JP2015081642A - Control device of continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To accurately estimate a magnitude of torque Tin inputted to a belt-type CVT more than before, to maximally reduce belt grip pressure which is controlled on the basis of the estimation, and to further reduce a power loss.SOLUTION: In a hydraulic control circuit for controlling belt grip pressure, when an oil pump is interposed in a power transmission passage to a pulley at a drive side from an engine, a variation amount of a drive load of the oil pump for gear change motion is calculated when a gear change ratio of a CVT is changed by the gear change motion (step ST2), and the belt grip pressure is controlled by adding the variation amount thereto (step ST10).

Description

  The present invention relates to a control device for a continuously variable transmission using an endless transmission member such as a belt or a chain, and more particularly to control of a clamping pressure of an endless transmission member by a pulley.

  2. Description of the Related Art Conventionally, in a vehicle such as an automobile, a continuously variable transmission (CVT) capable of continuously changing the rotation input from the engine as a transmission that transmits the output of an engine that is a power source to driving wheels. Are known. Further, as a CVT, a belt type in which a transmission belt is wound between pulleys on the driving side and the driven side is generally put into practical use.

  In such a belt-type CVT, the transmission belt must be clamped with a large force in the V groove on the outer periphery of each pulley to suppress the slippage, but the friction between the transmission belt and the transmission belt increases as the clamping pressure increases. The loss increases, and high oil pressure is also required to maintain a large clamping pressure, so the engine fuel efficiency tends to deteriorate.

  In this regard, for example, the CVT described in Patent Document 1 indirectly detects the torque output from the engine, subtracts the hydraulic pump drive loss from this engine torque, and calculates the input torque to the CVT. The belt clamping pressure (the line pressure of the hydraulic control circuit of the CVT in this document) is controlled according to the input torque. If it carries out like this, it will suppress that a pinching pressure becomes large more than needed, and reduction of a fuel consumption will be aimed at.

JP-A-2-236052

  By the way, in the conventional technique, as a hydraulic pump drive loss, a control map is created by experimentally determining the hydraulic pump drive load corresponding to the engine speed and hydraulic pressure (the target secondary pressure in the same document). Keep it. A drive load (hereinafter also referred to as a pump drive load) calculated with reference to the control map based on the detected engine speed and hydraulic pressure while the vehicle is running is defined as a hydraulic pump drive loss.

  However, the pump drive load set in the control map as described above corresponds to a steady state in which the CVT gear ratio is kept substantially constant, and during the speed change operation in which the CVT gear ratio changes. There is no consideration in the prior art regarding transient fluctuations in the pump drive load. That is, in general, during the speed change operation, the flow speed of the hydraulic oil flowing through the oil passage of the hydraulic control circuit increases, so that the flow resistance increases and the pump driving load increases.

  If the pump driving load is increased, the input torque to the CVT is reduced accordingly, so that the belt clamping pressure can be reduced. The fluctuation of the pump drive load during the shifting operation is not taken into consideration. Therefore, in the conventional technique, there is still room for further reducing the clamping pressure and reducing the power loss in the CVT.

  Accordingly, an object of the present invention is to reduce the pinching pressure of the endless transmission member as much as possible and to further reduce the loss by estimating the magnitude of the torque input to the CVT more accurately than in the past.

  In order to achieve the above-mentioned object, the present invention has an endless transmission member wound between pulleys on the driving side and the driven side, and the endless transmission member is sandwiched according to the magnitude of torque input to the pulley on the driving side. A control device for a continuously variable transmission (hereinafter referred to as CVT) configured to control pressure is assumed.

  In the hydraulic control circuit for controlling the clamping force of the endless transmission member, when a hydraulic pump is interposed in the power transmission path from the driving power source of the vehicle to the driving pulley, In a shift operation in which the transmission ratio of the CVT changes, a fluctuation amount of the driving load of the hydraulic pump for the shift operation is calculated, and the clamping pressure of the endless transmission member is controlled in consideration of the fluctuation amount. It is a thing.

  In the CVT as described above, first, the magnitude of the torque input from the power source for driving the vehicle to the pulley on the driving side is obtained by, for example, subtracting the driving loss of the hydraulic pump from the torque output from the power source. It is determined in consideration of the inertia of the transmission member. Of course, if there is a torque converter in the power transmission path, torque amplification by this is also considered. Then, the clamping pressure is controlled so that the endless transmission member does not slip with respect to the magnitude of the input torque.

  In addition, during the CVT shift operation, the clamping pressure is controlled in consideration of the increase in the driving load of the hydraulic pump. In other words, during the CVT shift operation, the flow rate of the hydraulic fluid flowing through the oil passage of the hydraulic control circuit increases, so that the input load to the CVT decreases as the drive load (ie, drive loss) of the hydraulic pump increases. Therefore, the clamping pressure of the endless transmission member can be further reduced.

  In other words, according to the specific matter described above, the clamping force of the endless transmission member can be further reduced by reflecting the fluctuation of the pump drive load due to the CVT shift operation, thereby reducing the power loss in the CVT. Can be reduced.

  By the way, in order to reduce the clamping pressure as much as possible, it is necessary to accurately determine how much the driving load of the hydraulic pump is increased during the shift operation of the CVT. If the drive load increment is overestimated, the pinching pressure will be excessively reduced accordingly, which may cause slippage of the endless transmission member. Therefore, the loss reduction effect becomes low.

