (No), the primary rotation speed Np substantially matches the steady target rotation speed Nps (i),
Therefore, since the CVT 10 is considered to be in a substantially steady state, the F / B control (feedback control) of the primary rotation speed Np is performed in steps # 9 and # 10. That is, the target duty pressure Pd is calculated in step # 9,
Subsequently, in step # 10, the duty ratio D corresponding to the target duty pressure Pd is calculated according to the oil temperature THO, the oil pressure Poil, etc. using a predetermined map or table, and the duty ratio D is determined by the first duty solenoid. 51 is applied. Since the F / B control itself is generally known and is not the gist of the present application, a detailed description thereof will be omitted. However, in step # 9, the rotation speed deviation dn is calculated by the following equation 1 and the rotation speed deviation is calculated. The target duty pressure Pd is calculated by a normal PID operation based on dn. Note that after step # 10 is executed, the process returns to step # 1.

[Formula 1] dn = Npt (i) −Np …………………………………………………… Equation 1

On the other hand, in step # 3, │Nps (i) -Np│
If it is determined that ≧ C1 (Yes), it is considered that the CVT 10 is in a transient state (acceleration state or deceleration state), so basically F / F control (feedforward control)
Is performed. Specifically, first, in step # 4, it is compared / determined whether the absolute value | Npt (i) −Np | of the deviation of the primary rotational speed Np from the transient target rotational speed Npt (i) is equal to or greater than a predetermined value C2. To be done. Here, C2 is set to a boundary value such that the absolute value of the deviation │Npt (i) -Np│ cannot be more than this in a normal state (state where the control system is not failed). ing. In step # 4, │Npt
If (i) -Np│ ≧ C2 is determined (Yes), it is considered that a failure has occurred in the control system.
The routines for dealing with the failure in steps # 5, # 6 and # 7 are executed. If F / F control is performed when the control system fails, the deviation of the primary rotation speed Np from the transient target rotation speed Npt increases with the passage of time, as shown in FIG. 10, for example. Therefore, in the present embodiment, when the deviation becomes equal to or larger than the predetermined value C2, the control is forcibly F / F after the predetermined time C3 has elapsed from this point.
The control is switched to the F / B control.

Specifically, in step # 5, the absolute value of the deviation | Npt (i-1) -Np | of the primary rotational speed Np from the previous transient target rotational speed Npt (i-1) is the above-mentioned predetermined value. C2
It is compared and judged whether or not it is less than | Npt (i-1)-
If it is determined that Np│ <C2 (Yes), it means that the absolute value of the deviation becomes C2 or more for the first time this time, and therefore the initial value C3 is set in the timer TM in step # 6. On the other hand, if it is determined that | Npt (i−1) −Np | ≧ C2 (No), the absolute value of the deviation is already C2 or more before the previous time, so the timer TM has already counted. It has started. Therefore, step # 7
Then, it is compared / determined whether the timer TM has not yet counted up, that is, whether TM> 0. Here, when it is determined that TM ≦ 0, (N
o), because the timer TM has already counted up,
Steps # 9 and # 10 are executed and control is forced to F /
The F control is switched to the F / B control.

On the other hand, in step # 4, | Npt (i) -Np |
If it is determined to be <C2 (No), no failure has occurred in the control system, so the F / F control of the primary rotation speed Np is performed in steps # 8 and # 10. That is, the target duty pressure Pd is calculated in step # 8,
Then, in step # 10, the duty ratio D corresponding to the target duty pressure Pd is calculated, and this duty ratio D is applied to the first duty solenoid 51. Since the F / F control itself is generally known and is not the gist of the present application, a detailed description thereof will be omitted. However, in step # 8, the target shift speed Rv is calculated by the following equation 2, and then the target duty is calculated. The pressure Pd is calculated by Equation 3.

## EQU2 ## Rv = Npt (i) / Ns-Npt (i-1) / Ns ………………………… Equation 2

## EQU00003 ## Pd = Gd (Rv, Ps) + Pmd .................. Formula 3 In addition, in Formula 3, Gd (Rv, Ps) is the target shift speed Rv.
And a predetermined function Gd (R with the line pressure Ps as a parameter
v, Ps), and Pmd is the neutral position duty pressure stored while being updated by the learning control.

The target speed calculation subroutine called in step # 2 will be described below with reference to the flowchart shown in FIG. This target rotation speed calculation subroutine is a subroutine for calculating the steady target rotation speed Nps (i) and the transient target rotation speed Npt (i). When the subroutine is called, first in step # 11, the throttle opening degree at the present time is calculated. TVO, that is, the engine load and the rate of change of the throttle opening TVO (engine load) with time dTV
O / dt, that is, the differential value (difference value) of the throttle opening TVO
Based on the above, the throttle opening predicted value TVO '(engine load predicted value) after a predetermined time Td is calculated.

