JP2010210026A - Control device for continuously variable transmission - Google Patents

Abstract

The present invention provides a control device for a continuously variable transmission that does not have insufficient driving force due to a reduction in a gear ratio and that can prevent belt slippage in an operation region in which the closing control is conventionally performed.
In a belt-type continuously variable transmission, when a vehicle starts or at a low vehicle speed, both an upshift transmission control valve 301 and a downshift transmission control valve 302 are closed to actuate a primary pulley 41 on a hydraulic actuator 413. Shut off oil supply / discharge. The hydraulic oil leakage amount from the upshift transmission control valve 301 to the hydraulic actuator 413 and the hydraulic oil leakage amount from the hydraulic actuator 413 are recognized in advance so that the hydraulic pressure does not cause belt slip or upshift. The line pressure is set and the hydraulic pressure in the hydraulic actuator 413 is managed.
[Selection] Figure 2

Description

  The present invention relates to a control device for a continuously variable transmission mounted on a vehicle. In particular, the present invention relates to measures for optimizing the gear ratio of a continuously variable transmission, for example, when the vehicle starts.

  Conventionally, in a vehicle equipped with an engine, the transmission ratio between the engine and the drive wheel is automatically set as a transmission that appropriately transmits the torque and rotation speed generated by the engine to the drive wheel according to the traveling state of the vehicle. An automatic transmission that is optimally set for the motor is known. Also known as this automatic transmission is a continuously variable transmission (CVT) that allows the gear ratio to be adjusted steplessly. One example is a belt-type continuously variable transmission.

  In this belt type continuously variable transmission, a belt is wound around a primary pulley (input pulley) having a pulley groove (V groove) and a secondary pulley (output pulley), and the width of the pulley groove of one pulley is increased. Simultaneously with the expansion, the width of the pulley groove of the other pulley is narrowed to continuously change the belt wrapping radius (effective diameter) for each pulley, thereby adjusting the transmission ratio steplessly. ing.

  The speed change in the belt-type continuously variable transmission is calculated by calculating a target speed ratio (target speed on the input side of the transmission) based on, for example, an accelerator operation amount indicating a driver's required output amount and a vehicle speed, and the target speed ratio. In order to make the actual gear ratio coincide with each other, the movable sheave of the primary pulley is moved by a hydraulic actuator provided on the back side thereof, and feedback control is performed to enlarge / reduce the groove width of the pulley groove.

  In such a belt-type continuously variable transmission, for example, as disclosed in Patent Document 1 below, a speed ratio is controlled using an upshift transmission control valve and a downshift transmission control valve provided in a hydraulic circuit. Is controlling. Line pressure is supplied to these two shift control valves as the original pressure.

  FIG. 12 is a diagram schematically showing a part of a hydraulic circuit for controlling the speed ratio of the belt type continuously variable transmission. As shown in this figure, shift control solenoids c and d are connected to the upshift transmission control valve a and the downshift transmission control valve b, respectively, and the shift is performed according to the upshift transmission command or the downshift transmission command. The control solenoids c and d are operated, and the upshift transmission control valve a and the downshift transmission control valve b are switched by the control hydraulic pressure output from the transmission control solenoids c and d. As a result, the amount of hydraulic oil (see the arrow indicated by the solid line in the figure) supplied to the hydraulic actuator f of the primary pulley e via the upshift transmission control valve a and the downshift from this hydraulic actuator f The amount of hydraulic oil (see the arrow indicated by the broken line in the figure) discharged through the shift control valve b is controlled. By controlling the amount of hydraulic oil flowing into and out of the hydraulic actuator f of the primary pulley e in this way, the groove width of the primary pulley e, that is, the winding radius of the belt g in the primary pulley e is changed, and the transmission ratio is controlled. The

  A belt clamping pressure control valve (not shown) is connected to the hydraulic actuator i of the secondary pulley h. A line pressure is supplied to the belt clamping pressure control valve, and the control pressure output from a linear solenoid valve (not shown) is controlled as a pilot pressure to be supplied to the hydraulic actuator i of the secondary pulley h, thereby the belt clamping pressure. Is controlled.

  The line pressure used for the above shift control and belt clamping pressure control is generated by adjusting the hydraulic pressure generated by the oil pump with the line pressure control valve (primary regulator valve). This line pressure control valve is configured to operate with a control hydraulic pressure output from a linear solenoid valve for line pressure control as a pilot pressure.

  By the way, in this type of belt-type continuously variable transmission, there is known one in which hydraulic control called closing control is performed when the vehicle starts or travels at low speed (see Patent Document 2 and Patent Document 3 below). ). Hereinafter, this confinement control will be described.

  For example, when the engine is stopped, since the input shaft and output shaft of the belt-type continuously variable transmission are stopped, the speed ratio of the belt-type continuously variable transmission that is the rotational speed ratio of these shafts cannot be recognized. Therefore, the gear ratio adjustment operation based on the feedback control described above cannot be executed. For this reason, when the vehicle restarts, hydraulic oil is continuously discharged from the downshift transmission control valve b so that the transmission ratio is maximized so that the transmission ratio is maximized. I am doing so. However, when the vehicle travels by traction or the like while the engine is stopped, the gear ratio may be deviated from the maximum gear ratio. For example, there is a case where a hydraulic force is generated in the hydraulic actuator f of the primary pulley e due to a centrifugal force acting on the hydraulic oil as it is pulled, and the gear ratio deviates from the maximum gear ratio. As described above, when the vehicle is restarted in a state where the gear ratio deviates from the maximum gear ratio, as described above, when the hydraulic oil is discharged from the downshift transmission control valve b, There is a possibility that the operating hydraulic pressure in the hydraulic actuator f of the pulley e is lowered and the belt g slips. In order to avoid such a situation, the closing control is performed. That is, a check valve j for reducing the line pressure and feeding it to the hydraulic actuator f of the primary pulley e is provided, and a minute hydraulic pressure is applied to the hydraulic actuator f of the primary pulley e as shown by the dashed line arrow in FIG. By causing it to act, the belt g is prevented from slipping.