  In this respect, it is preferable that the drive load of the hydraulic pump is expressed as a function having the oil pressure and the rotation speed as variables (hereinafter referred to as a drive load function), and the amount of change is expressed as a change in the drive load due to a change in the oil pressure. In other words, it is calculated by adding up the fluctuation of the driving load due to the change in the rotational speed.

  Then, the fluctuation of the driving load due to the change in the oil pressure can be calculated as the product of the partial differential coefficient value due to the oil pressure of the driving load function and the amount of change in the oil pressure. Similarly, the fluctuation of the drive load due to the change in the rotational speed can be calculated as the product of the partial differential coefficient value due to the rotational speed of the drive load function and the amount of change in the rotational speed.

  In this case, if the partial differential coefficient value based on the hydraulic pressure or the rotational speed of the drive load function is obtained by numerical differentiation, it is not necessary to derive the drive load function from the mathematical model of the hydraulic control circuit. For example, in the same manner as the control map for determining the pump drive loss in the above-described prior art, a map in which the drive load of the hydraulic pump is set in correspondence with the operation state (hydraulic pressure and rotation speed) is created. What is necessary is just to approximate the difference value of the driving load at a certain operating point on the map as a partial differential coefficient value.

  As described above, according to the control device for a continuously variable transmission (CVT) according to the present invention, the clamping pressure of an endless transmission member such as a belt is controlled in accordance with the magnitude of input torque from the power source of the vehicle. In addition, it is possible to estimate the input torque more accurately than before by taking into account the drive load fluctuation of the hydraulic pump for shifting operation, thereby reducing the pinching pressure further and reducing the CVT. Power loss can be reduced.

It is a schematic block diagram which shows an example of the power train of the vehicle to which this invention is applied. It is a circuit block diagram of the gear ratio control part of a hydraulic control circuit, and a clamping pressure control part. It is a figure which shows an example of the transmission ratio control map of CVT. It is a figure which shows an example of the map of the pump drive loss which set the drive load (drive loss) of the oil pump corresponding to the line pressure and the engine speed. It is explanatory drawing which shows the force concerning the hydraulic oil which flows through a pipe line, (a) shows the case of a steady flow, (b) shows the case of an unsteady flow, respectively. It is a flowchart which shows an example of the flow of control of the belt clamping pressure during gear shifting operation. It is an image figure which shows an example of the map which computed and set the partial differential coefficient value by the line pressure of a drive load function previously. FIG. 8 is a diagram corresponding to FIG. 7 of a map in which partial differential coefficient values according to engine speed are set. It is a data figure showing distribution of the operating state of the engine in the case of a run test.

  Hereinafter, embodiments of the present invention will be described with reference to the drawings. As an example, in the present embodiment, a case will be described in which the present invention is applied to a power train mounted horizontally on a vehicle as schematically shown in FIG. Note that the description of the present embodiment is merely an example, and the configuration and application of the present invention are not limited.

(Schematic configuration of powertrain)
As schematically shown in FIG. 1, the power train of this embodiment includes an engine 1, which is a power source for traveling, a torque converter 2, a forward / reverse switching mechanism 3, a continuously variable transmission mechanism 4, a reduction gear mechanism 5, A differential gear mechanism 6 and the like are provided. The crankshaft 11 of the engine 1 is connected to the torque converter 2, and the output is transmitted from the torque converter 2 to the differential gear mechanism 6 via the forward / reverse switching mechanism 3, the continuously variable transmission mechanism 4 and the reduction gear mechanism 5. The left and right drive wheels 7 are distributed.

  The engine 1 is a multi-cylinder gasoline engine as an example, and includes an engine speed sensor 101 for calculating the engine speed ne. Further, in the illustrated example, the throttle valve 12 that adjusts the intake air amount of the engine 1 is operated by the throttle motor 13, and the opening degree (throttle opening degree Th) is such that the target intake air amount can be obtained. It is controlled by an ECU (Electronic Control Unit) 8. The throttle opening degree Th is detected by a throttle opening degree sensor 102.

-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 a one-way clutch 24, and is provided between the pump impeller 21 and the turbine runner 22. Power is transmitted by hydraulic oil (ATF). The pump impeller 21 is connected to the crankshaft 11 of the engine 1, while the turbine runner 22 is connected to the forward / reverse switching mechanism 3 via the turbine shaft 25.

  The torque converter 2 also includes a lockup clutch 26 that directly connects the input side and the output side thereof. The lock-up clutch 26 controls the differential pressure (lock-up differential pressure) between the hydraulic pressure in the engagement-side oil chamber and the hydraulic pressure in the release-side oil chamber, thereby enabling full engagement and half-engagement (engagement in the slip state). ) Or released state.

  In the released state of the lock-up clutch 26, power is transmitted from the pump impeller 21 to the turbine runner 22 by the hydraulic oil (ATF) as described above, but the rotational speed of the turbine runner 22 (turbine rotational speed Nt) is pumped. In a state lower than the rotational speed of the impeller 21 (same as the engine rotational speed ne), the output torque to the turbine shaft 25 is amplified according to the rotational difference.

  In the illustrated example, the torque converter 2 is provided with a mechanical oil pump 9 (hydraulic pump) that is connected to and driven by a pump impeller 21. The oil pump 9 is, for example, a gear pump or a vane pump, and is driven by the crankshaft 11 of the engine 1 via the pump impeller 21 (therefore, the rotational speed of the oil pump 9 becomes the engine rotational speed ne), which will be described later. The hydraulic oil is also supplied to the hydraulic control circuit 20.

