JP2014152889A - Control device of continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a control device of a continuously variable transmission which can properly obtain an integration term in the gear change feedback control of the continuously variable transmission.SOLUTION: When an absolute value (|dlintgt|) of a target sheave position change amount exceeds a prescribed value A1, and when an absolute value (|lintgt-linact|) of a value which is obtained by subtracting an actual sheave position from a target sheave position exceeds a prescribed value B1, the hold of the sheave position is executed, and an integration term hold execution flag XFBIHLD is turned on. When an absolute value (|dlinact|) of an actual sheave position change amount is smaller than a prescribed value A2 in an integration term hold state, and when a time during which the integration term hold execution flag XFBIHLD is kept to be turned on continues for a prescribed time or longer, the hold of the integration term is released, and the integration term hold execution flag XFBIHLD is turned off.

Description

  The present invention relates to a control device for a continuously variable transmission. In particular, the present invention relates to an improvement in shift feedback control.

  Conventionally, as disclosed in, for example, Patent Document 1 and Patent Document 2, in a belt-type continuously variable transmission (CVT), in order to obtain a target gear ratio, a hydraulic actuator of a movable sheave A target value is set for the hydraulic pressure supplied to the hydraulic pressure, and hydraulic control is performed by feedback control based on a deviation between the target hydraulic pressure and the actual hydraulic pressure. As this feedback control, PI control and PID control are widely known.

  In the PI control and PID control, since an integral term is included, the control can be stabilized by taking into account the influence of past deviations. On the other hand, the hydraulic oil cannot be supplied promptly to the hydraulic actuator. Thus, when the actual oil pressure does not reach the target oil pressure and the deviation is accumulated and then the actual oil pressure reaches the target oil pressure, an overshoot or undershoot of the oil pressure occurs.

  In view of this point, Patent Document 1 discloses that the integral gain of the feedback control is set to be small when the deviation between the target speed ratio and the actual speed ratio exceeds a predetermined value. Patent Document 2 discloses that the accumulation of the integral term is interrupted when the shift speed is equal to or higher than a predetermined value. This suppresses excessive accumulation of integral terms.

JP 2007-132420 A Japanese Patent Laying-Open No. 2005-257067

  However, in the techniques of each patent document, when the release timing of the operation for suppressing the excessive accumulation of the integral term (the operation for interrupting the accumulation of the integral term) is not properly obtained, the overshoot of the hydraulic pressure is performed. And undershoot may occur.

  In other words, if the operation for suppressing the excessive accumulation of the integral term is canceled immediately after the start of the gear shift, the target value and the actual value of the gear ratio are caused by the existence of the hydraulic response delay period immediately after the start of the gear shift. Accumulation of the integral term is started in a state where the deviation from is large. For this reason, there is a possibility that the integral term is accumulated excessively and the hydraulic overshoot or undershoot occurs.

  The present invention has been made in view of such points, and an object of the present invention is to provide a control device for a continuously variable transmission that can appropriately obtain an integral term in the shift feedback control of the continuously variable transmission. There is to do.

-Solution principle of the invention-
The solution principle of the present invention taken to achieve the above object is that when the integral term is fixed and the integral term is fixed in the shift feedback control, the integral term is fixed for a predetermined time or more. By continuing, the integration term is prevented from being accumulated in a state where the deviation between the target speed ratio and the actual speed ratio is increased due to the response delay.

-Solution-
Specifically, the present invention presupposes a control device for a continuously variable transmission that performs feedback control so that the actual gear ratio follows the target gear ratio. When the change amount of the target gear ratio is equal to or greater than a predetermined value and the integral term fixing condition is satisfied for this continuously variable transmission control device, the deviation between the actual gear ratio and the target gear ratio is integrated. The integral term of the proportional feedback control is fixed, and the integral term is fixed when the change amount of the actual gear ratio is less than a predetermined value and the time during which the integral term is fixed is continued for a predetermined time or more. The release condition is satisfied and the integral term is released from being fixed.

  After fixing the integral term because the integral term fixing condition is satisfied due to this specific matter, the integration term is fixed unless the integral term fixed release condition is satisfied for a predetermined time or longer after the integration term is fixed. The fixation of is maintained. In other words, the time for fixing the integral term is secured for a predetermined time or more, and accumulation of the integral term is started in a state where the actual gear ratio is close to the target gear ratio (for example, the response delay is eliminated). Yes. Thereby, it is possible to prevent the accumulation of the integral term from being started in a state where the deviation between the actual speed ratio and the target speed ratio is large, and it is possible to prevent overshoot and undershoot of the speed ratio.

  Specifically, as the integral term fixed release condition, when the physical quantity correlated with the change amount of the actual gear ratio is equal to or less than a predetermined value and the time during which the integral term is fixed is continued for a predetermined time or more. It shall be established.

  Here, as the physical quantity correlated with the change amount of the actual gear ratio, specifically, the change amount of the actual gear ratio can be mentioned. In particular, in the case of a belt-type continuously variable transmission, examples include an actual sheave position change amount (actual sheave position change speed), and a sheave (particularly, primary sheave) rotational speed change amount. When the physical quantity correlated with the change amount of the actual gear ratio is equal to or less than the predetermined value, even if an error occurs in the feedforward control, the time during which the integral term is fixed is the predetermined time. By continuing the calculation (accumulation) of the integral term after the above is continued, the actual gear ratio can be brought close to the target gear ratio without causing overshoot or undershoot.

  Another example of the integral term fixed release condition is that the physical quantity correlated with the deviation between the actual gear ratio and the target gear ratio is equal to or less than a predetermined value.

  Here, as a physical quantity having a correlation with the deviation between the actual gear ratio and the target gear ratio, specifically, in the case of a belt-type continuously variable transmission, the deviation between the target sheave position and the actual sheave position, The deviation between the target rotational speed of the primary sheave) and the actual rotational speed of the sheave is exemplified. If the physical quantity correlated to the deviation between the actual gear ratio and the target gear ratio is less than or equal to a predetermined value, the integral term will not be excessive even if the integral term is released. By starting (accumulation), it becomes possible to bring the actual gear ratio closer to the target gear ratio without causing overshoot or undershoot.

  Specifically, as the integral term fixed condition, a physical quantity correlated with the change amount of the target gear ratio is a predetermined value or more, and a physical quantity correlated with a deviation between the actual gear ratio and the target gear ratio is a predetermined value. It is assumed that this is the case.

  Here, specifically, in the case of a belt-type continuously variable transmission, the amount of change in the target sheave position and the change in the target rotational speed of the sheave (especially the primary sheave) as a physical quantity correlated with the amount of change in the target gear ratio. Amount and the like. Considering that there is a concern that the integral term will become excessive due to a response delay of control, etc., when the physical quantity correlated with the change amount of the target gear ratio is a predetermined value or more, In such a case, the integral term is fixed on condition that the physical quantity correlated with the deviation between the actual speed ratio and the target speed ratio is equal to or greater than a predetermined value. This suppresses the integral term from becoming excessive.

  In the present invention, when the integral term is fixed and the integral term is fixed in the shift feedback control, the state where the integral term is fixed is continued for a predetermined time or more. As a result, accumulation of integral terms is suppressed in a state where the deviation between the target speed ratio and the actual speed ratio is large.

It is a schematic block diagram which shows the power transmission system of the vehicle by which the belt type continuously variable transmission which concerns on embodiment is mounted. It is a circuit block diagram of the hydraulic control circuit which controls the hydraulic actuator of a primary pulley. It is a circuit block diagram of the hydraulic control circuit which controls the clamping pressure of a belt. It is a figure which shows an example of the map used for the shift control of a belt-type continuously variable transmission. It is a figure which shows an example of the map used for the belt clamping pressure control of a belt-type continuously variable transmission. It is a block diagram which shows the structure of control systems, such as ECU. It is a flowchart figure which shows the procedure of the integral term control in transmission feedback control. It is a timing chart figure showing an example of each time change of a target sheave position, an actual sheave position, and an integral term hold execution flag.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. This embodiment demonstrates the case where this invention is applied to the belt-type continuously variable transmission mounted in the motor vehicle.