  Not only towing while the engine is stopped, even if the engine stops after sudden braking of the vehicle, the engine may stop before the speed ratio reaches the maximum speed ratio, and the speed ratio becomes the maximum speed ratio. There is a possibility that the vehicle is started in a state deviated from the above. Also in this case, the same closing control as described above is performed to prevent the belt from slipping.

JP-A-63-74736 JP 2005-42799 A JP 2008-309271 A

  However, in the conventional closing control described above, a minute hydraulic pressure continues to be applied to the hydraulic actuator of the primary pulley until the vehicle speed reaches a predetermined value or more. It gradually becomes narrower and the gear ratio becomes smaller.

  For this reason, the driving force applied to the drive wheels also decreases, and for example, even when a situation in which a large driving force is required by the driver depressing the accelerator pedal (a vehicle acceleration request is made) Therefore, it is difficult to quickly respond to the driver's request because the gear ratio is not obtained. That is, in the conventional closing control, it is difficult to quickly respond to the driver's subsequent acceleration request.

  The present invention has been made in view of the above points, and the object of the present invention is to provide a belt that does not have insufficient driving force due to a reduction in the gear ratio in an operation region in which the closing control is conventionally performed. An object of the present invention is to provide a control device for a continuously variable transmission that can prevent slippage.

-Principle of solving the problem-
The solution principle of the present invention devised to achieve the above object is that the working fluid chamber in which working fluid is supplied and discharged when the gear ratio is changed when the gear ratio is in an unrecognizable traveling state. The circuit is cut off from the circuit, and the working fluid enters and exits the working fluid chamber only for leakage and leakage. In this state, by adjusting the line pressure and appropriately adjusting the leakage amount of the working fluid, the hydraulic pressure management is performed so as not to cause belt slip and upshift.

-Solution-
Specifically, the present invention includes a primary pulley provided on the input shaft, a secondary pulley provided on the output shaft, and a belt wound around the primary pulley and the secondary pulley, and includes a working fluid circuit. The groove width of the pulley is changed and the belt winding radius is changed by the supply operation of the working fluid obtained by adjusting the line pressure as the original pressure to the working fluid chamber and the discharge operation of the working fluid from the working fluid chamber. Assuming that the control device for the continuously variable transmission is configured to change the transmission gear ratio steplessly. When the rotational speed of the output shaft is less than or equal to a predetermined value with respect to the control device for the continuously variable transmission, the operation fluid supply operation to the operation fluid chamber and the discharge operation are both stopped and the operation fluid chamber is supplied to the operation fluid chamber. The working fluid chamber pressure management means for managing the working fluid pressure in the working fluid chamber within a predetermined range by adjusting the amount of leakage of the working fluid by the line pressure is provided.

  Here, the amount of leakage of the working fluid into the working fluid chamber is, for example, hydraulic oil that leaks (leaks into) the working fluid chamber from a gap between a valve body and a spool in a valve provided in the working fluid circuit. Is the amount. The predetermined range in which the working fluid pressure in the working fluid chamber is managed is a working oil pressure range in which belt slippage or upshift does not occur in the continuously variable transmission. In other words, the working fluid chamber pressure management means is configured so that the working fluid pressure in the working fluid chamber exceeds the working fluid pressure at which belt slip occurs and is less than the working fluid pressure at which the gear ratio is changed to the upshift side. Set the line pressure.

  Due to this specific matter, when the rotational speed of the output shaft is not more than a predetermined value (for example, when the vehicle starts or travels at a low speed), both the supply operation and the discharge operation of the working fluid to the working fluid chamber are stopped. As a result, the working fluid enters and exits the working fluid chamber only by the leakage and the leakage. In this state, by adjusting the line pressure, a leakage amount of the working fluid corresponding to the leakage amount of the working fluid from the working fluid chamber is obtained. Specifically, the amount of leakage of working fluid from the working fluid chamber hardly changes even if the line pressure changes, so if the line pressure is set high, the amount of working fluid leaking into the working fluid chamber Increases relatively, and the working fluid pressure in the working fluid chamber increases. Conversely, if the line pressure is set low, the amount of working fluid leaking into the working fluid chamber will be relatively small, and the working fluid pressure in the working fluid chamber will be low. In this way, by adjusting the amount of leakage of the working fluid with the line pressure, the working fluid pressure in the working fluid chamber is within a range where belt slippage or upshift does not occur in the continuously variable transmission. It becomes possible to adjust. As a result, it is possible to eliminate the check valve that was necessary when performing the conventional closing control, and it is possible to eliminate the components of the working fluid circuit, while preventing belt slippage. It is possible to improve the durability of the motor and to solve the shortage of driving force by not causing an upshift.

  As a more specific configuration of the working fluid chamber pressure management means, a variation in the amount of leakage of the working fluid into the working fluid chamber and a variation in the amount of leakage of the working fluid from the working fluid chamber (with a constant line pressure) In the case of a continuously variable transmission in which a predetermined fluctuation range exists in the actual leakage amount obtained by subtracting the leakage amount of the working fluid from the leakage amount of the working fluid due to the variation in each amount) The working fluid pressure when the actual leakage amount is the largest in the above-mentioned fluctuation range exceeds the working fluid pressure at which belt slip occurs and the actual leakage amount is the smallest in the above-mentioned fluctuation range. The line pressure is set so that the working fluid pressure in the working fluid chamber is managed within a range where the pressure is less than the working fluid pressure at which the gear ratio is changed to the upshift side.

  Here, the variation in the leakage amount and the variation in the leakage amount are not limited to variations in the same continuously variable transmission depending on its operating state, environmental conditions, etc., but each of a plurality of continuously variable transmissions having the same configuration. It is a concept that includes variations caused by individual differences (such as variations in processing dimensions). Here, the leakage amount of the working fluid from the working fluid chamber is, for example, the amount of working oil that leaks to the outside of the working fluid chamber from a gap between the movable sheave of the pulley and the cylinder member.

  In this way, the amount of working fluid leaked into the working fluid chamber is adjusted by the line pressure, so that the actual leakage amount is any amount (any amount within the fluctuation range), that is, Even if the actual leakage amount fluctuates due to variations in the leakage amount of the working fluid into the working fluid chamber or variations in the leakage amount of the working fluid from the working fluid chamber, the working fluid pressure at which belt slip occurs is changed. It is possible to set the working fluid pressure in the working fluid chamber to be greater than the working fluid pressure that is changed to the upshift side and to prevent both belt slip and upshift.