-Forward / reverse switching mechanism-
The forward / reverse switching mechanism 3 includes a double pinion planetary gear mechanism 30, a forward clutch C1, and a reverse brake B1. The sun gear 31 of the planetary gear mechanism 30 is connected to the turbine shaft 25 of the torque converter 2, and a turbine rotation speed sensor 104 that detects the rotation speed of the turbine shaft 25 is disposed in the vicinity of the forward clutch C1. On the other hand, the carrier 33 of the planetary gear mechanism 30 is connected to the input shaft 40 of the continuously variable transmission mechanism 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. That is, when the forward clutch C1 is engaged and the reverse brake B1 is released, the forward / reverse switching mechanism 3 rotates integrally to establish a forward power transmission path. The driving force in the direction is transmitted to the continuously variable transmission mechanism 4 side.

  On the other hand, when the reverse brake B1 is engaged and the forward clutch C1 is released, the reverse drive mechanism is established by the forward / reverse switching mechanism 3. In this state, the input shaft 40 rotates in the reverse direction with respect to the turbine shaft 25, and the driving force in the reverse direction is transmitted to the continuously variable transmission mechanism 4 side. Further, when both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching mechanism 3 enters a neutral state in which power transmission is interrupted.

-Continuously variable transmission mechanism-
The continuously variable transmission mechanism 4 of the present embodiment is a belt type continuously variable transmission (stepless transmission) that can output the rotation input from the engine 1 via the torque converter 2 and the forward / reverse switching mechanism 3 in a stepless manner. CVT: Continuously Variable Transmission), hereinafter referred to as CVT4. The CVT 4 includes an input side (drive side) primary pulley 41, an output side (driven side) secondary pulley 42, and a metal transmission belt 43 wound between the primary pulley 41 and the secondary pulley 42. (Endless transmission member: including a chain belt).

  A primary pulley rotation speed sensor 105 is disposed in the vicinity of the primary pulley 41. From the output signal of the primary pulley rotation speed sensor 105, the input shaft rotation speed Nin of the CVT 4 can be calculated. Further, a secondary pulley rotational speed sensor 106 is disposed in the vicinity of the secondary pulley 42, and the output shaft rotational speed Nout of the CVT 4 can be calculated from the output signal. The vehicle speed spd can be calculated from the output shaft speed Nout.

  Specifically, the primary pulley 41 includes a fixed sheave 411 that is fixed to the input shaft 40 and a movable sheave 412 that is disposed on the input shaft 40 so as to be slidable only in the axial direction. Then, the winding diameter (effective diameter) of the transmission belt 43 is changed by changing the V groove width between the fixed sheave 411 and the movable sheave 412 by the hydraulic actuator 413 disposed on the movable sheave 412 side. It has become so.

  Similarly, the secondary pulley 42 also includes a fixed sheave 421 that is fixed to the output shaft 44 and a movable sheave 422 that is slidably disposed on the output shaft 44 in the axial direction, and is disposed on the movable sheave 422 side. By changing the width of the V-groove between the fixed sheave 421 and the movable sheave 422 by the hydraulic actuator 423, the winding diameter (effective diameter) of the transmission belt 43 is changed.

  Then, by controlling the hydraulic actuator 413 of the primary pulley 41 and changing the V groove width of each of the primary pulley 41 and the secondary pulley 42, the effective diameters of both the pulleys 41 and 42 are continuously changed, The speed ratio γ can be continuously changed. The gear ratio γ is defined as γ = input shaft rotational speed Nin / output shaft rotational speed Nout. For example, when the effective diameter of the primary pulley 41 increases and the effective form of the secondary pulley 42 decreases (upshift). The gear ratio γ becomes small.

  Thus, when the hydraulic actuator 413 of the primary pulley 41 is controlled to change the effective diameters of the primary pulley 41 and the secondary pulley 42, the hydraulic actuator 423 of the secondary pulley 42 does not slip the transmission belt 43. Thus, a predetermined clamping pressure is generated. That is, as will be described in detail later, the belt clamping pressure is determined according to the input torque Tin to the CVT 4, and the hydraulic actuator 423 is controlled to generate this belt clamping pressure.

(Hydraulic control circuit)
Next, the hydraulic control circuit 20 that controls the torque converter 2, the forward / reverse switching mechanism 3, the CVT 4 and the like will be described. The hydraulic control circuit 20 of the present embodiment includes a transmission ratio control unit 20a that mainly controls the hydraulic pressure of the hydraulic actuator 413 of the primary pulley 41 and a secondary pulley mainly when changing the transmission ratio γ of the CVT 4 as described above. And a clamping pressure control unit 20b that controls the hydraulic pressure of the hydraulic actuators 423. Although not shown in FIG. 1, the hydraulic control circuit 20 controls the hydraulic pressure for controlling the line pressure and engaging and releasing the lockup clutch 26, and the forward clutch C1 and the reverse drive of the forward / reverse switching mechanism 3. The hydraulic control for engaging and releasing the brake B1 is also performed.

  Specifically, as shown in FIG. 2, the hydraulic control circuit 20 includes a primary regulator valve 201, a select valve 202, a line pressure modulator valve 203, a solenoid modulator valve 204, a linear solenoid valve (SLP) 205, a linear solenoid valve (SLS) 206, A shift control valve 207 and a belt clamping pressure control valve 208 are provided.