Here, the predetermined time Td is set so as to match the response delay time (hydraulic response delay time) generated by setting the transient target rotation speed Npt (i) (see FIG. 14). That is, as will be described later, the transient target speed Np
t (i) is calculated by performing a smoothing process on the steady target rotation speed Nps (i). For this reason, in the conventional control method, for example, as shown in FIG. 14, the transitional target rotation speed Npt (i) rises late with respect to the change in the throttle opening, and the initial responsiveness in gear shift deteriorates. Therefore, in the present embodiment, the throttle opening after the response delay time due to the smoothing process is predicted, and the steady target rotation speed Nps (i) is calculated based on the predicted throttle opening TVO ′. The above-mentioned response delay is compensated to improve the initial response in shifting. Note that such a decrease in responsiveness is particularly problematic at the initial stage of acceleration or the initial stage of deceleration. Therefore, the use of the throttle opening predicted value TVO ′ is within a predetermined period (eg, within 1 second) after the start of acceleration or the start of deceleration. It may be limited to.

Next, at step # 12, the steady-state target rotation speed Nps (i) is the shift position signal Range detected by the shift position sensor 62, and the throttle opening predicted value TVO 'calculated at step # 11. , And the secondary rotation speed Ns detected by the secondary rotation speed sensor 64, for example, is read from a map as shown in FIG. 9. Then, in step # 13, the gear ratio deviation Δ is the absolute value | Nps of the difference between the current Nps (i) and the previous Npt (i-1).
It is calculated by dividing (i) -Npt (i-1) | by the secondary rotation speed Ns. That is, the gear ratio deviation Δ is obtained in the form of the absolute value of the difference between the steady target gear ratio and the previous transient target gear ratio. By using such a gear ratio deviation Δ, the engine speed and The influence of vehicle speed can be eliminated, and the control calculation is simplified.

Thereafter, at step # 14, the current transient target rotation speed Npt (i) of the primary pulley 31 is calculated. Here, the transient target speed Npt (i) is calculated by the following equation 4 during acceleration (downshift) and by equation 5 during deceleration (upshift).

## EQU00004 ## Npt (i) = min [Npt (i-1) + Ns.FA (.DELTA.), Nps (i)] ......... Equation 4

## EQU00005 ## Npt (i) = max [Npt (i-1) -Ns.FA (.DELTA.), Nps (i)] ..... Equation 5 In equations 4 and 5, min [α, β] means the smaller of α and β, and max [α, β] means the larger of α and β. It should be noted that in Equations 4 and 5, Nps (i) is merely introduced as a countermeasure for erroneous control or a hunting prevention measure, and normally Npt (i) = Npt (i−1) + Ns · FA during acceleration. (Δ), and Npt (i) = Npt (i−1) −Ns · FA (Δ) during deceleration.

In Expression 4 or Expression 5, FA (Δ) at the time of acceleration or deceleration is a function having the gear ratio deviation Δ as a parameter, and is set as shown in FIG. 11 or 12, respectively. Therefore, the current transient target speed Npt (i) is set by the smoothing process based on the previous transient target speed Npt (i-1) and the gear ratio deviation Δ. Then, as is clear from FIG. 11 or FIG. 12, the function FA (Δ) is set to a large value when the gear ratio deviation Δ is large, such as immediately after the transition to the transient state, and the transition to the other transient state. The value is set to an extremely small value when the gear ratio deviation Δ is small, such as when a certain amount of time has elapsed. By setting the function FA (Δ) in this way, during acceleration, for example, as shown in FIG. 7, the transient target speed is gradually increased while the transient target speed is gradually increased with respect to a sudden increase in the slot opening. The number of torques consumed by the rotational inertia of the engine can be reduced to obtain a good feeling of acceleration. Further, during deceleration, for example, as shown in FIG. 8, it is possible to gradually decrease the transitional target rotation speed and match the steady target rotation speed with the rapid decrease in the slot opening degree.

FIG. 13 shows the throttle opening during acceleration when the shift control (FIGS. 5 and 6) is performed.
An example of the characteristics of the target rotation speed, the duty ratio, and the actual primary rotation speed with respect to time is shown, and FIG. 14 shows a similar diagram when the conventional control method is used. As is clear from FIG. 14, when the transient target rotational speed is calculated by the conventional control method, that is, the smoothing process based on the actual throttle opening, the hydraulic response delay Td (t ₀ ~ t ₂ ) and a hydraulic response delay Th (t _{2 to} t ₃ ) caused by other causes, and eventually the change in the primary rotational speed has a response delay of (Td + Th) with respect to the change in the throttle opening. Become. On the other hand, in the control method according to the present invention, that is, when the transient target rotation speed is calculated by the smoothing process based on the throttle opening predicted value after Td, as is clear from FIG. The resulting response delay does not occur, so the response delay is Th (t ₀ ~
Only t ₁ ), and the responsiveness is greatly improved.