  Examples of the working fluid chamber pressure management means include the following configurations. In other words, due to variations in the amount of working fluid leaked into the working fluid chamber and variations in the amount of working fluid leaked from the working fluid chamber (variations in amounts when the line pressure is constant), The actual leakage amount is the largest of the above fluctuation ranges for a continuously variable transmission in which a predetermined fluctuation range exists in the actual leakage amount obtained by subtracting the leakage amount of the working fluid from the leakage amount of the working fluid. If the line pressure is set to exceed the working fluid pressure at which belt slip occurs, the change in the gear ratio as the vehicle travels is recognized. A line pressure learning operation is performed to decrease the predetermined amount.

  In this case, after the line pressure has been reduced by a predetermined amount by executing the line pressure learning operation, the change in the gear ratio as the vehicle travels is recognized again. It is preferable to further reduce the line pressure in a stepwise manner by this line pressure learning operation until it is further reduced by a fixed amount and no change in the transmission gear ratio accompanying the running of the vehicle occurs.

  According to this, first, when the gear ratio becomes small when the vehicle travels after setting the line pressure to obtain a working fluid pressure exceeding the working fluid pressure at which belt slip occurs, the set value of the line pressure is high. Therefore, it can be determined that the working fluid pressure in the working fluid chamber has become high enough to cause an upshift. For this reason, the working fluid pressure in the working fluid chamber is lowered by reducing the line pressure by a predetermined amount. By repeating such a line pressure learning operation, the working fluid pressure in the working fluid chamber can be set to a value that does not cause an upshift. In this way, the working fluid pressure in the working fluid chamber at the time when the upshift is not changed to the state where the upshift is not caused is set to be relatively high, and belt slip does not occur. In other words, the line pressure that is initially set in this learning operation is set to a value that provides a working fluid pressure that exceeds the working fluid pressure at which belt slip occurs, and the operation in the working fluid chamber that does not cause an upshift. The line pressure learning operation ends when the fluid pressure is reached. For this reason, even if this line pressure learning operation is repeated, the line pressure does not fall below the working fluid pressure at which belt slip occurs, and belt slip does not occur.

  This solution means that the actual leakage amount exceeds the working fluid pressure at which belt slip occurs when it is assumed that the actual leakage amount is the largest in the fluctuation range, and the actual leakage amount is the smallest in the fluctuation range. This is effective when there is no line pressure in a range where the gear ratio is less than the working fluid pressure that is changed to the upshift side (when there is no line pressure setting solution).

  In the present invention, when the output shaft rotational speed of the continuously variable transmission is less than or equal to a predetermined value, the working fluid chamber to which the working fluid is supplied and discharged is shut off from the working fluid circuit, and leakage of the working fluid in the working fluid chamber is leaked. The working fluid pressure in the working fluid chamber is managed by adjusting the line pressure only for the intake and leakage. For this reason, the working fluid pressure in the working fluid chamber can be adjusted to a range in which belt slippage or upshift does not occur in the continuously variable transmission.

1 is a schematic configuration diagram of a vehicle equipped with a belt-type continuously variable transmission according to an embodiment. It is a circuit block diagram which shows the hydraulic control circuit which controls the hydraulic actuator of a primary pulley. It is a circuit block diagram which shows the hydraulic control circuit which controls the hydraulic actuator of a secondary pulley. 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 figure which shows the relationship between the line pressure and primary pressure for demonstrating the setting method of the line pressure in 1st Embodiment. It is a figure which shows the relationship between the line pressure and primary pressure for demonstrating the control method of the line pressure in 2nd Embodiment. It is a flowchart figure which shows the driving | running | working time integration routine in the complete confinement control which concerns on 2nd Embodiment. It is a flowchart figure which shows the line pressure control routine in the complete confinement control which concerns on 2nd Embodiment. It is a timing chart figure which shows an example of the change of the gear ratio at the time of complete closing control concerning a 2nd embodiment, and the change of vehicle speed. It is a figure which shows typically a part of hydraulic circuit for controlling a gear ratio in the conventional belt-type continuously variable transmission.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. In this embodiment, a case where the present invention is applied to a belt type continuously variable transmission mounted on a vehicle will be described.

  FIG. 1 is a schematic configuration diagram of a vehicle according to the present embodiment. As shown in FIG. 1, the vehicle of the present embodiment is an FF (front engine / front drive) type vehicle, which is an engine (internal combustion engine) 1 that is a power source for traveling, and a torque converter 2 that is a fluid transmission device. , A forward / reverse switching device 3, a belt type continuously variable transmission (CVT) 4, a reduction gear device 5, a differential gear device 6, an ECU (Electronic Control Unit) 8, and the like are mounted. A control device for the belt type continuously variable transmission is realized by the hydraulic control circuit (working fluid circuit) 20, the primary pulley rotation speed sensor 105, the secondary pulley rotation speed sensor 106, and the like.

  A crankshaft 11, which 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 (not shown).

  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 supplied 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. Further, when both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 3 becomes neutral (interrupted state) for interrupting power transmission.

-Belt type continuously variable transmission-
The belt-type continuously variable transmission 4 includes an input-side primary pulley 41, an output-side secondary pulley 42, a metal belt 43 wound around the primary pulley 41 and the secondary pulley 42, and the like.

  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.

  On the movable sheave 412 side of the primary pulley 41, a hydraulic actuator (hydraulic actuator having a working fluid chamber in the present invention) 413 for changing the V groove width between the fixed sheave 411 and the movable sheave 412 is arranged. Yes. 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-described 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 shift speed 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 control unit. It is constituted by a valve 20f or the like.

  Further, a shift control solenoid (DS1) 304 and a shift control solenoid (DS2) 305, a linear solenoid (SLS) 202 for belt clamping pressure control, and a linear solenoid for line pressure control, which constitute the hydraulic pressure control circuit 20. A control signal from the ECU 8 is supplied to the (SLT) 201 and the duty solenoid (DSU) 307 for controlling the lockup engagement pressure.

