JP4479332B2 - Control device for continuously variable transmission - Google Patents

Description

  The present invention relates to a continuously variable transmission that changes speed by continuously changing a belt wrapping radius between a driving pulley and a driven pulley, and is particularly configured to control a transmission ratio and a belt clamping pressure by hydraulic pressure. The present invention relates to a control device for a continuously variable transmission.

  In a belt-type continuously variable transmission, each of a drive pulley and a driven pulley is constituted by a fixed sheave and a movable sheave that approaches and separates from the fixed sheave, and a groove formed between the fixed sheave and the movable sheave. The belt is wound and the belt is sandwiched between a fixed sheave and a movable sheave to set a predetermined transmission torque capacity. Conventionally, a hydraulic pressure for setting a belt clamping pressure (hereinafter simply referred to as a clamping pressure) determined based on an input torque and a gear ratio is supplied to a driven pulley. An apparatus having a configuration in which it is set by changing the hydraulic pressure is known.

  Therefore, since the gear ratio is determined according to the relative position between the fixed sheave and the movable sheave, conventionally, the gear shift control is performed by controlling the flow rate of the pressure oil with respect to the drive pulley. Technology to do it has been developed. In this case, the clamping pressure is controlled based on the input torque or the throttle opening degree related thereto, so that the gear ratio setting hydraulic pressure is controlled in accordance with the clamping pressure to improve the followability to the target gear ratio. The apparatus which was made is described in patent document 1. FIG.

The device described in Patent Document 1 obtains the actual gear ratio from the rotation speed of each pulley, and obtains the target gear ratio from the engine load and the rotation speed so that the actual gear ratio matches the target gear ratio. The hydraulic pressure of the primary pulley (for example, drive pulley) is controlled, and the clamping pressure corresponding to the input torque is controlled by the hydraulic pressure supplied to the secondary pulley (for example, driven pulley). The primary side hydraulic pressure is controlled together with the secondary side hydraulic pressure control. Then, the gear ratio control means multiplies the deviation between the target gear ratio and the actual gear ratio by a predetermined gain, integrates the value divided by the variable corresponding to the secondary hydraulic pressure, and adds the above variable to the integral value. Integral calculation means for outputting the multiplied value is provided. According to the invention described in Patent Document 1, even when the input torque is increased by the driver's accelerator operation and the secondary hydraulic pressure is increased accordingly, the gear ratio is hardly changed. The followability to the target value is improved.
JP-A-8-326857

  The device described in Patent Document 1 is configured to determine the feedback pressure by PID control, and particularly improves the stability of the transmission ratio when the pinching pressure changes according to the change of the input torque. It is configured as follows. For this reason, the accumulation of integral terms is large, and this may be a cause of a decrease in responsiveness at the time of shifting. That is, in the invention of the above-mentioned Patent Document 1, when the target gear ratio changes stepwise due to some factor and the deviation from the actual gear ratio increases, the deviation is gradually accumulated, and the actual gear ratio becomes the target gear ratio. There is a possibility that the feedback pressure at the time of approaching increases, and this causes an increase in the shift control amount and an overshoot of the actual gear ratio. On the other hand, when the torque and the rotational speed temporarily change due to disturbance, and this appears as a shift, in the invention of Patent Document 1, the feedback pressure is relatively insufficient until the deviation becomes large to some extent. The return to the target gear ratio is delayed, and as a result, drivability may deteriorate.

  The present invention has been made paying attention to the above technical problem, and in controlling the speed ratio, prevents overshoot of the actual speed ratio with respect to the target speed ratio, and is excellent in stability against disturbance. An object of the present invention is to provide a control device for a step transmission.

In order to achieve the above object, the invention of claim 1 has a drive pulley and a driven pulley capable of changing the groove width around which the belt is wound by hydraulic pressure, and by hydraulic pressure supplied to one pulley. In a control device for a continuously variable transmission that controls a transmission ratio and controls a clamping pressure of the belt by a hydraulic pressure supplied to the other pulley, an axial load in which the driving pulley clamps the belt and a driven pulley are Means for obtaining a steady travel-required hydraulic pressure to be supplied to the one pulley in a normal state where no speed change occurs based on a thrust ratio that is a ratio of an axial load that sandwiches the belt and the clamping pressure; the shifting time required oil pressure corresponding to the target shift speed for shifting the gear ratio to a target, means for determining based on a predetermined coefficient, the rate of change in the actual speed ratio and the target speed change rate Means for determining a feedback pressure by multiplying the inverse of the coefficient of the deviation, means for determining the hydraulic pressure supplied to the one of the pulleys from the previous SL feedback pressure in addition to the steady running necessary hydraulic and the transmission time required oil pressure It is provided with the control apparatus characterized by these.