  FIG. 1 is a schematic configuration diagram showing a power transmission system of a vehicle equipped with a belt type continuously variable transmission according to the present embodiment. As shown in FIG. 1, the vehicle according to the present embodiment is an FF (front engine / front drive) type vehicle, and is an engine (internal combustion engine) 1 that is a driving power source, and a torque converter as a fluid transmission device. 2, a forward / reverse switching device 3, a belt-type continuously variable transmission 4, a reduction gear device 5, a differential gear device 6, an ECU (Electronic Control Unit) 8, and the like are mounted. The control device according to the present invention is realized by the ECU 8, the belt type continuously variable transmission 4, the hydraulic control circuit 20 described later, and the like.

  A crankshaft 11 that is an output shaft of the engine 1 is connected to the torque converter 2, and the output of the engine 1 is transmitted from the torque converter 2 through the forward / reverse switching device 3, the belt-type continuously variable transmission 4, and the reduction gear device 5. Are transmitted to the differential gear device 6 and distributed to the left and right drive wheels 10, 10.

  The parts of the engine 1, the torque converter 2, the forward / reverse switching device 3, the belt-type continuously variable transmission 4, and the ECU 8 will be described below.

-Engine-
The engine 1 is a multi-cylinder gasoline engine, for example. The amount of intake air taken into the engine 1 is adjusted by an electronically controlled throttle valve 12. The throttle valve 12 can electronically control the throttle opening independently of the driver's accelerator pedal operation, and the opening (throttle opening) is detected by the throttle opening sensor 102. Further, the coolant temperature of the engine 1 is detected by a water temperature sensor 103.

  The throttle opening of the throttle valve 12 is driven and controlled by the ECU 8. Specifically, the optimum intake air amount (in accordance with the operating state of the engine 1 such as the engine speed Ne detected by the engine speed sensor 101 and the accelerator pedal depression amount (accelerator operation amount Acc) of the driver ( The throttle opening of the throttle valve 12 is controlled so as to obtain a target intake air amount. More specifically, the actual throttle opening of the throttle valve 12 is detected using the throttle opening sensor 102, and the actual throttle opening coincides with the throttle opening (target throttle opening) at which the target intake air amount can be obtained. Thus, the throttle motor 13 of the throttle valve 12 is feedback-controlled.

-Torque converter-
The torque converter 2 includes an input-side pump impeller 21, an output-side turbine runner 22, a stator 23 that develops a torque amplification function, and the like, and fluid is passed between the pump impeller 21 and the turbine runner 22. Transmit power. The pump impeller 21 is connected to the crankshaft 11 of the engine 1. The turbine runner 22 is connected to the forward / reverse switching device 3 via the turbine shaft 27.

  The torque converter 2 is provided with a lockup clutch 24 that directly connects the input side and the output side of the torque converter 2. The lock-up clutch 24 controls the differential pressure (lock-up differential pressure) between the hydraulic pressure in the engagement side oil chamber 25 and the hydraulic pressure in the release side oil chamber 26 to achieve full engagement and half engagement (in the slip state). Engagement) or release.

  By completely engaging the lockup clutch 24, the pump impeller 21 and the turbine runner 22 rotate integrally. Further, by engaging the lockup clutch 24 in a predetermined slip state (half-engaged state), the turbine runner 22 rotates following the pump impeller 21 with a predetermined slip amount during driving. On the other hand, by setting the lockup differential pressure to be negative, the lockup clutch 24 is released.

  The torque converter 2 is provided with a mechanical oil pump (hydraulic pressure generating source) 7 that is connected to and driven by the pump impeller 21.

-Forward / reverse switching device-
The forward / reverse switching device 3 includes a double pinion type planetary gear mechanism 30, a forward clutch C1, and a reverse brake B1.

  The sun gear 31 of the planetary gear mechanism 30 is integrally connected to the turbine shaft 27 of the torque converter 2, and the carrier 33 is integrally connected to the input shaft 40 of the belt type continuously variable transmission 4. The carrier 33 and the sun gear 31 are selectively connected via the forward clutch C1, and the ring gear 32 is selectively fixed to the housing via the reverse brake B1.

  The forward clutch C1 and the reverse brake B1 are hydraulic friction engagement elements that are engaged and released by a hydraulic control circuit 20 to be described later. The forward clutch C1 is engaged and the reverse brake B1 is released. Thus, the forward / reverse switching device 3 is integrally rotated to establish (achieve) the forward power transmission path, and in this state, the forward driving force is transmitted to the belt-type continuously variable transmission 4 side. .

  On the other hand, when the reverse brake B1 is engaged and the forward clutch C1 is released, the forward / reverse switching device 3 establishes (achieves) a reverse power transmission path. In this state, the input shaft 40 rotates in the reverse direction with respect to the turbine shaft 27, and the driving force in the reverse direction is transmitted to the belt type continuously variable transmission 4 side. When both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 3 is in a neutral state (blocking state) that blocks power transmission.

-Belt type continuously variable transmission-
The belt-type continuously variable transmission 4 receives power from the engine 1, shifts the rotational speed of the input shaft 40, and transmits it to the drive wheels 10, 10 side. The output side secondary pulley 42 and a metal belt 43 wound around the primary pulley 41 and the secondary pulley 42 are provided.

  The primary pulley 41 is a variable pulley having a variable effective diameter, and a fixed sheave 411 fixed to the input shaft 40 and a movable sheave 412 disposed on the input shaft 40 in a state in which sliding is possible only in the axial direction. And is composed of. Similarly, the secondary pulley 42 is a variable pulley whose effective diameter is variable, and is a fixed sheave 421 fixed to the output shaft 44 and a movable sheave arranged on the output shaft 44 so as to be slidable only in the axial direction. 422.

  A hydraulic actuator 413 for changing the V groove width between the fixed sheave 411 and the movable sheave 412 is disposed on the movable sheave 412 side of the primary pulley 41. Similarly, a hydraulic actuator 423 for changing the V groove width between the fixed sheave 421 and the movable sheave 422 is also arranged on the movable sheave 422 side of the secondary pulley 42.

  In the belt-type continuously variable transmission 4 having the above structure, by controlling the hydraulic pressure of the hydraulic actuator 413 of the primary pulley 41, the V groove widths of the primary pulley 41 and the secondary pulley 42 change, and the engagement diameter of the belt 43 ( The effective gear ratio) is changed, and the gear ratio γ (γ = primary pulley rotation speed (input shaft rotation speed) Nin / secondary pulley rotation speed (output shaft rotation speed) Nout) continuously changes. The hydraulic pressure of the hydraulic actuator 423 of the secondary pulley 42 is controlled such that the belt 43 is clamped with a predetermined clamping pressure that does not cause belt slip. These controls are executed by the ECU 8 and the hydraulic control circuit 20.

-Hydraulic control circuit-
As shown in FIG. 1, the hydraulic control circuit 20 includes a gear ratio control unit 20a, a belt clamping pressure control unit 20b, a line pressure control unit 20c, a lockup engagement pressure control unit 20d, a clutch pressure control unit 20e, and a manual It is constituted by a valve 20f or the like.

  Further, a shift control solenoid (linear solenoid; SLP) 304 and a shift control solenoid (linear solenoid; SLP) 305, a linear solenoid (SLS) 202 for belt clamping pressure control, a line, which constitute the hydraulic control circuit 20 Control signals from the ECU 8 are supplied to the linear solenoid (SLT) 201 for pressure control and the duty solenoid (DSU) 307 for lockup engagement pressure control.