  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 speed 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.

  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, by switching the clutch apply control valve 204, the control hydraulic pressure from the linear solenoid (SLS) 202 is supplied to the primary regulator valve 203, and the line pressure PL may be 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 adjusted 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. A 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 (DS1) 304 and a shift control solenoid (DS2) to be described later. ) 305 and the belt clamping pressure control valve 303 and the like.

[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. 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 (DS1) 304 that outputs a control hydraulic pressure in accordance with a current value determined by a duty signal (DS1 shift duty (upshift duty)) output from the ECU 8. The control hydraulic pressure output from the shift control solenoid (DS1) 304 is applied to the first hydraulic pressure port 315. The second hydraulic pressure port 316 is connected to a shift control solenoid (DS2) 305 that outputs a control hydraulic pressure according to a current value determined by a duty signal (DS2 shift duty (downshift duty)) output from the ECU 8. The control hydraulic pressure output from the shift control solenoid (DS2) 305 is applied to the second hydraulic 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 shift control solenoid (DS1) 304 is connected to the first hydraulic port 326, and the control hydraulic pressure output from the shift control solenoid (DS1) 304 is applied to the first hydraulic port 326. The shift control solenoid (DS2) 305 is connected to the second hydraulic port 327, and the control hydraulic pressure output from the shift control solenoid (DS2) 305 is applied to the second hydraulic port 327.

  Further, the downshift transmission control valve 302 is formed with an input / output port 324 and a discharge port 325. In the 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.

  2, the shift control solenoid (DS1) 304 operates in response to the DS1 shift duty (upshift shift command) output from the ECU 8, and the control hydraulic pressure output from the shift control solenoid (DS1) 304. Is supplied to the first hydraulic pressure port 315 of the upshift transmission control valve 301, the spool 311 is moved 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 (DS1) 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, the shift control solenoid (DS2) 305 is operated in response to the DS2 shift duty (downshift shift command) output from the ECU 8, and the control hydraulic pressure output from the shift control solenoid (DS2) 305 is the upshift shift control valve 301. Is supplied to the second hydraulic pressure port 316, the spool 311 is moved to the downshift position side (lower side in FIG. 2) by a thrust according to the control hydraulic 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. The hydraulic fluid 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 reduced, the V groove width of the input side variable pulley 42 is increased, and the transmission ratio γ is increased (downshift).

  When the control hydraulic pressure output from the shift control solenoid (DS2) 305 is supplied to the second hydraulic 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 described above, when the control hydraulic pressure is output from the shift control solenoid (DS1) 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. Upshifted. Further, when the control hydraulic pressure is output from the shift control solenoid (DS2) 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. Downshifted.

  In this example, as shown in FIG. 4, for example, the input 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 larger 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.

[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 cylinder 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. This is done by controlling the pressure of the oil. 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.

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

  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.

  The 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, A CVT oil temperature sensor 108, a brake pedal sensor 109, and a lever position sensor 110 for detecting a lever position (operation position) of the shift lever 9 are connected. The output signals of the respective sensors, that is, the engine 1 rotation speed (engine rotation speed) Ne, the throttle valve 12 throttle opening θth, the engine 1 cooling water temperature Tw, and the turbine shaft 27 rotation speed (turbine rotation 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 degree) Acc, oil temperature of hydraulic control circuit 20 (CVT oil temperature Thc) ), Signals indicating the presence / absence of operation of the foot brake as a service brake (brake ON / OFF) and the lever position (operation position) of the shift lever 9 are 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 lever position sensor 110 is, for example, a parking position “P”, a reverse position “R”, a neutral position “N”, a drive position “D”, a manual position “M”, an upshift position, a downshift position, or a range position. A plurality of ON / OFF switches for detecting that the shift lever 9 is operated are provided. In order to change the gear ratio γ manually, a downshift switch, an upshift switch, or a lever can be provided on the steering wheel or the like separately from the shift lever 9.

Then, the ECU 8 controls the output of the engine 1, the shift speed control and belt clamping pressure control of the belt-type continuously variable transmission 4, and the lock-up clutch 24 based on the output signals of the various sensors described above. Execute release / 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.

-Complete confinement control-
Next, a plurality of embodiments regarding complete confinement control, which is a control characteristic of this embodiment, will be described. This complete confinement control is control performed when the vehicle starts or when the vehicle speed is extremely low (when the rotational speed of the output shaft 44 is a predetermined value or less). In other words, before starting the vehicle (when the vehicle is stopped), the primary pulley 41 cannot detect the rotation speed of the primary pulley 41 (or the rotation speed of the input shaft 40 can be detected by the turbine rotation speed sensor 104), and the secondary pulley. The rotational speed of the secondary pulley 42 cannot be detected by the rotational speed sensor 106. For this reason, the gear ratio cannot be calculated (the winding radius of the belt 43 cannot be specified). Further, when the vehicle is running at a very low vehicle speed, the primary pulley rotation speed sensor 105 cannot sufficiently detect the rotation speed of the primary pulley 41 (or the rotation speed detection of the input shaft 40 by the turbine rotation speed sensor 104), and the secondary pulley The accuracy of the rotation speed detection of the secondary pulley 42 by the pulley rotation speed sensor 106 cannot be obtained sufficiently. For this reason, sufficient calculation accuracy of the gear ratio cannot be obtained.

  In such a situation, in the conventional closing control, a minute hydraulic pressure is continuously applied to the hydraulic actuator of the primary pulley until the vehicle speed reaches a predetermined value or more (shown by a one-dot chain line in FIG. 12). As shown in the arrows), the gear ratio is reduced by this hydraulic pressure, and sufficient driving force is not applied to the driving wheels, making it difficult to quickly respond to the driver's acceleration request and the like.

  Therefore, in the present embodiment, complete confinement control as described below is performed. In this complete closing control, both the supply operation and the discharge operation of the hydraulic oil to the hydraulic actuator 413 of the primary pulley 41 are stopped. Then, the leakage amount of hydraulic oil to the hydraulic actuator 413 of the primary pulley 41 and the leakage amount of hydraulic oil from the hydraulic actuator 413 (leakage amount and leakage due to leakage of hydraulic oil from the location to be sealed if originally intended) The line pressure PL is controlled according to the amount of discharge. Each embodiment will be specifically described below.