According to a second aspect of the present invention, in the first aspect of the invention, an upper limit guard for restricting an upper limit of the target shift speed and a lower limit guard for restricting a lower limit of the target shift speed are set. This is a control device characterized by that.
Further, the feedback of the invention of claim 3, in the invention of claim 1 or 2, further comprising means for filtering the pre-Symbol feedback pressure, means for determining the hydraulic pressure supplied to said one pulley, that is the filter processing The control apparatus is configured to obtain a hydraulic pressure to be supplied to the one pulley from a hydraulic pressure, the steady traveling necessary hydraulic pressure, and the shifting necessary hydraulic pressure.

  According to the first aspect of the present invention, the hydraulic pressure for maintaining the gear ratio during normal traveling without shifting, that is, the hydraulic pressure of the one pulley (steady traveling hydraulic pressure) is based on the thrust ratio and the clamping pressure. Is required. The thrust ratio can be determined in advance. Further, the hydraulic pressure required for shifting to change the gear ratio to the target gear ratio is obtained based on the gear speed and a predetermined coefficient for achieving the target gear ratio. The coefficient, or the relationship between the required hydraulic pressure during shift, the coefficient, and the target shift speed can be obtained in advance based on experiments or the like. These relationships also hold between the deviation or error between the target shift speed and the actual shift speed and the hydraulic pressure for solving this, so that the feedback hydraulic pressure is obtained by multiplying the shift speed deviation by the reciprocal of the coefficient. Desired. In addition, since the hydraulic pressure for the one pulley for setting the gear ratio is obtained based on the steady travel required hydraulic pressure, the shift required hydraulic pressure, and the feedback hydraulic pressure, the target gear ratio is changed stepwise. However, it is possible to prevent or suppress the actual gear ratio from overshooting or the delay in following the target gear ratio.

According to the invention of claims 2 and 3, it is possible to obtain the same effect as the invention of claim 1, particularly according to the invention of claim 3, it improves the stability of the speed change ratio control In addition, the durability of the hydraulic control device can be improved. That is, the feedback hydraulic pressure reflects the shift speed, which is the amount of change in the gear ratio per unit time, but the feedback hydraulic pressure corresponding to the change due to the disturbance in the gear ratio is removed by the filter, so that the shift control by the disturbance is performed. Frequent occurrences are avoided.

  Next, the present invention will be described based on specific examples. First, a continuously variable transmission 1 and its hydraulic control system to which the present invention can be applied will be described. FIG. 6 schematically shows a basic configuration of the belt-type continuously variable transmission 1, in which the driving pulley 2 and the driven pulley 3 are arranged at predetermined intervals with their central axes parallel to each other. ing. The drive pulley 2 can change the width of a so-called V-groove around which the belt 4 is wound. For this reason, the drive pulley 2 moves back and forth in the axial direction with a fixed sheave (fixed pulley piece) 5 fixed in the axial direction. And a movable sheave (movable pulley piece) 6 that approaches and separates from the fixed sheave 5. A hydraulic actuator (specifically a hydraulic cylinder) 7 for moving the movable sheave 6 back and forth is provided on the back side of the movable sheave 6 (opposite to the fixed sheave 5).

  The driven pulley 3 is basically the same as the driving pulley 2 except that the positions of the fixed sheave and the movable sheave are opposite to the driving pulley 2 described above, and are opposed to each other on the same axis. A fixed sheave 8 and a movable sheave 9 that moves back and forth with respect to the fixed sheave 8. The width of a so-called V-groove formed between the opposed surfaces of the fixed sheave 8 and the movable sheave 9 is changed in a wide and narrow manner. The wrapping radius of the belt 4 is changed. A movable sheave 9 on the outer driven pulley 3 is arranged in the radial direction of the fixed sheave 5 in the driving pulley 2, and a fixed sheave 8 on the outer driven pulley 3 is arranged in the radial direction of the movable sheave 6 in the driving pulley 2. ing. Further, a hydraulic actuator (specifically, a hydraulic cylinder) 10 for moving the movable sheave 9 back and forth is provided on the back side of the movable sheave 9 (on the opposite side to the fixed sheave 8).