  The hydraulic control circuit 20 is provided with an electric oil pump 90 using the electric motor 91 as a power source in addition to the mechanical oil pump 9 driven by the driving force from the engine 1 as described above. The electric oil pump 90 is disposed in parallel with the mechanical oil pump 9 and operates, for example, when the engine 1 is idle stopped.

  In this hydraulic pressure control circuit 20, first, the hydraulic pressure generated by the mechanical oil pump 9 (electric oil pump 90 at the time of idling stop) is regulated by, for example, a relief type primary regulator valve 201 to become the line pressure PL. . As an example, the primary regulator valve 201 is supplied with a control hydraulic pressure output from a linear solenoid valve (SLS) 206 via a select valve 202, and operates with the control hydraulic pressure as a pilot pressure.

  Thus, the line pressure PL regulated by the primary regulator valve 201 is supplied to the line pressure modulator valve 203 to be regulated to the lower modulated line pressure LPM2 in addition to the shift control valve 207 and the belt clamping pressure. The control valve 208 is supplied as it is as the line pressure PL. The shift control valve 207 and the belt clamping pressure control valve 208 will be described later.

  On the other hand, the modulated line pressure LPM2 regulated by the line pressure modulator valve 203 is supplied to a solenoid modulator valve 204, a linear solenoid valve (SLP) 205, and a linear solenoid valve (SLS) 206 as shown in FIG. The solenoid modulator valve 204 regulates the modulated line pressure LPM2 to a lower modulator hydraulic pressure PSM and supplies it to the shift control valve 207, the belt clamping pressure control valve 208, and the like.

  In addition, the linear solenoid valve (SLP) 205 and the linear solenoid valve (SLS) 206 are normally open type electromagnetic solenoid valves (may be normally closed type) in the illustrated example, and are respectively transmitted from the ECU 8. The control oil pressure is output according to the duty ratio of the control signal to output the control oil pressure using the modulated line pressure LPM2 as a source pressure.

  The control hydraulic pressure output in this way is supplied to the transmission control valve 207 as described below, and is supplied to the transmission ratio control of the CVT 4 and is supplied to the belt clamping pressure control valve 208 to be used for clamping pressure control. Is done. The control hydraulic pressure from the linear solenoid valve (SLS) 206 is also supplied to the primary regulator valve 201 as described above.

-Gear ratio control unit-
Next, the gear ratio control unit 20a that controls the hydraulic pressure of the hydraulic actuator 413 of the primary pulley 41 will be described in detail. As shown in FIG. 2, a shift control valve 207 for controlling the hydraulic pressure is connected to the hydraulic actuator 413 of the primary pulley 41 of the CVT 4 (hereinafter also referred to as the primary hydraulic actuator 413).

  The speed change control valve 207 includes a cylindrical spool valve as an example, and includes a spool 271 that is movable in the axial direction. A compression coil spring 272 is in a compressed state at one end (the lower end in FIG. 2). A control hydraulic pressure port 273 is provided. A control hydraulic pressure from a linear solenoid valve (SLP) 205 is applied to the control hydraulic pressure port 273.

  Further, the transmission control valve 207 is provided with an input port 274 to which the line pressure PL is supplied and an output port 275 connected to the primary side hydraulic actuator 413. The shift control valve 207 regulates the line pressure PL using the control hydraulic pressure output from the linear solenoid valve (SLP) 205 as a pilot pressure, and supplies the pressure to the primary hydraulic actuator 413 from the output port 275.

  That is, the output hydraulic pressure of the shift control valve 207 adjusted according to the control hydraulic pressure from the linear solenoid valve (SLP) 205 is supplied to the primary hydraulic actuator 413. As a result, the oil pressure Pin of the primary hydraulic actuator 413 (hereinafter also referred to as primary sheave oil pressure Pin) is regulated, and the gear ratio γ of the CVT 4 is controlled.

  Specifically, when the control hydraulic pressure from the linear solenoid valve (SLP) 205 increases while a predetermined hydraulic pressure is being supplied to the primary hydraulic actuator 413, the spool 271 of the shift control valve 207 is moved upward in FIG. Displacement increases the output hydraulic pressure, and the primary sheave hydraulic pressure Pin also increases. As a result, the V groove width of the primary pulley 41 becomes narrower, and the speed ratio γ becomes smaller (upshift).

  On the other hand, if the control hydraulic pressure from the linear solenoid valve (SLP) 205 decreases, the spool 271 of the shift control valve 207 is displaced downward in FIG. 2 and the output hydraulic pressure decreases, and the primary sheave hydraulic pressure Pin also decreases. As a result, the V-groove width of the primary pulley 41 is increased, and the gear ratio γ is increased (downshift).

  The control of the gear ratio control unit 20a as described above is performed by the ECU 8. That is, for example, according to a control map (see FIG. 3) to be described later, the ECU 8 calculates the target gear ratio of the CVT 4 (in this example, the target input speed Nint), and according to the deviation between the target gear ratio and the actual gear ratio γ. A control signal is generated. In response to this control signal, the control hydraulic pressure from the linear solenoid valve (SLP) 205 is regulated as described above, and the primary sheave hydraulic pressure Pin of the CVT 4 is controlled.

-Pinch pressure control part-
Next, the clamping pressure control unit 20b will be described in detail in the same manner as the gear ratio control unit 20a. The hydraulic actuator 423 of the secondary pulley 42 (hereinafter also referred to as the secondary hydraulic actuator 423) controls its hydraulic pressure. A belt clamping pressure control valve 208 is connected to the belt.