(First embodiment)
When the rotational speed of the secondary pulley 42 detected by the secondary pulley rotational speed sensor 106 increases from “0 (vehicle stopped state)” to the rotational speed increase (start of the vehicle), and the rotational speed of the secondary pulley 42 is predetermined. When the rotational speed is equal to or lower than the number of revolutions (for example, 5 km / h or lower when converted to a vehicle speed), the complete closing control is started. The value of this vehicle speed is not limited to this.

  When the complete closing control is started, first, both control hydraulic pressures output from the shift control solenoid (DS1) 304 and the shift control solenoid (DS2) 305 are set to “0”. That is, the duty ratio of the duty signal output from the ECU 8 is set to “0”, and the control hydraulic pressure is not output from the shift control solenoid (DS1) 304 and the shift control solenoid (DS2) 305.

  As a result, in the upshift transmission control valve 301, the spool 311 is in the closed position (the left position in FIG. 2) by the biasing force of the spring 312 and the input port 313 is closed. That is, the supply of the line pressure PL to the hydraulic actuator 413 is interrupted. In the downshift transmission control valve 302, the spool 321 is moved to the closed position (the right position in FIG. 2) by the urging force of the spring 322, and the input / output port 324 is closed. That is, the discharge of hydraulic oil from the hydraulic actuator 413 is blocked.

  In this way, the hydraulic oil is not supplied to or discharged from the hydraulic actuator 413. In this case, inflow / outflow of the hydraulic oil in the hydraulic actuator 413 is caused by leakage from a gap between the valve body of the upshift transmission control valve 301 and the spool 311, the movable sheave 412 of the primary pulley 41, the cylinder member (the hydraulic pressure described above) Only leakage from the gap with the hydraulic oil chamber of the actuator 413).

  The amount of hydraulic fluid that leaks into the hydraulic actuator 413 depends on the valve body of the upshift transmission control valve 301 and the machining tolerance of the spool 311. In other words, assuming that the line pressure is constant, if the machining tolerance is relatively large and the gap between the two is large, the amount of hydraulic fluid leaked per unit time will increase, When the gap is small, the amount of hydraulic oil leaked per unit time is small. On the other hand, the amount of hydraulic fluid leaking from the hydraulic actuator 413 depends on the movable sheave 412 of the primary pulley 41 and the machining tolerance of the cylinder member. In other words, when the machining tolerance is relatively large and the gap between the two is large, the amount of hydraulic fluid leaking per unit time increases, and when the gap between the two is small, the hydraulic fluid per unit time is large. The amount of leakage is reduced. The leakage of hydraulic oil to the hydraulic actuator 413 acts in a direction to increase the internal pressure (hydraulic pressure) of the hydraulic actuator 413. Conversely, the leakage of hydraulic oil to the hydraulic actuator 413 decreases the internal pressure of the hydraulic actuator 413. Act on.

  In the complete confinement control in this embodiment, the hydraulic oil leakage amount corresponding to the leakage amount is adjusted by the line pressure, and the internal pressure of the hydraulic actuator 413 is managed within a predetermined range. ing.

  The amount of leakage of the hydraulic oil is substantially constant according to the gap between the movable sheave 412 of the primary pulley 41 and the cylinder member as described above. That is, even if the line pressure changes, the leakage amount of the hydraulic oil hardly changes. On the other hand, the leakage amount of the hydraulic oil is greatly influenced not only by the size of the gap between the valve body of the upshift transmission control valve 301 and the spool 311 but also by the line pressure. It also changes depending on the pressure level. In other words, even if the gap is constant, the amount of hydraulic fluid leaking increases as the line pressure increases. That is, by adjusting the line pressure, the amount of hydraulic oil leaking into the hydraulic actuator 413 can be adjusted.

  In the complete closing control in the present embodiment, in view of this point, the relative amount of the hydraulic oil leakage amount with respect to the hydraulic oil leakage amount is adjusted by the line pressure, whereby the internal pressure of the hydraulic actuator 413 is set to a predetermined value. It is possible to manage within the range (management of the working fluid pressure by the working fluid chamber pressure management means).

  More specifically, the amount of leakage of the hydraulic fluid is recognized in advance through experiments, simulations, etc., and the relationship between the line pressure and the amount of hydraulic fluid leakage is recognized, and the leakage of hydraulic fluid is determined. The line pressure is set so that the difference between the discharge amount and the leak amount can manage the internal pressure of the hydraulic actuator 413 within a predetermined range.

  Hereinafter, a method for setting the line pressure will be described.

  When a plurality of belt-type continuously variable transmissions 4 of the same type are manufactured, the machining dimensions of each part are within the above-described machining tolerance. However, the leakage of hydraulic oil described above occurs when the machining tolerance is machined to the maximum value (machining dimension where the gap is the largest) and when the machining tolerance is machined to the minimum value (machining dimension where the gap is the smallest). The amount to be dispensed will be different. Moreover, even if the line pressure is the same, the amount of hydraulic fluid leaking will be different.

  For example, in the case of the hydraulic oil leakage amount, when the clearance between the movable sheave 412 of the primary pulley 41 and the cylinder member is processed to the minimum processing tolerance, the leakage amount is also minimized, but this clearance is the processing tolerance. In the case where the maximum value is processed, the amount of leakage becomes the maximum. Further, the amount of leakage varies depending on the viscosity of the hydraulic oil. That is, the amount of leakage changes depending on the temperature of the hydraulic oil.

  Similarly, in the case of the hydraulic oil leakage amount, when the clearance between the valve body of the upshift transmission control valve 301 and the spool 311 is processed to the minimum processing tolerance, the leakage amount is also minimum. However, when this gap is machined to the maximum machining tolerance, the amount of leakage is the largest. The amount of leakage also varies depending on the viscosity of the hydraulic oil. That is, the amount of leakage changes depending on the temperature of the hydraulic oil.