  The continuously variable transmission 1 can be employed as a transmission for a vehicle. Therefore, the drive pulley 2 is connected to a power source (not shown) such as an internal combustion engine or an electric motor via a starting clutch or a torque converter. Z). The driven pulley 3 is connected to driving wheels (not shown) via an output shaft, a differential, a propeller shaft, or the like.

  In the belt 4, a large number of metal pieces (sometimes referred to as blocks) sandwiched between the V grooves of the pulleys 2 and 3 are arranged in a ring shape, and these metal pieces are referred to as hoops. It is formed by binding with a band. Therefore, since the entire length is limited by the hoop, when the belt 4 is sandwiched between the pulleys 2 and 3, a force is applied to push the belt 4 outward in the radial direction by the inclined surface of the V groove. As a result, the belt 4 is applied with tension, and a contact pressure between the belt 4 and the pulleys 2 and 3 is generated. A torque is generated between the belt 4 and the pulleys 2 and 3 by a frictional force determined by the contact pressure and a friction coefficient. Communicated. Thus, the pressure which pinches | interposes the belt 4 is a clamping pressure, Comprising: It sets with the hydraulic pressure supplied to the said driven pulley 3, for example.

  On the other hand, when the pressure sandwiching the belt 4 in either one of the pulleys 2 and 3 is relatively increased or decreased, the belt 4 has a radius of the pulleys 2 and 3 against the tension of the belt 4. The belt 4 is pushed inward in the radial direction or pushed outward in the radial direction at the same time in the other pulleys 3, 2. Such a change in the winding radius is execution of the speed change, which is executed by, for example, hydraulic pressure supplied to the drive pulley 2.

  That is, the continuously variable transmission 1 is configured to perform the clamping pressure control and the shift control by pressure control of the hydraulic pressure supplied to the pulleys 2 and 3. Briefly describing the hydraulic control system for that purpose, a pressure regulating valve 11 for regulating the hydraulic pressure supplied to the drive pulley 2 is provided. The pressure regulating valve 11 is a spool type valve, and a spring 12 that presses a spool (not shown) in the axial direction thereof is disposed at one end of the spool and is pressed in the same direction as the pressing direction by the spring 12. A control port 13 for supplying a signal pressure so as to apply pressure is provided, and an electromagnetic valve 14 is communicated with the control port 13.

  Also, the line pressure PL, which is the original pressure of the entire hydraulic system, is supplied to the input port 15 that is opened and closed by the spool, and when the spool moves in the axial direction, the input port 15 is connected to the output port 16 and the drain port. The communication is switched to 17. Further, a feedback port 18 is formed on the opposite side of the control port 13 in the axial direction of the spool. The hydraulic pressure of the output port 16 is applied to the feedback port 18 via the orifice 19.

  Therefore, the pressure regulating valve 11 balances the axial pressing force against the spool due to the spring 12 and the signal pressure supplied to the control port 13 and the axial pressing force against the spool due to the output pressure acting on the feedback port 18. In this way, the output pressure is regulated. In order to supply this output pressure to the drive pulley 2, the output port 16 communicates with the hydraulic cylinder 7 in the drive pulley 2. The solenoid valve 14 is for outputting a signal pressure applied to the control port 13, and uses the line pressure PL or a solenoid modulator pressure obtained by adjusting the line pressure PL as a base pressure in accordance with a duty ratio or a current value. This is a known solenoid valve that outputs hydraulic pressure. In addition, a hydraulic pressure sensor 20 that detects the hydraulic pressure in the hydraulic cylinder 7 on the drive pulley 2 side is provided.