  The belt clamping pressure control valve 208 is also formed of a spool valve similar to the speed change control valve 207, and includes a spool 281 that is movable in the axial direction. A compression coil spring 282 is provided on one end side (lower end side in FIG. 2). Are arranged in a compressed state, and a control hydraulic pressure port 283 is provided. The control hydraulic pressure from the linear solenoid valve (SLS) 206 is applied to the control hydraulic pressure port 283.

  Further, the belt clamping pressure control valve 208 is formed with an input port 284 to which the line pressure PL is supplied and an output port 285 connected to the hydraulic actuator 423 of the secondary pulley 42. The belt clamping pressure control valve 208 regulates the line pressure PL using the control hydraulic pressure output from the linear solenoid valve (SLS) 206 as a pilot pressure, and supplies the pressure to the hydraulic actuator 423 of the secondary pulley 42 from the output port 285.

  In other words, the output hydraulic pressure of the belt clamping pressure control valve 208 adjusted in accordance with the control hydraulic pressure from the linear solenoid valve (SLS) 206 is supplied to the secondary hydraulic actuator 423, thereby the hydraulic pressure of the secondary hydraulic actuator 423. Pd (hereinafter also referred to as secondary sheave oil pressure Pd) is regulated to control the belt clamping pressure of CVT 4.

  Specifically, when the control hydraulic pressure from the linear solenoid valve (SLS) 206 increases while a predetermined hydraulic pressure is being supplied to the secondary hydraulic actuator 423, the spool 281 of the belt clamping pressure control valve 208 is shown in FIG. Displacement to the upper side increases the output hydraulic pressure Pd, and the secondary sheave hydraulic pressure Pd also increases. This also increases the belt clamping pressure. On the other hand, if the control hydraulic pressure from the linear solenoid valve (SLS) 206 decreases, the spool 281 of the belt clamping pressure control valve 208 is displaced downward in FIG. 2, and the belt sheave pressure Pd decreases as a result of the secondary sheave hydraulic pressure Pd decreasing. Reduced.

  The control of the clamping pressure control unit 20b as described above is performed by the ECU 8. That is, as an example, the ECU 8 calculates the target value (belt clamping pressure command value Pdi) of the secondary sheave hydraulic pressure Pd for obtaining a required belt clamping pressure according to an arithmetic expression described later, and linearly outputs the target hydraulic pressure. A solenoid valve (SLS) 206 is controlled. Thereby, the hydraulic pressure Pd of the secondary side hydraulic actuator 423 can be suitably controlled as described above.

  When the primary sheave oil pressure Pin and the secondary sheave oil pressure Pd are controlled independently as described above, the thrust ratio τ (τ = [secondary sheave oil pressure Pd × pressure receiving area Aout of the secondary side hydraulic actuator 423] / [primary sheave]. The primary sheave oil pressure Pin and the secondary sheave oil pressure Pd are controlled so that the hydraulic pressure Pin × the pressure receiving area Ain of the primary hydraulic actuator 413 can be maintained.

  The above-described control by the transmission ratio control unit 20a and the clamping pressure control unit 20b of the hydraulic control circuit 20 receives the control signals from the ECU 8, and the linear solenoid valves 207 and 208 of the control units 20a and 20b operate as described above. This is realized by suitably controlling the hydraulic pressure of the hydraulic actuators 413 and 423 of the CVT 4. That is, in the present embodiment, the control device for the continuously variable transmission according to the present invention is configured by a predetermined program executed by the ECU 8 as described later and the clamping pressure control unit 20b of the hydraulic control circuit 20.

(ECU)
Although not shown in the figure, the ECU 8 is a known unit including a CPU (Central Processing Unit), a ROM (Read Only Memory), a RAM (Random Access Memory), a backup RAM, and the like. The CPU executes various arithmetic processes based on various control programs and maps stored in the ROM. In addition, the RAM temporarily stores calculation results from the CPU, data input from each sensor, and the like, and the backup RAM stores data to be saved when the engine 1 is stopped, for example.

  1 is connected to the input interface of the ECU 8 such as the engine speed sensor 101, the throttle opening sensor 102, the turbine speed sensor 104, the primary pulley speed sensor 105, the secondary pulley speed sensor 106, and the like. ing. On the other hand, in addition to the throttle motor 13 and hydraulic control circuit 20 shown in FIG. 1, the output interface is connected to a fuel injection device, an ignition device, and the like of the engine 1. Based on the output signal and the like, control of the engine 1, control of the torque converter 2, control of the forward / reverse switching mechanism 3, control of the CVT 4 and the like are executed.

For example, as operation control of the engine 1, control signals are output to the throttle motor 13, the fuel injection device, the ignition device, and the like, and the intake air amount, fuel injection amount, ignition timing, and the like are controlled. Engagement and disengagement of the lockup clutch 26 is controlled for the torque converter 2, and engagement and disengagement of the forward clutch C1 and the reverse brake B1 are controlled for the forward / reverse switching mechanism 3. Then, the ECU 8 is controlled as CVT4. As an example, the target speed Nint is calculated with reference to the speed ratio control map shown in FIG. 3, and the speed ratio γ is controlled so that the actual input shaft speed Nin becomes the target speed Nint. The ECU 8 outputs a control signal for the target rotational speed Nint to the transmission ratio control unit 20a of the hydraulic control circuit 20, and controls the hydraulic pressure of the primary hydraulic actuator 413 as described above to continuously set the transmission ratio γ of the CVT 4. Change it.