  As described above, the amount of hydraulic oil that leaks and the amount of leakage varies depending not only on the individual differences of each belt-type continuously variable transmission 4 but also on the operating temperature of the continuously variable transmission 4. Varies depending on temperature (viscosity of hydraulic fluid).

  A straight line L1 in FIG. 7 is the belt type continuously variable transmission 4 that is assumed to have the smallest leakage amount (leakage amount−leakage amount) from the hydraulic actuator 413 among the belt type continuously variable transmission 4 to be manufactured. The relationship between the line pressure PL and the primary pressure Pin (internal pressure of the hydraulic actuator 413) in a situation where the amount of leakage due to the influence of the viscosity of the hydraulic oil is the smallest is shown. When the leakage amount is small as described above, the increase rate of the primary pressure Pin with respect to the increase amount of the line pressure PL is increased. On the other hand, a straight line L2 in FIG. 7 is a belt type continuously variable transmission that is assumed to have the largest amount of leakage (leakage amount−leakage amount) from the hydraulic actuator 413 among the belt type continuously variable transmission 4 to be manufactured. 4 shows the relationship between the line pressure PL and the primary pressure Pin in a situation where the amount of leakage due to the influence of the viscosity of the hydraulic oil is also greatest. Thus, when the amount of leakage is large, the rate of increase of the primary pressure Pin with respect to the amount of increase of the line pressure PL is small.

  That is, the leakage amount (leakage amount−leakage amount) from the hydraulic actuator 413 in each of the plurality of belt-type continuously variable transmissions 4 to be manufactured exists within the range of the straight line L1 and the straight line L2. .

  Further, a straight line M2 indicated by a one-dot chain line in FIG. 7 is a lower limit value of the primary pressure Pin at which no belt slip occurs. That is, when the primary pressure Pin is lower than the pressure of the straight line M2, the belt 43 slips. A straight line M1 indicated by a one-dot chain line in FIG. 7 is an upper limit value of the primary pressure Pin at which no upshift occurs in the belt type continuously variable transmission 4. In other words, when the primary pressure Pin exceeds the pressure of the straight line M1, the upshift occurs, which leads to a situation where it is difficult to obtain a sufficient driving force for the driving wheels.

  In the present embodiment, even if the belt type continuously variable transmission 4 is assumed to have the least amount of leakage from the hydraulic actuator 413, no upshift occurs and the amount of leakage from the hydraulic actuator 413 is the smallest. Even in the belt-type continuously variable transmission 4 that is assumed to be large, a range indicated by dots in the figure is set as a range in which belt slip does not occur, and the line pressure is set within this range.

  Specifically, the straight line L1 and the straight line M1 are used as limit line pressures at which no upshift occurs even in the belt-type continuously variable transmission 4 (straight line L1) that is assumed to have the least amount of leakage from the hydraulic actuator 413. Is set to the line pressure PL2 corresponding to the intersection point X. Further, even if the belt type continuously variable transmission 4 (straight line L2) is assumed to have the largest amount of leakage from the hydraulic actuator 413, the intersection of the straight line L2 and the straight line M2 as a limit line pressure at which no belt slip occurs. A line pressure PL1 corresponding to Y is set. As a result, the line pressure set in the fully closed control is set within the range of PL1 to PL2 in the figure.

  In this case, the higher the line pressure PL is set, the longer the driving time of the oil pump 7 is, and there is a possibility that the fuel consumption rate of the engine 1 is deteriorated. Therefore, it is as low as possible in the above range (PL1 to PL2). It is preferable to set the line pressure. In other words, the set line pressure is preferably set to PL1 in the figure, but when this line pressure PL1 is set, the belt type continuously variable transmission 4 (with the largest amount of leakage from the hydraulic actuator 413). If the line pressure is slightly reduced, belt slippage occurs (below the straight line M2). Therefore, in practice, it is preferable to set the line pressure slightly higher than the line pressure PL1 so that belt slip does not occur even if the line pressure fluctuates somewhat.

  As described above, in this embodiment, belt slippage or upshift can be prevented by managing the hydraulic pressure of the hydraulic actuator 413 by the line pressure when the vehicle starts or at an extremely low vehicle speed. For this reason, it becomes possible to eliminate the check valve (check valve indicated by the symbol j in FIG. 12) that has been necessary when performing the conventional closing control, while reducing the number of components of the hydraulic control circuit 20. Further, it is possible to improve the durability of the continuously variable transmission 4 by not causing belt slip and to solve the shortage of driving force by not causing an upshift.

(Second Embodiment)
Next, a second embodiment will be described. In the present embodiment, a line pressure learning operation (hereinafter also referred to as line pressure learning) is executed in order to obtain the same effect as that of the first embodiment described above. In this line pressure learning, first, the vehicle is run in a state where the line pressure exceeds the working fluid pressure at which belt slip occurs, and a change in the gear ratio due to the running of the vehicle is recognized, so that the gear ratio decreases. If this occurs (when upshifting), the line pressure at the next vehicle start-up or the like is reduced by a predetermined amount. This will be specifically described below.

  FIG. 8 is a diagram corresponding to FIG. 7, and the straight line L <b> 1 indicates that the leakage amount (leakage amount−leakage amount) from the hydraulic actuator 413 is the smallest in the belt type continuously variable transmission 4 to be manufactured. In the assumed belt type continuously variable transmission 4, the relationship between the line pressure PL and the primary pressure Pin in a situation where the amount of leakage due to the effect of the viscosity of the hydraulic oil is the smallest is shown. The straight line L2 is the largest amount of leakage due to the effect of the viscosity of the hydraulic oil in the belt-type continuously variable transmission 4 that is assumed to have the largest amount of leakage from the hydraulic actuator 413 among the belt-type continuously variable transmission 4 to be manufactured. The relationship between the line pressure PL and the primary pressure Pin in many situations is shown. Further, a straight line M2 indicated by a one-dot chain line in FIG. 8 is a lower limit value of the primary pressure Pin at which belt slip does not occur. Further, a straight line M1 indicated by a one-dot chain line is an upper limit value of the primary pressure Pin at which no upshift occurs in the belt type continuously variable transmission 4.