  The hydraulic system for setting the clamping pressure by controlling the hydraulic pressure supplied to the hydraulic cylinder 10 of the driven pulley 3 has substantially the same configuration as the control system for controlling the hydraulic pressure supplied to the drive pulley 2 described above. A pressure regulating valve 21 that outputs a hydraulic pressure corresponding to the pressure is provided. The pressure regulating valve 21 has a spring 22 and a control port 23 as in the pressure regulating valve 11 described above. A solenoid valve 24 similar to the valve 14 is communicated. Further, the pressure regulating valve 21 includes an input port 25 to which the line pressure PL is supplied, an output port 26, a drain port 27, and a feedback port 28 communicated with the output port 26 via an orifice 29. A hydraulic sensor 30 that detects the hydraulic pressure in the hydraulic cylinder 10 of the driven pulley 3 is provided. An electronic control unit (ECU) 31 is provided that appropriately outputs a command signal to each of the electromagnetic valves 14 and 24 to control the hydraulic pressure on the driving pulley 2 side and the hydraulic pressure on the driven pulley 3 side.

  The hydraulic pressure of the hydraulic cylinder 7 in the drive pulley 2 is controlled to a pressure for setting a predetermined gear ratio. The speed ratio is basically a speed ratio determined in advance according to the running state of the vehicle. For example, when an internal combustion engine is used as a power source, the speed at which the fuel consumption is minimized is set. The transient hydraulic pressure is controlled so as to execute a shift at a speed satisfying the acceleration / deceleration request during acceleration / deceleration. This hydraulic pressure control is executed by, for example, feedforward feedback control. On the other hand, the hydraulic pressure in the hydraulic cylinder 10 of the driven pulley 3 is for setting a clamping pressure, and is therefore controlled mainly according to the input torque. The control is basically feedback control.

  FIG. 1 is a block diagram showing an example of control executed by the control device of the present invention, and is a control block diagram for mainly setting the predetermined gear ratio by controlling the hydraulic pressure on the drive pulley 2 side. When setting the gear ratio by hydraulic pressure control, first, the steady-state hydraulic pressure Pinc is calculated (block B1). In the example shown in FIG. 6, the gear ratio is controlled by the hydraulic pressure on the drive pulley (primary pulley) 2 side, and therefore, “steady-state primary pulley hydraulic pressure” is described in FIG.

FIG. 2 shows a flowchart for calculating the steady-state hydraulic pressure Pinc. In this routine, first, an axial load by the driven pulley 3, that is, an actual secondary pulley thrust WOUT is calculated (step S0011). Since the thrust WOUT is generated by the hydraulic cylinder 10 on the driven pulley 3 side,
WOUT = (POUT + PSCH) × AOUT
Is calculated by Here, POUT is the hydraulic pressure in the hydraulic cylinder 10 on the driven pulley 3 side, PSCH is the centrifugal hydraulic pressure in the hydraulic cylinder 10, and AOUT is the pressure receiving area of the piston in the hydraulic cylinder 10.

Further, the thrust to be generated in the driven pulley 3, that is, the theoretical secondary pulley thrust WOUTMIN is calculated (step S0012). As described above, in the hydraulic cylinder 10 on the driven pulley 3 side, thrust is generated so as to set a necessary and sufficient clamping force to transmit the input torque without causing slippage, but the actual thrust is Since some safety factor (safety factor SF) is expected, in step S0012, a thrust of SF = 1 is calculated. In particular,
WOUTMIN = TT × cos α / (2 × μ × RIN)
Is calculated by Here, TT is an estimated input torque, and can be obtained based on a map using the throttle opening as a parameter when, for example, an internal combustion engine is used as a power source. α is the included angle of the belt 4 in the drive pulley 2, μ is the coefficient of friction between the belt 4 and the drive pulley 2, and RIN is the winding radius of the belt 4 in the drive pulley 2. Since the winding radius RIN is a value that is geometrically determined by the gear ratio, it may be prepared in advance as a gear ratio map and read from the map.

A safety factor SF is obtained from the actual thrust WOUT and the theoretical thrust WOUTMIN in the driven pulley 3 (step S0013).
SF = WOUT / WOUTMIN

  Furthermore, the steady travel thrust ratio τ is calculated (step S0014). This steady-state driving thrust ratio τ is determined by the driving pulley 2 for driving the continuously variable transmission 1 at a predetermined gear ratio (ratio between the rotational speed NIN of the driving pulley 2 and the rotational speed NOUT of the driven pulley 3) γ. This is the ratio of the thrust WIN and the thrust WOUT at the driven pulley 3, and can be obtained in advance as a map according to the gear ratio γ and the input torque. Note that a map in which the safety factor SF is further added as a parameter may be used. In step S0014, the thrust ratio τ can be calculated using the map. FIG. 3 conceptually shows the relationship between the input torque TT and the thrust ratio τ at a predetermined gear ratio γ and safety factor SF by a diagram.