  Next, in the hydraulic control circuit 20, the hydraulic control circuit of the hydraulic actuator 413 of the primary pulley 41 of the belt-type continuously variable transmission 4 (specific hydraulic circuit configuration of the transmission ratio control unit 20 a), and the secondary pulley 42 A hydraulic control circuit of the hydraulic actuator 423 (specific hydraulic circuit configuration of the belt clamping pressure control unit 20b) will be described with reference to FIGS. 2 and 3 are examples of a hydraulic control circuit. The hydraulic control circuit applicable to the present invention is not limited to those shown in these drawings, and various circuits can be applied.

  First, as shown in FIG. 3, the hydraulic pressure generated by the oil pump 7 is regulated by the primary regulator valve 203 to generate the line pressure PL. The primary regulator valve 203 is supplied with the control hydraulic pressure output from the linear solenoid (SLT) 201 via the clutch apply control valve 204, and operates with the control hydraulic pressure as a pilot pressure.

  Note that there is a case where the control hydraulic pressure from the linear solenoid (SLS) 202 is supplied to the primary regulator valve 203 by switching the clutch apply control valve 204, and the line pressure PL is regulated using the control hydraulic pressure as a pilot pressure. The linear solenoid (SLT) 201 and the linear solenoid (SLS) 202 are supplied with the hydraulic pressure regulated by the modulator valve 205 using the line pressure PL as a source pressure.

  The linear solenoid (SLT) 201 outputs a control hydraulic pressure according to a current value determined by a duty signal output from the ECU 8. The linear solenoid (SLT) 201 is a normally open type solenoid valve.

  Further, the linear solenoid (SLS) 202 outputs a control hydraulic pressure according to a current value determined by a duty signal output from the ECU 8. The linear solenoid (SLS) 202 is also a normally open type solenoid valve like the linear solenoid (SLT) 201.

  2 and 3, the modulator valve 206 adjusts the hydraulic pressure output from the modulator valve 205 to a constant pressure, so that a shift control solenoid (SLP) 304 and a shift control solenoid (SLP), which will be described later. ) 305 and the belt clamping pressure control valve 303 and the like.

(Circuit configuration for shift control)
Next, a hydraulic control circuit for the hydraulic actuator 413 of the primary pulley 41 will be described. As shown in FIG. 2, an upshift transmission control valve 301 is connected to the hydraulic actuator 413 of the primary pulley 41.

  The upshift transmission control valve 301 is provided with a spool 311 that is movable in the axial direction. A spring 312 is disposed on one end side (the upper end side in FIG. 2) of the spool 311, and a first hydraulic port 315 is formed at the end opposite to the spring 312 across the spool 311. Further, a second hydraulic port 316 is formed on the one end side where the spring 312 is disposed.

  The first hydraulic pressure port 315 is connected to a shift control solenoid (SLP) 304 that outputs a control hydraulic pressure in accordance with a current value determined by a signal output from the ECU 8, and a control that the shift control solenoid (SLP) 304 outputs. Oil pressure is applied to the first oil pressure port 315. The second hydraulic pressure port 316 is connected to a shift control solenoid (SLP) 305 that outputs a control hydraulic pressure in accordance with a current value determined by a signal output from the ECU 8, and a control that the shift control solenoid (SLP) 305 outputs. Oil pressure is applied to the second oil pressure port 316.

  Further, the upshift transmission control valve 301 is formed with an input port 313 to which the line pressure PL is supplied, an input / output port 314 and an output port 317 that are connected (communication) to the hydraulic actuator 413 of the primary pulley 41. When the spool 311 is in the upshift position (the right position in FIG. 2), the output port 317 is closed, and the line pressure PL is supplied from the input port 313 to the hydraulic actuator 413 of the primary pulley 41 via the input / output port 314. . On the other hand, when the spool 311 is in the closed position (the left position in FIG. 2), the input port 313 is closed, and the hydraulic actuator 413 of the primary pulley 41 communicates with the output port 317 via the input / output port 314.

  The downshift transmission control valve 302 is provided with a spool 321 that is movable in the axial direction. A spring 322 is disposed on one end side (the lower end side in FIG. 2) of the spool 321 and a first hydraulic port 326 is formed on one end side thereof. A second hydraulic port 327 is formed at the end opposite to the spring 322 across the spool 321. The first hydraulic pressure port 326 is connected to the shift control solenoid (SLP) 304, and the control hydraulic pressure output from the shift control solenoid (SLP) 304 is applied to the first hydraulic pressure port 326. The shift hydraulic control solenoid (SLP) 305 is connected to the second hydraulic pressure port 327, and the control hydraulic pressure output from the shift control solenoid (SLP) 305 is applied to the second hydraulic pressure port 327.

  Further, the downshift transmission control valve 302 is formed with an input port 323, an input / output port 324, and a discharge port 325. A bypass control valve 306 is connected to the input port 323, and a hydraulic pressure obtained by adjusting the line pressure PL by the bypass control valve 306 is supplied. In such a downshift transmission control valve 302, the input / output port 324 communicates with the discharge port 325 when the spool 321 is in the downshift position (left side position in FIG. 2). On the other hand, when the spool 321 is in the closed position (right side position in FIG. 2), the input / output port 324 is closed. The input / output port 324 of the downshift transmission control valve 302 is connected to the output port 317 of the upshift transmission control valve 301.

  In the hydraulic control circuit of FIG. 2 described above, the shift control solenoid (SLP) 304 is actuated in response to the upshift gear shift command output from the ECU 8, and the control hydraulic pressure output from the shift control solenoid (SLP) 304 is the shift gear for upshift. When supplied to the first hydraulic port 315 of the control valve 301, the spool 311 moves to the upshift position side (upper side in FIG. 2) by a thrust according to the control hydraulic pressure. By the movement of the spool 311 (movement toward the upshift side), hydraulic oil (line pressure PL) is supplied from the input port 313 to the hydraulic actuator 413 of the primary pulley 41 through the input / output port 314 at a flow rate corresponding to the control hydraulic pressure. At the same time, the output port 317 is closed and the flow of hydraulic oil to the downshift transmission control valve 302 is blocked. As a result, the transmission control pressure is increased, the V groove width of the primary pulley 41 is reduced, and the transmission ratio γ is reduced (upshift).

  When the control hydraulic pressure output from the shift control solenoid (SLP) 304 is supplied to the first hydraulic port 326 of the downshift transmission control valve 302, the spool 321 moves upward in FIG. Closed.

  On the other hand, a shift control solenoid (SLP) 305 is operated in accordance with a downshift gear shift command output from the ECU 8, and the control hydraulic pressure output from the shift control solenoid (SLP) 305 is the second hydraulic port of the upshift shift control valve 301. When supplied to 316, the spool 311 moves to the downshift position side (the lower side in FIG. 2) by the thrust according to the control oil pressure. By this movement of the spool 311 (movement toward the downshift side), the hydraulic oil in the hydraulic actuator 413 of the primary pulley 41 flows into the input / output port 314 of the upshift transmission control valve 301 at a flow rate corresponding to the control hydraulic pressure.

  When the control hydraulic pressure output from the shift control solenoid (SLP) 305 is supplied to the second hydraulic pressure port 327 of the downshift transmission control valve 302, the spool 321 moves downward in FIG. And the discharge port 325 communicate with each other.

  As a result, the hydraulic oil flowing into the upshift transmission control valve 301 is discharged from the discharge port 325 through the output port 317 and the input / output port 324 of the downshift transmission control valve 302. As a result, the transmission control pressure is lowered, the V groove width of the primary pulley 41 is increased, and the transmission ratio γ is increased (downshift).

  As described above, when the control hydraulic pressure is output from the shift control solenoid (SLP) 304, hydraulic oil is supplied from the upshift shift control valve 301 to the hydraulic actuator 413 of the primary pulley 41, and the shift control pressure is continuously increased. Ascend and upshift. Further, when the control hydraulic pressure is output from the shift control solenoid (SLP) 305, the hydraulic oil in the hydraulic actuator 413 of the primary pulley 41 is discharged from the discharge port 325 of the downshift shift control valve 302, and the shift control pressure continues. Down and downshifted.