  8, the line corresponding to the intersection (Y in the figure) between the straight line L2 and the straight line M2 with respect to the line pressure PL3 corresponding to the intersection (X in the figure) between the straight line L1 and the straight line M1. The pressure PL4 is located on the high pressure side. In other words, even if the belt type continuously variable transmission 4 that is assumed to have the least amount of leakage from the hydraulic actuator 413 is set to PL3 so that no upshift occurs, the actual belt type If the stage transmission 4 has the largest amount of leakage from the hydraulic actuator 413, belt slipping may occur. Similarly, even if the belt type continuously variable transmission 4 that is assumed to have the largest amount of leakage from the hydraulic actuator 413 is set to PL4 so that belt slip does not occur, the actual belt type If the continuously variable transmission 4 has the smallest amount of leakage from the hydraulic actuator 413, an upshift may occur. In such a situation, for example, the belt type continuously variable transmission 4 manufactured by setting the processing tolerance to be relatively large (setting the inclination angle of the straight line L1 to be large and the inclination angle of the straight line L2 to be small). It is easy to occur in. In this case, the line pressure range shown in FIG. 7 described in the first embodiment (the range with dots in FIG. 7) does not exist, and the line pressure setting solution (set line pressure ranges PL1 to PL1 obtained from the top of FIG. 8). PL2) does not exist. In the present embodiment, even in such a case, line pressure learning described later is performed in order to appropriately set the line pressure. Hereinafter, the line pressure learning will be specifically described.

  FIG. 9 and FIG. 10 are flowcharts showing the control procedure of this line pressure learning. FIG. 9 is a travel time integration routine for determining whether or not the execution condition of the line pressure learning is satisfied. For example, every few milliseconds). FIG. 10 shows a line pressure control routine when line pressure learning is executed.

  First, in step ST1 of FIG. 9, it is determined whether or not the vehicle has stopped (the vehicle speed has reached 0 km / h). Until the vehicle is temporarily stopped, a NO determination is made in step ST1, and this routine is temporarily ended. Then, when the vehicle is temporarily stopped and YES is determined in step ST1, the process proceeds to step ST2, and it is determined whether or not the vehicle has started after stopping (Nout> 0). While the vehicle is stopped, NO is determined in step ST2. Then, when the vehicle starts and YES is determined in step ST2, the process proceeds to step ST3, in which it is determined whether or not the vehicle speed is less than a predetermined vehicle speed αkm / h (for example, 5 km / h) (including the case where the vehicle stops again). Is done. If the vehicle speed is less than the predetermined vehicle speed αkm / h, a YES determination is made in step ST3, and the process proceeds to step ST4 to include the duration time of this low vehicle speed state (including the duration time in the stopped state if the vehicle stops again). ) Is started. Alternatively, the accumulation of the travel distance in the low vehicle speed state may be started. On the other hand, when the vehicle speed is equal to or higher than the predetermined vehicle speed αkm / h, NO is determined in step ST3, the process proceeds to step ST5, and the integrated value of the duration (or the integrated travel distance in the low vehicle speed state) is cleared.

  That is, in this flowchart, after the vehicle has started, the duration (or travel distance) of the low vehicle speed state (or the stopped state when the vehicle stops again) during the period in which the vehicle speed is continuously less than α km / h. Accumulation is performed, and the operation is continued until the vehicle speed reaches α km / h or more.

  On the other hand, in step ST11 of FIG. 10, the duration of the low vehicle speed state (or the accumulated travel distance in the low vehicle speed state) integrated in the flowchart of FIG. 9 is equal to or greater than a certain value (for example, the duration of the low vehicle speed state is 5 seconds). It is determined whether or not the cumulative travel distance in the low vehicle speed state is 10 m or more (these values are not limited to this). If the duration of the low vehicle speed state (or the accumulated travel distance in the low vehicle speed state) is equal to or greater than a predetermined value and YES is determined in step ST11, the process proceeds to step ST12, where the gear ratio γ before the vehicle stops is set. It is determined whether or not the difference between the vehicle speed and the transmission gear ratio γ is equal to or greater than a certain value (transmission ratio is 0.3: this value is not limited to this). That is, when the vehicle is traveling at a low vehicle speed, it is determined whether or not the gear ratio has changed by a predetermined value or more on the upshift side.

  If this gear ratio has changed by more than a predetermined value on the upshift side (when an upshift occurs due to the line pressure set in the line pressure learning being too high), a YES determination is made in step ST12, and the process proceeds to step ST13. Then, it is determined whether or not the current set value of the line pressure is not a preset lower limit value. The lower limit value of the line pressure is set to prevent the line pressure from excessively decreasing due to mislearning. For example, the lower limit value of the line pressure is set to a value that can sufficiently obtain the belt clamping pressure in the belt clamping pressure control described above. .

  If the current set value of the line pressure is not the lower limit value and YES is determined in step ST13, the process proceeds to step ST14, where the set pressure of the line pressure is changed from the current line pressure to one level (for example, to the current line pressure). Lower by 10%). As a result, the line pressure set at the next start of the vehicle is set lower than the current line pressure. That is, the line pressure is changed in the direction in which the upshift is eliminated.

  On the other hand, if NO is determined in any of the above steps ST11, ST12, ST13, the current line pressure is maintained as the line pressure at the next start of the vehicle without changing the line pressure. In particular, if NO is determined in step ST12, no upshift occurs, and the line pressure is appropriately set by line pressure learning.

  By repeating the line pressure learning to change the line pressure in this way, the working fluid pressure in the hydraulic actuator 413 can be converged to a value that does not cause an upshift. For example, in the case of the belt type continuously variable transmission 4 with the least amount of leakage from the hydraulic actuator 413, first, the line pressure at the start of line pressure learning is set to PL4 in FIG. The primary pressure Pin at this time is at a point Z in the figure. In this case, an upshift occurs as the vehicle travels. Then, by performing the line pressure learning a plurality of times, the line pressure is converged to PL3 in FIG.

  Further, in this line pressure learning, the working fluid pressure in the hydraulic actuator 413 at the time when the upshift does not occur from the state where the upshift has occurred is set to be relatively high in a range where the upshift does not occur. As a result, belt slip does not occur.