The thrust WIN in the drive pulley 2 is the product of the hydraulic pressure Pinc in the hydraulic cylinder 7 and the pressure receiving area Ain. Therefore, the hydraulic pressure of the drive pulley 2 during steady running (steady-time primary pulley hydraulic pressure) Pinc is
Pinc = WOUT / (τ × Ain)
(Step S0015).

On the other hand, when shifting is performed, the thrust WIN at the driving pulley 2 is changed, and accordingly, the groove width at the driving pulley 2 becomes narrower (the winding radius increases) and the groove width at the driven pulley 3 becomes smaller. wider (by winding radius becomes smaller) upshift occurs, or groove width of the drive pulley 2 is widened in the opposite (winding radius decreases) narrows the groove width at the driven pulley 3 with A downshift occurs (with increased wrapping radius). Accordingly, the hydraulic pressure (primary pulley hydraulic pressure) Pins of the drive pulley 2 side hydraulic cylinder 7 by the shift is calculated (block B2). A routine for the calculation is shown in the flowchart of FIG.

First, the time change dgdt of the gear ratio γ is calculated (step S0021). Between change dgdt time of the gear ratio γ of this is,
dgdt = (γtgt (i) −γtgt (i-1)) / ΔT
It is. Here, ΔT is a sampling time, and γtgt is a target speed ratio, which is a ratio (NINT / NOUT) between a target input rotational speed (target drive pulley rotational speed) NINT and the rotational speed NOUT of the driven pulley 3.

The time change dgdt of the speed ratio γ thus obtained is proportional to the hydraulic pressure Pins of the driving pulley 2 due to the speed change, and the proportional constant K is the actual speed ratio γreal, the input rotational speed NIN, and the hydraulic cylinder 10 on the driven pulley 3 side. It is known that it depends on the hydraulic pressure POUT and the type of upshift / downshift. That is, the proportionality constant K can be prepared in advance as data or a map using these as parameters. A proportionality constant K is calculated from the data or map (step S0022). Then, the hydraulic pressure Pins of the drive pulley 2 side hydraulic cylinder 7 due to the speed change is calculated from the proportionality constant K and the time change dgdt of the speed ratio γ by the following equation (step S0023).
Pins = dgdt / K

Further, an upper limit guard and a lower limit guard are applied to the hydraulic pressure Pins of the drive pulley 2 side hydraulic cylinder 7 by the speed change (step S0024). As shown in FIG. 1, the oil pressure Pinc of the driving pulley 2 at the time of steady running described above and the oil pressure Pins of the driving pulley 2 by shifting are added to obtain the final feedforward oil pressure Pinff, which is the final feedforward oil pressure Pinff. Since it is restricted to the upper limit value in terms of mechanism or control, the hydraulic pressure Pins of the drive pulley 2 side hydraulic cylinder 7 is limited by shifting so as not to exceed the upper limit value. On the other hand, in order to avoid slipping of the belt 4 caused by excessively lowering the final feedforward hydraulic pressure Pinff, the drive pulley 2 side hydraulic cylinder 7 by shifting so that the final feedforward hydraulic pressure Pinff maintains a pressure higher than the lower limit value Ping. There is a restriction on the hydraulic pressure Pins. The lower limit value Ping corresponds to the hydraulic pressure in the drive pulley 2 side hydraulic cylinder 7 that generates the theoretical secondary pulley thrust WOUTMIN, and the theoretical secondary pulley thrust WOUTMIN is expressed by the area AIN of the piston of the drive pulley 2 side hydraulic cylinder 7. Divided value.
Ping = TT × cos α / (2 × μ × RIN × AIN)

The control device according to the present invention for controlling the hydraulic pressure via the solenoid valves 14 and 24 shown in FIG. 6 further converts the final feedforward hydraulic pressure Pinff obtained as described above with the feedback hydraulic pressure based on the shift speed deviation. It is configured to correct. A portion corresponding to the control is a portion surrounded by a broken line in FIG. 1. First, the proportional speed K is multiplied by the hydraulic pressure Pins of the driving pulley 2 by the above-described shift to obtain the target shift speed dγtgt / dt (block). 3). As the hydraulic pressure Pins, the value to which the upper limit guard or the lower limit guard in step S0024 described above is applied may be used. In this case, the time change of the speed ratio γ calculated in step S0021 described above may be used. dgdt that is, it does not coincide with the goals shift speed.