  In this example, as shown in FIG. 4, for example, the target target rotational speed Nint is calculated from a preset shift map using the accelerator operation amount Acc representing the driver's requested output amount and the vehicle speed V as parameters, The shift control of the belt-type continuously variable transmission 4, that is, the hydraulic actuator 413 of the primary pulley 41 is controlled according to the deviation (Nint−Nin) so that the actual input shaft speed Nin matches the target speed Nint. The transmission control pressure is controlled by supplying and discharging the hydraulic oil, and the transmission ratio γ continuously changes. The map in FIG. 4 corresponds to the shift conditions and is stored in the ROM 82 (see FIG. 6) of the ECU 8.

  In the map of FIG. 4, the target rotational speed Nint is set such that the larger the vehicle speed V is and the greater the accelerator operation amount Acc is, the larger the gear ratio γ is. Further, since the vehicle speed V corresponds to the secondary pulley rotational speed (output shaft rotational speed) Nout, the target rotational speed Nint, which is the target value of the primary pulley rotational speed (input shaft rotational speed) Nin, corresponds to the target gear ratio, It is set within the range of the minimum speed ratio γmin and the maximum speed ratio γmax of the continuously variable transmission 4.

(Circuit configuration for belt clamping pressure control)
Next, a hydraulic control circuit of the hydraulic actuator 423 of the secondary pulley 42 will be described with reference to FIG.

  As shown in FIG. 3, a belt clamping pressure control valve 303 is connected to the hydraulic actuator 423 of the secondary pulley 42.

  The belt clamping pressure control valve 303 is provided with a spool 331 that is movable in the axial direction. A spring 332 is disposed on one end side (the lower end side in FIG. 3) of the spool 331, and a first hydraulic port 335 is formed on one end side thereof. A second hydraulic port 336 is formed at the end opposite to the spring 332 across the spool 331.

  A linear solenoid (SLS) 202 is connected to the first hydraulic pressure port 335, and a control hydraulic pressure output from the linear solenoid (SLS) 202 is applied to the first hydraulic pressure port 335. The hydraulic pressure from the modulator valve 206 is applied to the second hydraulic pressure port 336.

  Further, the belt clamping pressure control valve 303 is formed with an input port 333 to which the line pressure PL is supplied and an output port 334 connected (communication) to the hydraulic actuator 423 of the secondary pulley 42.

  In the hydraulic control circuit of FIG. 3, when the control hydraulic pressure output from the linear solenoid (SLS) 202 increases from the state where a predetermined hydraulic pressure is supplied to the hydraulic actuator 423 of the secondary pulley 42, the belt clamping pressure control valve 303 The spool 331 moves upward in FIG. In this case, the hydraulic pressure supplied to the hydraulic actuator 423 of the secondary pulley 42 increases, and the belt clamping pressure increases.

  On the other hand, when the control hydraulic pressure output from the linear solenoid (SLS) 202 decreases from the state in which the predetermined hydraulic pressure is supplied to the hydraulic actuator 423 of the secondary pulley 42, the spool 331 of the belt clamping pressure control valve 303 is lowered in FIG. Move to the side. In this case, the hydraulic pressure supplied to the hydraulic actuator 423 of the secondary pulley 42 decreases, and the belt clamping pressure decreases.

  In this way, the belt clamping pressure is increased or decreased by adjusting the line pressure PL using the control hydraulic pressure output from the linear solenoid (SLS) 202 as a pilot pressure and supplying it to the hydraulic actuator 423 of the secondary pulley 42.

  In this example, as shown in FIG. 5, for example, the accelerator opening Acc and the gear ratio γ (γ = Nin / Nout) corresponding to the transmission torque are used as parameters, and it is necessary to set in advance so that belt slip does not occur. By controlling the control hydraulic pressure output from the linear solenoid (SLS) 202 according to a map of hydraulic pressure (equivalent to belt clamping pressure), the belt clamping pressure of the belt-type continuously variable transmission 4, that is, the hydraulic actuator 423 of the secondary pulley 42 is controlled. Regulates the hydraulic pressure of The map in FIG. 5 corresponds to the clamping pressure control condition and is stored in the ROM 82 (see FIG. 6) of the ECU 8.

(Hydraulic control)
Next, the hydraulic pressure adjusting operation for the belt clamping pressure control and the hydraulic pressure adjusting operation for the shift control will be specifically described. That is, the primary pulley 41 is used to adjust the hydraulic pressure (secondary pulley command pressure) supplied to the hydraulic actuator 423 of the secondary pulley 42 in order to suppress slippage of the belt 43 with respect to the pulleys 41 and 42 and the target gear ratio. The operation of adjusting the hydraulic pressure (primary pulley command pressure) supplied to the hydraulic actuator 413 will be specifically described.

<Secondary pulley command pressure>
Adjustment of the hydraulic pressure supplied to the hydraulic actuator 423 of the secondary pulley 42 is performed based on this hydraulic pressure command pressure TPout. This command pressure TPout is based on a target thrust TWout, which is a target value of the thrust of the secondary pulley 42, and a pressure receiving area Aout of the pulley 42 (movable sheave 422) that receives the hydraulic pressure acting on the secondary pulley 42. 1).

TPout = TWout / Aout (1)
The target thrust TWout in the equation (1) is an estimated input torque Tt that is an estimated value of the torque input to the input shaft 40 of the belt type continuously variable transmission 4 and an actual speed ratio Rγ of the belt type continuously variable transmission 4. Based on this, the value is calculated so as not to cause a slip between the belt 43 and the pulleys 41 and 42. The estimated input torque Tt used here is obtained based on the throttle opening of the engine 1 and the like, and the actual gear ratio Rγ is based on the input shaft rotational speed Nin and the output shaft rotational speed Nout of the belt-type continuously variable transmission 4. Ratio (Nin / Nout). Then, by adjusting the hydraulic pressure acting on the hydraulic actuator 423 of the secondary pulley 42 based on the command pressure TPout calculated using the equation (1), the slip of the belt 43 with respect to the pulleys 41 and 42 is suppressed. .

<Primary pulley command pressure>
On the other hand, adjustment of the hydraulic pressure supplied to the hydraulic actuator 413 of the primary pulley 41 is performed based on the hydraulic pressure command pressure Pin for setting the actual transmission gear ratio Rγ of the belt-type continuously variable transmission 4 to the target transmission gear ratio Tγ. The target gear ratio Tγ of the belt-type continuously variable transmission 4 is based on the accelerator pedal depression amount (accelerator operation amount Acc), the vehicle speed V, etc., as described above with reference to FIG. Is set as a transmission ratio. The vehicle speed V is obtained based on the output shaft rotation speed Nout. The command pressure Pin is the same based on the feedforward pressure Pinff calculated as the hydraulic pressure required to achieve the target gear ratio Tγ and the deviation Δγ between the actual gear ratio Rγ and the target gear ratio Tγ. Based on the feedback pressure Pinfb that increases or decreases to reduce the deviation Δγ, the calculation is performed using the following equation (2).

Pin = Pinff + Pinfb (2)
By adjusting the hydraulic pressure supplied to the hydraulic actuator 413 of the primary pulley 41 based on the command pressure Pin calculated using this equation (2), the primary pulley 41 is driven (the movable sheave 412 slides), and the pulley The distance (hanging diameter) from the rotation center of the belt 41 to the belt 43 is adjusted, whereby the actual transmission gear ratio Rγ of the belt type continuously variable transmission 4 approaches the target transmission gear ratio Tγ.

  If the command pressure Pin is calculated only from the feedforward pressure Pinff, when the hydraulic pressure supplied to the hydraulic actuator 413 of the primary pulley 41 is adjusted based on the command pressure Pin, the actual gear ratio Rγ is set to the target gear ratio Tγ. There is a possibility that it cannot be approached. This is inevitable that individual differences and aging deterioration occur in the belt-type continuously variable transmission 4, and the calculated command pressure Pin (= feedforward pressure Pinff) is equivalent to such individual differences and aging deterioration. This is because the actual speed ratio Rγ deviates from an appropriate value as the target speed ratio Tγ.