  An example of the change in the gear ratio and the vehicle speed in the above-described line pressure learning is shown in FIG. In this timing chart, the vehicle temporarily stops at timing T1 (timing determined as YES in step ST1) and starts immediately after that (timing determined as YES in step ST2). The gear ratio at this time is set to γ1 in the figure.

  Then, the low vehicle speed state continues, and an upshift occurs in which the gear ratio becomes small at timing T2 (timing determined as YES in step ST12). The gear ratio at this time is γ2 in the figure. By recognizing this upshift, the line pressure setting pressure is lowered by one step from the current line pressure, and the line pressure set at the next start of the vehicle is set lower than the current line pressure. Become.

  As described above, according to the line pressure learning according to the present embodiment, even if the above-described line pressure setting solution (the line pressure ranges PL1 to PL2 obtained as in the first embodiment) does not exist. The line pressure can be set appropriately, the durability of the continuously variable transmission 4 can be improved by not causing belt slip, and the lack of driving force can be eliminated by not causing an upshift. Further, according to the line pressure learning of the present embodiment, it is possible to further increase the line pressure setting accuracy.

(Other embodiments)
In each of the embodiments described above, an example in which the present invention is applied to a continuously variable transmission provided with a shift control solenoid (DS1) 304 for upshifting and a shift control solenoid (DS2) 305 for downshifting has been shown. However, the present invention is not limited to this, and can also be applied to shift control of a continuously variable transmission equipped with other shift control means not including these shift control solenoids DS1 and DS2.

  In the second embodiment described above, when the line pressure is reduced by line pressure learning, the line pressure set at the next vehicle start is set lower than the current line pressure. The present invention is not limited to this, and when it is determined that the line pressure needs to be reduced by the line pressure learning (when an upshift occurs), the line pressure may be immediately reduced. However, in this case, since belt slip may occur when the line pressure is rapidly reduced, it is preferable to reduce the line pressure at a relatively low speed.

  Further, in each of the above embodiments, the case where the present invention is applied to a control device for a continuously variable transmission of a vehicle equipped with a gasoline engine has been described. However, the present invention is not limited to this, and other engines such as a diesel engine can be used. The present invention can also be applied to a control device for a continuously variable transmission of a vehicle equipped with an engine. In addition to the engine (internal combustion engine), the vehicle power source may be an electric motor, or a hybrid power source including both the engine and the electric motor.

  The present invention is not limited to FF (front engine / front drive) type vehicles, but can be applied to FR (front engine / rear drive) type vehicles and four-wheel drive vehicles.

  The present invention can be applied to the hydraulic control for the hydraulic actuator of the primary pulley when the vehicle equipped with the belt type continuously variable transmission starts or runs at a low speed.

4 Belt type continuously variable transmission 40 Input shaft 41 Primary pulley 413 Hydraulic actuator (working fluid chamber)
42 Secondary pulley 43 Belt 44 Output shaft 20 Hydraulic control circuit (working fluid circuit)
106 Secondary pulley speed sensor PL Line pressure

Claims (5)

A primary pulley provided on the input shaft, a secondary pulley provided on the output shaft, and a belt wound around these primary pulley and secondary pulley, and the line pressure, which is the original pressure of the working fluid circuit, is adjusted. By changing the groove width of the pulley and changing the belt winding radius by the supply operation of the pressurized working fluid to the working fluid chamber and the discharge operation of the working fluid from the working fluid chamber, the speed change ratio is stepless. In the control device for a continuously variable transmission configured to be changed to
When the rotation speed of the output shaft is equal to or less than a predetermined value, the amount of working fluid leaked into the working fluid chamber is determined as the line pressure with both the supply and discharge operations of the working fluid to the working fluid chamber stopped. A control device for a continuously variable transmission, comprising: a working fluid chamber pressure management means for managing the working fluid pressure in the working fluid chamber within a predetermined range by adjusting the pressure of the working fluid chamber. In the control device for the continuously variable transmission according to claim 1,
The working fluid chamber pressure management means controls the line pressure so that the working fluid pressure in the working fluid chamber exceeds the working fluid pressure at which belt slippage occurs and the gear ratio is changed to the upshift side. A control device for a continuously variable transmission, characterized in that: In the control device for continuously variable transmission according to claim 2,
Due to the variation in the leakage amount of the working fluid into the working fluid chamber and the variation in the leakage amount of the working fluid from the working fluid chamber, the leakage amount of the working fluid is subtracted from the leakage amount of the working fluid. For a continuously variable transmission with a specified fluctuation range in the actual leakage amount,
The working fluid chamber pressure management means is configured such that the working fluid pressure when the actual leak amount is the largest in the fluctuation range exceeds the working fluid pressure at which belt slip occurs, and the actual leak amount varies as described above. The line pressure is set so that the working fluid pressure in the working fluid chamber is managed within a range in which the working fluid pressure in the smallest width is less than the working fluid pressure at which the gear ratio is changed to the upshift side. A control device for a continuously variable transmission, characterized by being configured. In the control device for continuously variable transmission according to claim 2,
Due to the variation in the leakage amount of the working fluid into the working fluid chamber and the variation in the leakage amount of the working fluid from the working fluid chamber, the leakage amount of the working fluid is subtracted from the leakage amount of the working fluid. For a continuously variable transmission with a specified fluctuation range in the actual leakage amount,
The working fluid chamber pressure management means sets the line pressure so as to exceed the working fluid pressure at which belt slip occurs when the actual leakage amount is assumed to be the largest in the fluctuation range, and then the vehicle travels. A continuously variable transmission that is configured to perform a line pressure learning operation for recognizing a change in a gear ratio associated with the pressure ratio and lowering the line pressure by a predetermined amount when the gear ratio decreases. Control device. In the control device for continuously variable transmission according to claim 4,
The working fluid chamber pressure management means recognizes again the change in the gear ratio as the vehicle travels after the line pressure is reduced by a predetermined amount by executing the line pressure learning operation, and the gear ratio becomes smaller. Is characterized in that the line pressure is further reduced by a predetermined amount, and the line pressure is gradually reduced by this line pressure learning operation until the change of the transmission gear ratio accompanying the running of the vehicle does not occur. A control device for a continuously variable transmission.

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