On the other hand, the gear ratio γ (i) is continuously monitored every predetermined sample time ΔT, and the actual gear ratio changing speed dγreal / dt is obtained by approximation of the difference (block B4).
dγreal / dt = (γ (i) −γ (i−1)) / ΔT

If there is a deviation dγerr / dt between the target transmission speed dγtgt / dt and the actual transmission speed dγreal / dt, the actual transmission ratio γreal may be delayed in convergence to the target transmission ratio γtgt, or the actual transmission ratio γreal may be exceeded. There is a possibility of shooting. Therefore, a hydraulic pressure Pinfb necessary and sufficient to remove the shift speed deviation dγerr / dt is obtained (block B5). Specifically, by utilizing the relationship that the hydraulic pressure Pins of the drive pulley 2 side hydraulic cylinder 7 and the target shift speed dγtgt / dt are proportional to each other with a proportional constant K, the shift speed is set with the proportional constant K. By dividing the deviation dγerr / dt (accumulating the reciprocal of the proportionality constant K), the feedback hydraulic pressure Pinfb is obtained.
Pinfb = 1 / K × dγerr / dt

In the block B4, the actual speed ratio γreal is substantially differentiated to obtain the actual speed. For this reason, when there is a disturbance such as a temporary change in torque input from the driving wheels of the vehicle or an erroneous signal from the rotation speed sensor, this may be detected as a shift at a relatively high speed. In order to prevent this, a filter process for removing a disturbance component from the feedback hydraulic pressure Pinfb is performed (block B6). Specifically, this filter process is a low-pass filter process. Instead of this, the actual shift speed may be low-pass filtered.

  The feedback hydraulic pressure Pinfb is added to the final feedforward hydraulic pressure Pinff obtained from the steady-state hydraulic pressure Pinc and the shift hydraulic pressure Pins to determine the hydraulic pressure on the driving pulley 2 or its command value Pint, or the feedback hydraulic pressure Pinfb The low-pass filter processing value Pinfb 'is added to determine the hydraulic pressure on the driving pulley 2 side or its command value Pint (block B7).

  The operation of the control device according to the present invention will be described in comparison with a conventional example. In FIG. 5, the hydraulic pressure command value of the hydraulic cylinder 7 on the drive pulley 2 side is changed so as to change the speed ratio stepwise, and then the hydraulic cylinder 7 on the drive pulley 2 side is changed gradually so that the speed ratio is changed. It shows the change in the input rotational speed when a disturbance occurs in the process of changing the hydraulic pressure command value. FIG. 5A shows a conventional example in which feedback hydraulic pressure is determined by performing PID control on the position of the movable pulley constituting the driving pulley or the deviation amount between the actual value of the rotational speed of the driving pulley and the target value. As the gear ratio command value increases stepwise at time t1, the input rotational speed increases, but an overshoot occurs in which the actual input rotational speed is greater than the target input rotational speed. Also, when the gear ratio command value changes due to disturbance at the time t2, it vibrates with the actual input speed exceeding the target input speed and does not follow the change in the target input speed for a predetermined time. Yes.

  FIG. 5B is an example in the case where control is performed by the control device of the present invention, and is an example in which the low-pass filter processing in block B6 described above is not performed. The change in the gear ratio command value is the same as in the above-described conventional example, and is changed stepwise at the time t1, and is changed due to disturbance at the time t2. Since the feedback hydraulic pressure is not filtered, the high pressure component is included in the hydraulic pressure command value of the drive pulley side hydraulic cylinder, but the actual input speed when the gear ratio is changed stepwise at the time t1 Has converged to the target input speed without overshooting the target input speed. Even when there is a disturbance at the time t2, the actual input rotational speed changes following the target input rotational speed without greatly deviating from the target input rotational speed. Such a state is the same in the example shown in FIG. 5C in which the low-pass filter processing for the feedback hydraulic pressure is performed.