  By calculating the command pressure Pin as expressed by the above equation (2), correction by the feedback pressure Pinfb corresponding to individual differences of the belt-type continuously variable transmission 4 and aging deterioration is performed, and the actual gear ratio Rγ And the target pressure ratio Tγ can be eliminated by increasing or decreasing the feedback pressure Pinfb based on the deviation Δγ. The deviation Δγ is calculated by subtracting the target speed ratio Tγ from the actual speed ratio Rγ. Further, the feedback pressure Pinfb is increased when the deviation Δγ is a negative value, and is decreased when the deviation Δγ is a positive value. Thus, by increasing / decreasing the feedback pressure Pinfb, the deviation of the indicated pressure Pin from the appropriate value due to individual differences of the belt-type continuously variable transmission 4 or aging deterioration is eliminated, and the primary pulley 41 is moved to the primary pulley 41 based on the indicated pressure Pin. It is possible to make the actual speed ratio Rγ substantially coincide with the target speed ratio Tγ when the acting hydraulic pressure is adjusted.

  Note that the deviation of the actual transmission ratio Rγ with respect to the target transmission ratio Tγ due to individual differences of the belt-type continuously variable transmission 4 or deterioration over time is constant, and the steady deviation disappears through the increase or decrease of the feedback pressure Pinfb. The feedback pressure Pinfb converges to a value corresponding to the steady deviation.

  Next, a method for calculating the feedforward pressure Pinff used in Expression (2) will be described in detail.

  This feed-forward pressure Pinff has a steady term Pinc that is a hydraulic pressure that should be supplied to the hydraulic actuator 413 of the primary pulley 41 in order to maintain the actual gear ratio Rγ at the target gear ratio Tγ, and a target gear ratio Tγ that changes during gear shifting. At the same time, based on the shift term Pins, which is the hydraulic pressure to be supplied to the hydraulic actuator 413 of the primary pulley 41 in order to change the actual speed ratio Rγ, it is calculated using the following equation (3).

Pinff = Pinc + Pins (3)
The steady term Pinc of this equation (3) is the actual thrust RWout, which is the actual value of the thrust of the secondary pulley 42, and the pulley 41 (movable sheave 412) that receives the hydraulic pressure supplied to the hydraulic actuator 413 of the primary pulley 41. Based on the pressure receiving area Ain and the thrust ratio τ representing the ratio of the thrusts of the pulleys 41 and 42, the pressure is calculated using the following equation (4).

Pinc = RWout / (τ · Ain) (4)
The actual thrust RWout is obtained by multiplying the actual hydraulic pressure RPout, which is an actual value of the hydraulic pressure supplied to the hydraulic actuator 423 of the secondary pulley 42, by the pressure receiving area Aout of the pulley 42. The thrust ratio τ is calculated as a ratio (RWout / RWin) between the actual thrust RWout, which is the actual value of the thrust of the secondary pulley 42, and the actual thrust RWin, which is the actual value of the thrust of the primary pulley 41. Value. Specifically, the thrust ratio τ is based on the target speed ratio Tγ that is variably set as described above, and the safety factor SF that is a value representing the margin with respect to the necessary minimum value Woutmin of the actual thrust RWout of the secondary pulley 42. It is calculated with reference to a map predetermined by experiment or the like.

  The safety factor SF is calculated as, for example, “1”, “2”, “3”, “4”, “5” based on the ratio (RWout / Woutmin) between the actual thrust RWout and the necessary minimum value Woutmin. The larger the value is, the larger the margin is. The necessary minimum value Woutmin used for the calculation of the safety factor SF is a value representing the minimum value of the actual thrust RWout required for suppressing the slip of the belt 43 with respect to the pulleys 41 and 42. This required minimum value Woutmin is the estimated input torque Tt obtained from the throttle opening, the winding radius Rin of the belt 43 around the primary pulley 41, the friction coefficient μ between the belt 43 and the primary pulley 41, and the primary pulley 41 Based on the sandwiching angle α, it is calculated using the following equation (5).

Woutmin = Tt · cos α / (2 · μ · Rin) (5)
The winding radius Rin is a value representing the shortest distance from the center of rotation of the primary pulley 41 to the belt 43, and is calculated with reference to a map predetermined by experiment or the like based on the actual gear ratio Rγ. Further, the sandwiching angle α of the primary pulley 41 is a value representing the inclination angle of the contact surface between the pulley 41 and the belt 43 with respect to a plane orthogonal to the center line of the pulley 41.

  Further, the gear shift term Pins of the equation (3) is calculated using the following equation (6) based on the time change dgdt representing the change speed of the target gear ratio Tγ and the proportionality constant K.

Pins = dgdt / K (6)
The proportional constant K is a constant for obtaining, from the time change dgdt, the hydraulic pressure that acts on the primary pulley 41 that is required to change the actual speed ratio Rγ in accordance with the changing target speed ratio Tγ. The proportionality constant K is calculated with reference to a map set in advance through experiments or the like based on the actual speed ratio Rγ, the input rotation speed Nin, and the actual hydraulic pressure RPout acting on the secondary pulley 42.

  Next, a method for calculating the feedback pressure Pinfb used in Expression (2) will be described in detail.

  The feedback pressure Pinfb is calculated using the following equation (7) based on the deviation Δγ between the actual speed ratio Rγ and the target speed ratio Tγ, the proportional gain Gp, and the integral gain Gi.

Pinfb = Δγ · Gp + ΣΔγ · Gi (7)
ΣΔγ: Deviation integrated value “Δγ · Gp”, which is the first term on the right side in the equation (7), is a proportional term that takes a magnitude proportional to the deviation Δγ. The proportional term “Δγ · Gp” increases or decreases the feedback pressure Pinfb by an amount corresponding to the deviation Δγ to increase or decrease the hydraulic pressure acting on the hydraulic actuator 413 of the primary pulley 41, and the actual speed ratio Rγ is set to the target speed ratio Tγ. It is for getting closer. Further, the proportional gain Gp used in the proportional term “Δγ · Gp” is a constant obtained in advance through experiments or the like.

  Further, “ΣΔγ · Gi”, which is the second term on the right-hand side in the equation (7), is an actual speed ratio Rγ and a target speed ratio Tγ that cannot be canceled only by increasing / decreasing the feedback pressure Pinfb by the proportional term “Δγ · Gp”. Is an integral term to eliminate the residual deviation between This integral term “ΣΔγ · Gi” is used to increase or decrease the hydraulic pressure acting on the hydraulic actuator 413 of the primary pulley 41 by an amount corresponding to the residual deviation so as to match the actual gear ratio Rγ and the target gear ratio Tγ. It is. The deviation integrated value ΣΔγ used in the integral term “ΣΔγ · Gi” is a value obtained through integration processing in which the deviation Δγ is added at predetermined time intervals. In this integration process, a calculation “ΣΔγ ← previous ΣΔγ + Δγ” is executed at predetermined time intervals. Further, the integral gain Gi used in the integral term “ΣΔγ · Gi” is a constant obtained in advance through experiments or the like.

  The feedback pressure Pinfb calculated using the equation (7) increases / decreases so as to reduce the deviation Δγ based on the deviation Δγ between the actual transmission ratio Rγ and the target transmission ratio Tγ. Through the increase / decrease of the feedback pressure Pinfb, feedback control for setting the actual speed ratio Rγ to the target speed ratio Tγ is performed. When the feedback control is performed, even if a steady shift of the actual speed ratio Rγ with respect to the target speed ratio Tγ occurs due to individual differences of the belt-type continuously variable transmission 4 or aging deterioration, the steady shift Can be eliminated by increasing or decreasing the feedback pressure Pinfb.

  The above is the hydraulic pressure adjusting operation for the belt clamping pressure control and the hydraulic pressure adjusting operation for the shift control.

-ECU-
As shown in FIG. 6, the ECU 8 includes a CPU 81, a ROM 82, a RAM 83, a backup RAM 84, and the like.