  The gear ratio control map of FIG. 3 previously sets a gear ratio γ adapted by experiments and simulations, etc., using the throttle opening degree Th (accelerator opening degree) and the vehicle speed spd corresponding to the driver's output request as parameters. Which is stored in the ROM of the ECU 8. Since the vehicle speed spd corresponds to the output shaft rotational speed Nout, the target rotational speed Nint that is the target value of the input shaft rotational speed Nin is set as the target speed ratio γ in the control map.

(Control of belt clamping pressure)
As described below, the ECU 8 calculates the belt clamping pressure instruction value Pdi for suppressing the slippage of the transmission belt 43 according to the input torque Tin to the CVT 4 as described below. Is output to the clamping pressure control unit 20b of the hydraulic control circuit 20. Specifically, the ECU 8 adds a predetermined safety factor SF to the input torque Tin, and then calculates the belt clamping pressure instruction value Pdi by the following (Equation 1).

Pdi = (Tin + SF) × cos α / (2 × μ × Rin × Aout) (Formula 1)
In (Expression 1), Rin is the effective diameter of the primary pulley 41 (the winding radius of the transmission belt 43), and is a function of the speed ratio γ. Α is the sheave angle of the primary pulley 41 and the secondary pulley 42, and μ is a coefficient of friction between the pulleys 41, 42 and the transmission belt 43. Aout is the pressure receiving area of the secondary hydraulic actuator 423 as described above.

  Here, the magnitude of the input torque Tin used for the above calculation is obtained by subtracting the driving loss of the auxiliary machine such as the alternator and the oil pump 9 from the torque generated by the engine 1 (hereinafter referred to as engine torque Te). It is estimated in consideration of inertia such as the forward / reverse switching mechanism 3 and torque amplification by the torque converter 2. That is, the engine torque Te is calculated from the engine control map based on the throttle opening degree Th and the engine speed ne.

  The engine control map is well known and is not shown in the drawings, but experiments and simulations are performed in advance using the throttle opening degree Th (accelerator opening degree) and the engine speed ne corresponding to the driver's output request as parameters. The target engine torque that is more suitable is set, and is stored in the ROM of the ECU 8.

  In addition, for the driving loss of the auxiliary machine of the engine 1 and the oil pump 9, a map is prepared by previously obtaining the magnitude of the driving load corresponding to the operating state of the engine 1, the torque converter 2 and the CVT 4 by experiments and simulations. Keep it. Then, based on the engine speed ne during traveling of the vehicle, the line pressure PL of the hydraulic control circuit 20, and the like, the drive loss is calculated with reference to the map.

  For example, the driving load of the oil pump 9 is generally determined by the representative value of the operating hydraulic pressure in the hydraulic control circuit 20, that is, the line pressure PL, and the pump speed (in this embodiment, the engine speed ne). As shown, the magnitude of the pump drive load corresponding to the line pressure PL and the engine speed ne is examined by experiment and simulation, and set as a pump drive loss map.

In the example shown in FIG. 4, the pump drive load T _loss gradually increases as the engine speed ne increases, and more rapidly changes as the line pressure PL increases. Since the driving load of the oil pump 9 varies depending on the oil temperature in addition to the line pressure PL and the engine speed ne, at least two pump driving loss maps are prepared after unwarmed and warmed up. .

(Fluctuations in pump drive load during gear shifting)
Incidentally, the pump drive loss (pump drive load) set in the map as shown in FIG. 4 corresponds to a steady state in which the speed ratio γ of the CVT 4 is substantially constant. During the speed change operation in which the speed ratio γ of the CVT 4 is changed, the flow rate of the hydraulic oil flowing through the oil passage of the hydraulic control circuit 20 is increased, so that the inertia resistance and the viscosity loss are increased. The driving load increases.

  For this reason, the pump driving loss increases during the speed change operation, and the input torque to the CVT 4 is reduced accordingly, so that the belt clamping pressure necessary for suppressing the slippage of the transmission belt 43 is reduced. In consideration of this, in the present embodiment, the input torque Tin is estimated more accurately in consideration of the fluctuation of the driving loss (driving load) of the oil pump 9 during the shift operation, and more preferably based on this. The belt clamping pressure instruction value Pdi is calculated.

In the following, first, the fluctuation of the pump drive load will be described. As schematically shown in FIG. 5, the force F applied to the hydraulic oil (incompressible viscous fluid) flowing through the pipeline is shown in FIG. In the case of the one-dimensional steady flow shown in (a), it can be divided into the dynamic pressure f ₁ (inertia resistance) that pushes the fluid and the wall resistance f ₂ (viscosity loss) against this, If the density is ρ and the cross-sectional area of the flow path is Af, it is expressed as (Equation 2) below.