  The example shown in FIG. 5 is an example of a downshift that increases the gear ratio (input rotation speed), but the same effect occurs even in the case of an upshift. Therefore, the control device according to the present invention can prevent or suppress the overshoot of the actual input rotation speed or the actual gear ratio when the gear ratio is changed stepwise, and is excellent in stability against disturbances. When the continuously variable transmission is a target, it is possible to prevent the deterioration of the drivability of the vehicle or to improve the drivability.

  Here, the relationship between the above specific example and the present invention will be briefly described. The functional means for executing the control shown by the block B1 in FIG. 1 and shown by the flowchart in FIG. The functional means corresponding to “means for obtaining” and executing the control indicated by the block B2 in FIG. 1 and shown by the flowchart in FIG. The functional means for executing the control indicated by the blocks B3 to B5 or the control obtained by adding the control of the block B6 corresponds to the “means for obtaining the feedback hydraulic pressure” in the present invention, and the final feedforward hydraulic pressure Pinff and the feedback In the hydraulic cylinder on the drive pulley side from the oil pressure Pinfb or the oil pressure Pinfb 'that has been low-pass filtered. Functional means for obtaining the hydraulic pressure command value corresponds to the "means for determining the hydraulic pressure supplied to the pulley" of the present invention.

  The present invention is not limited to the above specific example. In the example shown in FIG. 1, the feedback hydraulic pressure that has undergone low-pass filtering is configured to be added to the final feedforward hydraulic pressure. You may comprise so that it may add to final feedforward oil_pressure | hydraulic. Further, according to the present invention, if the transmission ratio is controlled by one of the hydraulic actuator on the driving pulley side and the hydraulic actuator on the driven pulley side, and the clamping pressure is controlled by the other hydraulic pressure. Good.

It is a control block diagram for demonstrating an example of the control apparatus of this invention. It is a figure which shows an example of the flowchart for calculating | requiring the steady driving | running | working hydraulic pressure. It is a figure which shows notionally the relationship between estimated input torque and thrust ratio. It is a figure which shows an example of the flowchart for calculating | requiring the hydraulic pressure by the side of the drive pulley by gear shifting. It is a diagram which shows the change by the control apparatus of this invention, and the prior art example about the change of the input rotation speed when a gear ratio changes stepwise and when there is a disturbance. It is a figure which shows typically the continuously variable transmission made into object by this invention, and its hydraulic system.

Explanation of symbols

  DESCRIPTION OF SYMBOLS 1 ... Continuously variable transmission, 2 ... Drive pulley, 3 ... Driven pulley, 4 ... Belt, 11, 21 ... Pressure regulating valve, 14, 24 ... Electromagnetic valve, 31 ... Electronic control apparatus.

Claims (3)

A drive pulley and a driven pulley that can change the width of the groove around which the belt is wound by hydraulic pressure, the transmission ratio is controlled by hydraulic pressure supplied to one pulley, and the hydraulic pressure supplied to the other pulley In the control device for a continuously variable transmission that controls the clamping pressure of the belt,
Based on the thrust ratio, which is the ratio of the axial load with which the drive pulley pinches the belt and the axial load with which the driven pulley pinches the belt, and the pinching pressure, normal time when no shift occurs Means for obtaining a hydraulic pressure required for steady running to be supplied to the one pulley of
Means for obtaining a required hydraulic pressure at the time of shifting corresponding to a target shift speed for shifting to a target gear ratio based on a predetermined coefficient;
It means for determining a feedback pressure by multiplying the inverse of the coefficient of a deviation between the change rate of the actual gear ratio and the target speed change rate,
Control device for a continuously variable transmission, characterized in that it comprises a means for obtaining the hydraulic pressure supplied to the one of the pulleys from the previous SL feedback pressure in addition to the shift when required oil pressure and the steady-state running necessary hydraulic pressure. And the upper limit guard for restricting the upper limit of the target shift speed, continuously variable transmission according to claim 1, characterized in that you have been configured to set the lower limit guard for restricting the lower limit of the target shift speed Control device. Means for filtering the feedback oil pressure;
  The means for determining the hydraulic pressure to be supplied to the one pulley is configured to determine the hydraulic pressure to be supplied to the one pulley from the filtered feedback hydraulic pressure, the steady travel required hydraulic pressure, and the shift required hydraulic pressure.
The control device for a continuously variable transmission according to claim 1 or 2.

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