  The ROM 82 stores various control programs, maps that are referred to when the various control programs are executed, and the like. The CPU 81 executes arithmetic processing based on various control programs and maps stored in the ROM 82. The RAM 83 is a memory for temporarily storing calculation results in the CPU 81 and data input from each sensor. The backup RAM 84 is a non-volatile memory for storing data to be saved when the engine 1 is stopped. is there.

  These CPU 81, ROM 82, RAM 83, and backup RAM 84 are connected to each other via a bus 87, and are connected to an input interface 85 and an output interface 86.

  The input interface 85 of the ECU 8 includes an engine speed sensor 101, a throttle opening sensor 102, a water temperature sensor 103, a turbine speed sensor 104, a primary pulley speed sensor 105, a secondary pulley speed sensor 106, an accelerator position sensor 107, CVT oil temperature sensor 108, brake switch 109, lever position sensor 110 for detecting the lever position (operation position) of the shift lever 9, primary sheave position sensor 111 for detecting the sliding position of the movable sheave 412 of the primary pulley 41, the secondary A secondary sheave position sensor 112 for detecting the slide position of the movable sheave 422 of the pulley 42 is connected. Output signals of these sensors, that is, the rotational speed of the engine 1 (engine rotational speed) Ne, the throttle opening θth of the throttle valve 12, the cooling water temperature Tw of the engine 1, the rotational speed of the turbine shaft 27 (turbine rotational speed) Nt, primary Pulley rotational speed (input shaft rotational speed) Nin, secondary pulley rotational speed (output shaft rotational speed) Nout, accelerator pedal operation amount (accelerator opening) Acc, oil temperature of hydraulic control circuit 20 (CVT oil temperature Thc), normal use Whether the foot brake as a brake is operated (brake ON / OFF), the lever position (operating position) of the shift lever 9, the sliding position of the movable sheave 412 of the primary pulley 41 (corresponding to the actual gear ratio Rγ), the secondary pulley 42 A signal representing the slide position of the movable sheave 422 is supplied to the ECU 8.

  The output interface 86 is connected to the throttle motor 13, the fuel injection device 14, the ignition device 15, the hydraulic control circuit 20, and the like.

  Here, among the signals supplied to the ECU 8, the turbine rotational speed Nt coincides with the primary pulley rotational speed (input shaft rotational speed) Nin during forward travel in which the forward clutch C1 of the forward / reverse switching device 3 is engaged. The secondary pulley rotational speed (output shaft rotational speed) Nout corresponds to the vehicle speed V. The accelerator operation amount Acc represents the driver's requested output amount.

  The shift lever 9 includes a parking position “P” for parking, a reverse position “R” for reverse traveling, a neutral position “N” for interrupting power transmission, a drive position “D” for forward traveling, During forward running, the gear ratio γ of the belt type continuously variable transmission 4 is selectively operated to each position such as a manual position “M” where the manual operation can increase or decrease the speed ratio γ.

  The manual position “M” includes a plurality of ranges in which a downshift position and an upshift position for increasing / decreasing the speed ratio γ, or a plurality of speed ranges in which the upper limit of the speed range (the side where the speed ratio γ is smaller) are different can be selected. Position etc. are provided.

  The ECU 8 controls the output of the engine 1, the gear ratio control and the belt clamping pressure control of the belt-type continuously variable transmission 4 and the engagement of the lock-up clutch 24 based on the output signals of the various sensors.・ Execute release control. The output control of the engine 1 is executed by the throttle motor 13, the fuel injection device 14, the ignition device 15, the ECU 8, and the like.

−Integral term change operation−
Next, an operation for changing the integral term (ΣΔγ · Gi), which is an operation characteristic of the present embodiment, will be described. In this integral term change operation, the integral term is held (maintains the current value of the integral term) according to the speed of change of the gear ratio, etc., or the hold is released based on the integral term hold time, etc. (Accumulation)).

  The outline of the integral term changing operation will be described below.

  In the prior art (Patent Document 2), when the shift speed is equal to or higher than a predetermined value, the accumulation of the integral term is interrupted, thereby suppressing the excessive accumulation of the integral term, Undershoot was suppressed. However, if the release timing of the operation that suppresses the excessive accumulation of the integral term (the operation that interrupts the accumulation of the integral term) is not properly obtained, the hydraulic pressure may overshoot or undershoot. There is. For example, if the operation for suppressing the excessive accumulation of the integral term is canceled immediately after the start of gear shifting, the target value of the gear ratio (target gear shifting) is caused due to the existence of a hydraulic response delay period immediately after the gear shifting is started. The accumulation of the integral term is started in a state where the difference between the ratio Tγ) and the actual value (actual transmission ratio Rγ) is large. For this reason, there is a possibility that the integral term is excessively accumulated and the hydraulic overshoot or undershoot occurs.

  In view of this point, in the present embodiment, a predetermined integral term hold execution condition (for example, the change amount of the target gear ratio Tγ is equal to or greater than a predetermined value) is satisfied (“integral term fixed condition” in the present invention). If the above holds, the integral term is held (fixed). When the change amount of the actual gear ratio Rγ is equal to or less than a predetermined value and the time during which the integral term is held is continued for a predetermined time or longer, the integral term hold release condition is satisfied (referred to in the present invention). The integration term hold is released (assuming that the "integration term fixed release condition" is satisfied). This ensures that the integral term is held for a predetermined time or longer, and accumulation of the integral term is started in a state where the actual gear ratio Rγ is close to the target gear ratio Tγ (for example, the hydraulic response delay is eliminated). To be. That is, accumulation of integral terms is prevented from starting in a state where the deviation between the actual speed ratio Rγ and the target speed ratio Tγ is large.

  Hereinafter, the integral term changing operation (integral term control) will be specifically described with reference to the flowchart of FIG. The flowchart shown in FIG. 7 is executed every several milliseconds while the vehicle is traveling (for example, while the vehicle is traveling after the temperature of the hydraulic oil has risen above a predetermined temperature).

  First, in step ST1, information from each sensor is acquired. Specifically, primary pulley rotational speed (input shaft rotational speed) information from the primary pulley rotational speed sensor 105, secondary pulley rotational speed (output shaft rotational speed) information from the secondary pulley rotational speed sensor 106, the primary sheave The slide position information of the movable sheave 412 (movable sheave 412 of the primary pulley 41) from the position sensor 111 is acquired. Instead of the information from the primary sheave position sensor 111 or in addition to the information from the primary sheave position sensor 111, the movable sheave 422 from the secondary sheave position sensor 112 (the movable sheave 422 of the secondary pulley 42). The slide position information may be acquired.

  Thereafter, the process proceeds to step ST2, whether or not the integral term hold execution flag XFBIHLD stored in advance in the RAM 83 of the ECU 8 is OFF, that is, whether or not the integral term is held (fixed). Determine. This integral term hold execution flag XFBIHLD is turned on substantially simultaneously with the hold of the integral term when an integral term hold execution condition described later is satisfied. The integral term hold execution flag XFBIHLD is turned OFF substantially simultaneously with the release of the integral term hold when an integral term hold release condition described later is satisfied. When the vehicle starts to travel, the maximum transmission gear ratio γmax is set as the transmission gear ratio of the belt-type continuously variable transmission 4, and the integral term hold execution flag XFBIHLD is not normally satisfied. Is in an OFF state.

  If the integral term hold has not been executed (is in a non-execution state), the integral term hold execution flag XFBIHLD is in the OFF state, and if YES is determined in step ST2, the process proceeds to step ST3. In step ST3, it is determined whether or not the absolute value (| dlintgt |) of the target sheave position change amount exceeds a predetermined value A1.