[0047]

According to the first aspect of the invention, since the transient control target value is set based on the engine load predicted value after a predetermined time, the transient control target value that anticipates a change in the engine load is set. Is set. Therefore, the response delay to the engine load change of the gear shift operation is reduced, and the initial responsiveness in gear shift is improved.

According to the second invention, basically, the same operation and effect as those of the first invention can be obtained. Furthermore, since the transient control target value is set based on the engine load predicted value after the time corresponding to the response delay time of the control system, the response delay of the control system is compensated and the initial response in shifting is further enhanced. .

According to the third invention, basically, the same operation and effect as those of the first or second invention can be obtained. further,
Response delay is a particular problem.Since the transient control target value is set based on the predicted engine load value only within a predetermined period after the start of acceleration, unnecessary engine load is not predicted and the load on the control mechanism is reduced. To be done.

According to the fourth invention, basically, the same action and effect as any one of the first to third inventions can be obtained. Further, since the next transient control target value is calculated based on the deviation between the normal control target value and the transient control target value, the transient control target value gradually changes during the transition, and the normal control is soon achieved. Reach the target value. Therefore, the transition from the transient state to the steady state is facilitated.

[Brief description of drawings]

FIG. 1 is a block diagram showing a configuration of first to fourth inventions corresponding to claims 1 to 4.

FIG. 2 is a system configuration diagram of a power train including a gear ratio control device according to the present invention.

FIG. 3 is a system configuration diagram of a hydraulic mechanism.

FIG. 4 is a system configuration diagram of a transmission control mechanism.

FIG. 5 is a flowchart showing a control method of gear ratio control.

FIG. 6 is a flowchart of a target rotation speed calculation subroutine.

FIG. 7 is a diagram showing characteristics of a throttle opening degree and a target rotation speed during acceleration with respect to time.

FIG. 8 is a diagram showing characteristics of a throttle opening and a target rotational speed during deceleration with respect to time.

FIG. 9 is a diagram showing characteristics of a primary rotation speed with respect to a secondary rotation speed and a throttle opening.

FIG. 10 is a diagram showing characteristics of a timer value and a primary rotation speed with respect to time.

FIG. 11 is a diagram showing characteristics of a function FA (Δ) for acceleration.

FIG. 12 is a diagram showing characteristics of a function FA (Δ) for deceleration.

FIG. 13 is a diagram showing a control characteristic in the shift control according to the present invention.

FIG. 14 is a diagram showing control characteristics in conventional shift control.

[Explanation of symbols]

 CE ... Engine CT ... Transmission CU ... Control unit 10 ... Belt type continuously variable transmission (CVT) 31 ... Primary pulley 32 ... Secondary pulley 33 ... V belt

Continuation of the front page (51) Int.Cl. ⁵ Identification code Office reference number FI technical display area F16H 59:42 8009-3J

Claims (4)

[Claims] 1. A control target value setting means for setting a control target value of a predetermined physical quantity indicating a shift characteristic based on an engine load, and a shift control means for controlling the physical quantity so as to follow the control target value. Acceleration state detecting means for detecting whether the engine is in a predetermined acceleration state, and when the acceleration state detecting means detects that the engine is in the predetermined acceleration state, it is set by the control target value setting means. In a transmission ratio control device for a continuously variable transmission, which is provided with a transient control target value setting means for setting a transient control target value whose variation amount is smaller than the control target value, a predetermined time is determined based on the rate of change of the engine load. The engine load predicting means for calculating the engine load predicting value after that is provided, and the transient control target value setting means is an engine load predicting means for calculating the engine load predicting value. Gear ratio control device for a continuously variable transmission, characterized in that is adapted to set the transient control target value based on the down load prediction value. 2. The gear ratio control device for a continuously variable transmission according to claim 1, wherein the engine load predicting means determines the engine load predictive value after a time corresponding to the response delay of the transient control target value setting means. A gear ratio control device for a continuously variable transmission, characterized in that it is adapted to perform calculations. 3. The gear ratio control device for a continuously variable transmission according to claim 1 or 2, wherein the transient control target value setting means is based on the engine load predicted value only within a predetermined period after the start of acceleration. A transmission ratio control device for a continuously variable transmission, characterized in that a transient control target value is set. 4. The gear ratio control device for a continuously variable transmission according to claim 1, wherein the transient control target value setting means is a transient control target value and a control target value setting means. A transmission ratio control device for a continuously variable transmission, characterized in that the next transient control target value is set based on the deviation from the control target value set by.