F = (f ₁ −f ₂ ) × Af
= {(1/2) ρu ² -μ (∂u / ∂y)} × Af (Formula 2)
When the flow velocity u becomes high and becomes u ′ during the shift operation of the CVT 4, both the dynamic pressure f ₁ and the wall resistance f ₂ increase as shown in FIG. 5B (f ₁ → f ′ ₁ , f ₂ → f ′ ₂ ), the increment dF of the force applied to the hydraulic oil is expressed by the following (Equation 3). This (Equation 3) can be approximated to the following (Equation 4) if macrourin expansion is performed and higher-order terms are ignored.

dF = [(1/2) (ρu ′ ² −ρu ² ) +
μ {(∂u ′ / ∂y) − (∂u / ∂y)}] × A f (Equation 3)
dF ≒ {ρu (∂u / ∂x ) dx + μ (∂ 2 u / ∂y 2) dy} × Af ... ( Equation 4)
If the dF of (Equation 4) is calculated at various points in the hydraulic control circuit 20 and summed over the entire flow path of the hydraulic oil, the increment of the driving load of the oil pump 9 can be obtained. It is difficult to obtain all the flow velocities u and u ′, the pipe friction resistance μ, and the channel cross-sectional area Af in each part of the circuit 20, and it is not realistic to perform the above calculation during the shift operation of the CVT 4.

Here, as described above, the driving load of the oil pump 9 can be represented by the line pressure PL and the engine speed ne in a steady state. Therefore, the drive load of the oil pump 9 is expressed as a function T _loss (PL, ne) having the line pressure PL and the engine speed ne as variables (hereinafter referred to as drive load function), and the difference dT _loss is expressed as follows. calculate.

dT _loss = (∂T _loss / ∂PL) × dPL + (∂T _loss / ∂ne) × dne (Formula 5)
That is, in the present embodiment, the fluctuation amount of the pump driving load is divided into a fluctuation amount due to the change in the line pressure PL and a fluctuation amount due to the change in the engine speed ne, respectively, from the driving load function T _loss (PL, ne). Calculated and added together. The fluctuation due to the change in the line pressure PL is calculated as the product of the partial differential coefficient value (∂T _loss / ∂PL) due to the line pressure PL of the drive load function T _loss (PL, ne) and the change dPL in the line pressure. Similarly, the variation due to the change in the engine speed ne can be calculated as the product of the partial differential coefficient value (∂T _loss / ∂ne) due to the engine speed ne and the change dne in the engine speed.

The partial differential coefficient value of the drive load function T _loss (PL, ne) may be obtained by numerical differentiation. That is, in the pump drive loss map of FIG. 4, when the operating state of CVT 4 changes, the drive load function T _loss (PL, ne) at a certain operating point (a point determined by the line pressure PL and the engine speed ne) is calculated. The difference value can be regarded as a partial differential coefficient value. For example, when a sudden downshift is started at the operating point T1 (PL1, ne1) in FIG. 4, the line pressure PL is increased and the engine speed ne is increased, and the operating point T2 (PL2, ne2) is shifted to. Will be described.

In this case, the fluctuation of the pump drive load when the engine speed ne slightly changes from the operating point T1 to the operating point T2 is shown as the slope of the arrow a shown in the figure, and the operating point T1 (PL1, PL1, The partial differential coefficient value of the driving load function T _loss (PL, ne) in ne1) based on the engine speed ne is calculated as in the following (Equation 6). Note that Δne is a step size for numerical differentiation and is set in advance. T _map (PL, ne) is a value of the drive load function T _loss (PL, ne) calculated with reference to the drive load map.

(∂T _loss / ∂ne) = {T _loss (PL1, ne1 + Δne) −T _loss (PL1, ne1)} / (ne1 + Δne−ne1)
= {T _map (PL1, ne1 + Δne) −T _map (PL1, ne1)} / Δne (Equation 6) Similarly, the fluctuation of the pump driving load when the line pressure PL slightly changes from the operating point T1. Is calculated as a partial differential coefficient value by the line pressure PL of the driving load function T _loss (PL, ne) at the operating point T1 (PL1, ne1) as in the following (Expression 7).

(∂T _loss / ∂PL) = {T _loss (PL1 + ΔPL, ne1) −T _loss (PL1, ne1)} / (PL1 + ΔPL−PL1)
= {T _map (PL1 + ΔPL, ne1) −T _map (PL1, ne1)} / ΔPL (Expression 7)
In this way, the fluctuation amount of the pump driving loss during the shifting operation can be obtained by numerical calculation with reference to the steady state pump driving loss map as shown in FIG. As a result, the input torque Tin to the CVT 4 can be accurately estimated by reflecting the fluctuation of the transient pump drive loss during the shifting operation, and the belt clamping pressure can be suitably controlled as described below. become.

-Nipping pressure control during shifting operation-
Hereinafter, the belt clamping pressure control of the CVT 4 as described above will be specifically described with reference to FIGS. The routine shown in the flowchart of FIG. 6 is repeatedly executed in the ECU 8 at predetermined time intervals (for example, several tens of milliseconds).

  In step ST1 after the start in the flow of FIG. 6, first, necessary data such as the engine torque Te is read. As described above, the engine torque Te is calculated from the engine control map based on the throttle opening degree Th and the engine speed ne, and is stored in the RAM of the ECU 8. Further, the driving loss of the engine accessory and the oil pump 9 is calculated with reference to the map. These data are used for calculation of input torque Tin described later.

  Subsequently, in step ST2, a pump load fluctuation term included in the input torque Tin is calculated. That is, the fluctuation amount of the pump drive load caused by the change in the operating state of the CVT 4 is calculated using the above (formula 6) and (formula 7). Specifically, for example, when a downshift is started at the operating point T1 (PL1 = 1MPa, ne1 = 1500rpm) and the operation point T2 (PL2 = 1.25MPa, ne2 = 2000rpm) shifts, the pump load fluctuation term is In this way, it is calculated.