  The target sheave position change amount dintgt is a target value of the slide movement amount per unit time of the movable sheave 412 of the primary pulley 41, that is, a slide movement speed (mm / sec), and is obtained according to the map shown in FIG. This is the moving speed (target moving speed) of the movable sheave 412 corresponding to the changing speed of the target gear ratio Tγ. For example, when the vehicle is in a driving state in which the accelerator operation amount Acc changes greatly or in a vehicle traveling state in which the vehicle speed V changes greatly, the change speed of the target gear ratio Tγ increases accordingly. The absolute value of (| dintgt |) also increases.

  The threshold A1 is a value set in advance by experiment or simulation, and an example thereof is 6.0 mm / sec. This value is not limited to this, but the gear shift speed (sliding movement speed of the movable sheave 412) to the extent that the deviation between the actual gear ratio Rγ and the target gear ratio Tγ is large and the integral term is excessively concerned. The target movement speed) is set in advance as a value corresponding to the increase.

  If the absolute value (| dlintgt |) of the target sheave position change amount exceeds the predetermined value A1 and YES is determined in step ST3, the process proceeds to step ST4, where the value obtained by subtracting the actual sheave position from the target sheave position is It is determined whether or not the absolute value (| lintgt−linact |), that is, the interval between the target sheave position and the actual sheave position exceeds a predetermined value B1.

  The target sheave position is determined according to the target gear ratio Tγ described above. That is, the correspondence relationship between the gear ratio and the sheave position (the position of the movable sheave 412) is stored in advance, and the target sheave position is obtained by extracting the position of the movable sheave 412 corresponding to the target gear ratio Tγ. become.

  Further, the actual sheave position is obtained from the slide position information of the movable sheave 412 from the primary sheave position sensor 111.

  The threshold value B1 is a value set in advance by experiment or simulation, and an example thereof is 0.3 mm. This value is not limited to this, but is set in advance as a value corresponding to a case where the deviation (deviation between the target sheave position and the actual sheave position) has increased to such a degree that the integral term is excessive. ing.

  If the absolute value of the value obtained by subtracting the actual sheave position from the target sheave position (| lintgt-linact |) exceeds the predetermined value B1, and YES is determined in step ST4, the integral term hold execution condition is satisfied ( The integral term hold execution condition is satisfied when YES is determined in each of step ST3 and step ST4; in the present invention, “the physical quantity correlated with the change amount of the target gear ratio is equal to or greater than a predetermined value and the actual speed change is performed. This corresponds to “when the physical quantity correlated with the deviation between the ratio and the target speed ratio is equal to or greater than a predetermined value”), and the process proceeds to step ST5, where the integral term hold execution flag XFBIHLD is set to ON. At the same time, the integral term is held (fixed). That is, during the period in which the current value of the integral term is maintained and the integral term is held, the integral term is a constant value (the integral term is independent of the deviation Δγ between the actual gear ratio Rγ and the target gear ratio Tγ). (The value at the time when the hold execution condition is satisfied).

  Hereinafter, the reason why each of step ST3 and step ST4 is set as the integral term hold execution condition will be described.

  The followability of control when the movable sheave 412 is moved to the target sheave position (the sheave position corresponding to the target gear ratio) tends to deteriorate as the target sheave position change amount dintgt increases. This is due to a response delay of the hydraulic control circuit 20 and the hydraulic actuator 413. In this case, the deviation between the target sheave position (corresponding to the target gear ratio Tγ) and the actual sheave position (corresponding to the actual gear ratio Rγ), which is the basis of the calculation of the integral term, also tends to increase. In consideration of this, in step ST3, the absolute value (| dlintgt |) of the target sheave position change amount exceeds the predetermined value A1, that is, the integral term may become excessive due to a response delay or the like. If it is determined whether there is a situation of concern, and the absolute value (| dlintgt |) of the target sheave position change amount exceeds the predetermined value A1, there is a possibility that the integral term must be held. As a result, the determination is YES and the process proceeds to the determination in step ST4.

  In step ST4, whether or not the integral term is actually excessive is determined based on whether the absolute value (| lintgt-linact |) of the value obtained by subtracting the actual sheave position from the target sheave position exceeds the predetermined value B1. Judgment is made based on whether or not it is.

  As described above, when there is a concern that the integral term becomes excessive due to a response delay or the like, and when the deviation between the target sheave position and the actual sheave position is actually large, the integral term It is determined that the hold execution condition is satisfied, and the integral term is maintained at a constant value.

  However, when the absolute value (| dlintgt |) of the target sheave position change amount does not exceed the predetermined value A1 (when NO is determined in step ST3), the value obtained by subtracting the actual sheave position from the target sheave position is Even when the absolute value (| lintgt-linact |) exceeds the predetermined value B1 (the situation determined as YES in step ST4), the integral term is not held. This is due to the fact that there is a possibility that a steady deviation of the actual gear ratio Rγ from the target gear ratio Tγ may occur due to individual differences in the belt-type continuously variable transmission 4 or deterioration over time. This is to allow the actual speed ratio Rγ to approach the target speed ratio Tγ by allowing it.

  In the determination of step ST3, when the absolute value (| dintgt |) of the target sheave position change amount does not exceed the predetermined value A1, or when it is determined NO, or in the determination of step ST4, the actual sheave position is changed from the target sheave position. If the absolute value of the value obtained by subtracting (| lintgt−linact |) does not exceed the predetermined value B1 and NO is determined, the integral term hold execution condition is not satisfied and the integral term hold is executed. And the integral term hold execution flag XFBIHLD is maintained OFF and the process returns.

  If the integral term hold is executed in the determination of step ST2 and the integral term hold execution flag XFBIHLD is in the ON state, a NO determination is made in step ST2, and the process proceeds to step ST6. In this step ST6, it is determined whether or not the following integral term hold release condition is satisfied. The integral term hold release condition is the following condition A and condition B. If at least one of the condition A and condition B is satisfied, YES is determined in step ST6.

  Condition A: The absolute value of the value obtained by subtracting the actual sheave position from the target sheave position (| lintgt-linact |) is less than a predetermined value B2 (in the present invention, “correlation with the deviation between the actual gear ratio and the target gear ratio). Equivalent to the case where a certain physical quantity is equal to or less than a predetermined value).

  Condition B: The absolute value (| dlinact |) of the actual sheave position change amount is less than a predetermined value A2, and the time during which the integral term hold execution flag XFBIHLD is maintained ON is continuous for a predetermined time T or more. (Corresponding to “when the physical quantity correlated with the change amount of the actual gear ratio is equal to or less than a predetermined value and the time during which the integral term is fixed continues for a predetermined time or more” in the present invention)

  Here, the threshold value B2 is a value set in advance by experiment or simulation, and is a value smaller than the predetermined value B1, and an example thereof is 0.3 mm. This value is not limited to this, but is set in advance as a value corresponding to a deviation (deviation between the target sheave position and the actual sheave position) to the extent that there is no concern about excess of the integral term.

  The threshold A2 is a value set in advance by experiment or simulation, and is a value smaller than the predetermined value A1, and an example thereof is 6.0 mm / sec. This value is not limited to this, but the speed change rate (the slide movement speed of the movable sheave 412; the actual movement) is such that there is no concern about the excess of the integral term due to the deviation between the actual speed ratio Rγ and the target speed ratio Tγ. It is preset as a value corresponding to (speed).

  Further, the predetermined time T is a value set in advance by experiment or simulation, and an example thereof is 0.2 sec. This value is not limited to this, but is set in advance as a time for sufficiently eliminating the hydraulic response delay, for example. Whether or not the time during which the integral term hold execution flag XFBIHLD is kept ON is continued for a predetermined time T or more is determined by a counter that counts the elapsed time since the integral term hold execution condition is satisfied. Is provided in the ECU 8, and this is performed depending on whether or not the count value of the counter has reached the predetermined time T.

  If the integral term hold release condition is satisfied and YES is determined in step ST6, the process proceeds to step ST7, where the integral term hold execution flag XFBIHLD is set to OFF. In addition, when the integral term hold release condition is satisfied, the integral term hold is released. That is, the integral term (ΣΔγ · Gi) is calculated (added or subtracted) in accordance with the deviation Δγ between the actual speed ratio Rγ and the target speed ratio Tγ.

  That is, when the condition A is satisfied, since the absolute value (| lintgt-linact |) of the value obtained by subtracting the actual sheave position from the target sheave position is less than the predetermined value B2, the integral term hold is released. However, since the integral term does not become excessive, calculation (accumulation) of the integral term is started.

  Further, when the condition B is satisfied, even if an adjustment error of the feedforward pressure Pinff occurs, after a predetermined time T has elapsed since the integral term was held (the response delay is After the cancellation, the calculation of the integral term (cumulative) is started on the condition that the absolute value (| dlinact |) of the actual sheave position change amount is less than the predetermined value A2. By increasing the value, it becomes possible to bring the actual speed ratio Rγ closer to the target speed ratio Tγ.

  If it is determined in step ST6 that the integral term hold release condition is not satisfied and the determination is NO, the integral term hold is maintained, and the integral term hold execution flag XFBIHLD is maintained on and returned. The

  The above operation is repeated, and the integral term hold when the integral term hold execution condition is satisfied and the integral term hold cancel when the integral term hold release condition is satisfied are executed.

  FIG. 8 is a timing chart showing an example of temporal changes in the target sheave position, the actual sheave position, and the integral term hold execution flag XFBIHLD. FIG. 8 shows a case where the shift control is performed in which the target sheave position change amount (the change speed of the target sheave position; corresponding to the slope of the straight line shown by the alternate long and short dash line in the figure) exceeds a predetermined value (A1). Yes.

  A shift command is transmitted at timing t1 in the figure, and the change of the target sheave position starts. In addition, there is a delay in the change amount of the actual sheave position with respect to the change amount of the target sheave position, and the target sheave position change amount exceeds a predetermined value (A1) at timing t2 in the figure. (YES determination at step ST3 in the flowchart) When the value obtained by subtracting the actual sheave position from the target sheave position exceeds a predetermined value (B1) (YES determination at step ST4 in the flowchart), the integral term hold execution condition Is established. As the integral term hold execution condition is satisfied, the integral term hold execution flag is set to ON (step ST5 in the flowchart). From this point on, the integral term is held.

  At timing t3 in the figure, the actual sheave position change amount is less than a predetermined value (A2), and the time during which the integral term hold execution flag XFBIHLD is maintained ON is continuous for a predetermined time T or longer. Therefore, the above-described condition B (integral term hold release condition) is satisfied (YES in step ST6 in the flowchart), and the integral term hold execution flag is set to OFF (step ST7 in the flowchart). At this point, the integral term hold is released.

  According to this embodiment, the following effects can be achieved.

  When the shift speed is high, there is a concern that excessive accumulation of the integral term may occur. However, in this embodiment, when the shift speed is high (when the target sheave position change amount is large), the target sheave position and the actual sheave position are actually increased. The integral term is held (fixed) on condition that the deviation from the sheave position (deviation between the target gear ratio and the actual gear ratio) is a predetermined value or more. For this reason, excessive accumulation of integral terms can be suppressed, and overshoot and undershoot of the gear ratio can be prevented. Further, since one of the conditions for holding the integral term is that the gear shift speed is high (the target sheave position change amount is large), the actual gear ratio Rγ due to individual differences of the belt-type continuously variable transmission 4 or aging deterioration. In the case where there is a steady deviation with respect to the target speed ratio Tγ, accumulation of integral terms can be allowed. As a result, the actual speed ratio Rγ can be brought close to the target speed ratio Tγ, which can contribute to an improvement in the fuel consumption rate of the engine 1.

  In the present embodiment, as one of the conditions for releasing the hold of the integral term, the period during which the integral term is held is continuous for a predetermined period or longer. For this reason, it is avoided that the cancellation | release of the operation | movement which suppresses the excessive accumulation | storage of an integral term will be performed immediately after the start of gear shifting. That is, if accumulation of the integral term is started in a state where the deviation between the actual gear ratio Rγ and the target gear ratio Tγ is large due to the presence of the hydraulic response delay period immediately after the start of the gear shift, the integral term is There is a possibility that overshooting and undershooting of the gear ratio will occur due to excessive accumulation, but in this embodiment, the period during which the integral term is held is secured for a predetermined period or more. Therefore, overshoot and undershoot of the gear ratio can be prevented.

-Other embodiments-
In the embodiment described above, the example in which the present invention is applied as a control device for a continuously variable transmission of a vehicle equipped with the gasoline engine 1 has been described. However, the present invention is not limited to this, and other diesel engines and the like are used. It can also be applied to a vehicle equipped with an engine. In addition to the engine, the vehicle power source may be an electric motor or a hybrid power source including both the engine and the electric motor.

  In the above embodiment, the continuously variable transmission 4 is an example of a belt type CVT. However, the present invention is not limited to this, and various continuously variable transmissions such as a toroidal CVT and a hydrostatic pressure CVT can be used. It can be applied to control of the gear ratio of the machine.

  The shift control according to the present invention is also applicable to the case where only feedback control is performed without combining feedforward control.

  Further, the “physical quantity correlated with the change amount of the target gear ratio” in the present invention is not limited to the change amount (dlintgt) of the target sheave position, but the target gear ratio (the target gear ratio obtained from the map of FIG. 4). Ratio) itself or the target rotational speed change amount of the movable sheave 412 of the primary pulley 41 (engine rotational speed or vehicle speed and the target speed ratio change amount). Number change amount).

  Further, the “physical quantity correlated with the deviation between the actual gear ratio and the target gear ratio” in the present invention is not limited to the deviation (lintgt-linact) between the target sheave position and the actual sheave position, and the actual gear ratio. It may be a deviation between the target speed ratio (actual speed ratio calculated from the ratio of the input shaft speed Nin and the output shaft speed Nout) or the target speed of the movable sheave 412 of the primary pulley 41. The deviation between the actual number of rotations of the sheave 412 (the number of rotations detected by the primary pulley rotation number sensor 105) and the like may be used.

  Furthermore, the “physical quantity correlated with the change amount of the actual gear ratio” in the present invention is not limited to the actual sheave position change amount (dlinact), and the change amount of the actual gear ratio itself or the movement of the primary pulley 41 is movable. It may be the amount of change in the actual rotational speed of the sheave 412.

  The present invention can be applied to shift feedback control of a CVT mounted on a vehicle.

4 Belt type continuously variable transmission 8 ECU
105 Primary pulley rotational speed sensor 106 Secondary pulley rotational speed sensor 111 Primary sheave position sensor

Claims (4)

In a control device for a continuously variable transmission that performs feedback control so that the actual gear ratio follows the target gear ratio,
When the change amount of the target speed ratio is equal to or greater than a predetermined value and the integral term fixing condition is satisfied, the integral term of the feedback control that is proportional to the integral of the deviation between the actual speed ratio and the target speed ratio is fixed. ,
When the amount of change in the actual gear ratio is equal to or less than a predetermined value and the time during which the integral term is fixed continues for a predetermined time or longer, the integral term fixed release condition is satisfied and the integral term is fixed. A control device for a continuously variable transmission, characterized by being configured to be released. The control device for a continuously variable transmission according to claim 1,
The integral term fixed release condition is satisfied when the physical quantity correlated with the change amount of the actual gear ratio is equal to or less than a predetermined value and the time during which the integral term is fixed is continued for a predetermined time or more. A control device for a continuously variable transmission. The control device for a continuously variable transmission according to claim 1 or 2,
The control unit for a continuously variable transmission is characterized in that the integral term fixed release condition is satisfied even when a physical quantity correlated with a deviation between the actual speed ratio and the target speed ratio is equal to or less than a predetermined value. The control device for a continuously variable transmission according to claim 1, 2, or 3,
The integral term fixed condition is that the physical quantity correlated with the change amount of the target gear ratio is a predetermined value or more and the physical quantity correlated with the deviation between the actual gear ratio and the target gear ratio is a predetermined value or more. A control device for a continuously variable transmission.

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