First, if the step size of the numerical differentiation is ΔPL = 0.001 and Δne = 1, then (（T _loss / ∂PL) = {T _map (1.001,1500) −T _map (1.00,1500) )} / 0.001 = 2.25. Since the change amount dPL of the line pressure at the operating points T1 to T2 is 1.25−1.00 = 0.25, the fluctuation of the pump driving load due to the change of the line pressure PL is 2.25 × 0.25 = 0.56.

Similarly, the fluctuation of the pump drive load due to the change in the engine speed ne is (∂T _loss / ∂ne) = {T _map (1.00,1501) −T _map (1.00,1500)} / 1 = 4.0 × 10 ⁻⁴ , and the engine speed change amount dne at the operating points T1 to T2 is 2000-1500 = 500, so that 4.0 × 10 ⁻⁴ × 500 = 0.20. Therefore, the pump load fluctuation term is calculated as (dT _loss = 0.56 + 0.20 = 0.76) from the above (Formula 5).

  However, in the flow of the present embodiment, the calculation as described above is not performed in step ST2, but the partial differential coefficient values at a plurality of operating points are calculated in advance with reference to the pump drive loss map as shown in FIG. The map is set in association with the line pressure PL and the engine speed ne. In step ST2, a pump load fluctuation term is calculated with reference to the map.

As an example, FIG. 7 shows a pump load generated by a change in the partial differential coefficient value (∂T _loss / ∂PL) calculated by the above (Equation 7) at a plurality of operating points in the map of FIG. It is an image figure of the map which set the gradient of fluctuation. The illustrated graphs are for the case where the line pressure PL is 0.5, 1, 2, 3, 4, 5 (MPa). Referring to this map, the partial differential coefficient value (∂T _loss / ∂PL) of the operating point T1 (PL1 = 1MPa, ne1 = 1500rpm) is calculated as indicated by the arrow in the figure, and the change in line pressure dPL is calculated by this. Is multiplied.

Similarly, FIG. 8 is an image diagram of a map in which the partial differential coefficient value (∂T _loss / ∂ne) calculated by the above (Formula 6), that is, the gradient of the pump load fluctuation generated by the change in the engine speed ne is set. Show. Referring to this map, the partial differential coefficient value (図 T _loss / ∂ne) of the operating point T1 (PL1 = 1MPa, ne1 = 1500rpm) is calculated as shown by the arrow in the figure, and the engine speed ne What is necessary is just to multiply the change amount dne.

  In step ST3 following step ST2, for example, it is determined whether or not the shift operation of the CVT 4 is an upshift with a control signal output to the gear ratio control unit 20a. If the determination is negative (NO), step ST7 described later is performed. On the other hand, if the determination is affirmative (YES), the process proceeds to step ST4 to determine whether or not to change the safety factor SF. For this determination, for example, whether or not to change the vehicle type in advance is set and stored in the ROM of the ECU 8.

  And if it is a vehicle type which changes safety factor SF, it will progress to step ST5, and after changing to the small safety factor SFdown set beforehand, it will progress to below-mentioned step ST9 and will calculate input torque Tin to CVT4. On the other hand, if the vehicle type does not change the safety factor SF, the process proceeds to step ST6, the value of the pump load fluctuation term is set to zero (0), and then the process proceeds to step ST9 described later.

  On the other hand, in step ST7, which is determined as a negative determination (NO) that the shift is not an upshift in step ST3, it is determined whether or not the shift operation of the CVT 4 is a downshift. If the determination is negative (NO), the gear shifting operation is not in progress, so the value of the pump load fluctuation term is set to zero (0) and the process proceeds to step ST9. If the determination is affirmative (YES), the process proceeds to step ST9. .

  In step ST9, the auxiliary motor drive loss and the pump drive loss in the steady state (calculated with reference to the map of FIG. 4) are subtracted from the engine torque Te read in step ST1, and the torque generated by the torque converter 2 is subtracted. In consideration of inertia such as amplification and forward / reverse switching mechanism 3, input torque Tin to CVT 4 is calculated including the pump load fluctuation term calculated in step ST2. During the speed change operation, the lockup clutch 26 of the torque converter 2 is engaged, and the torque ratio is 1. Further, the inertia is calculated from the change amount per time of the input rotational speed Nin to the CVT 4, that is, the acceleration / deceleration.

  In other words, the input torque Tin to the CVT 4 can be estimated more accurately than before, taking into account that the drive loss of the oil pump 9 fluctuates during the shifting operation of the CVT 4. In step ST10, using the estimated input torque Tin, a belt clamping pressure instruction value Pdi that can suitably suppress slippage of the transmission belt 43 is calculated by the above-described (Equation 1).

  Therefore, in the CVT 4 according to the present embodiment, the hydraulic oil flow rate is increased in the hydraulic control circuit 20 during the downshift, and the input torque Tin is reduced by the increase in the driving load of the oil pump 9. Reflecting it suitably, the belt clamping pressure can be made as small as possible. As a result, the power loss in the CVT 4 can be further reduced than before, and the fuel consumption can be reduced by reducing the engine torque Te accordingly.

  Further, since the driving load of the oil pump 9 is reduced during the upshift, the value of the pump load fluctuation term is set to zero (0) as in step ST6 of the flow of FIG. Alternatively, as in step ST5, the safety factor SF is changed to a smaller value. Since the safety factor SF is originally set to a large value so that the transmission belt 43 does not slip even during the shifting operation of the CVT 4, as described above, the driving load of the oil pump 9 is reduced by the calculation. Even if the value of the input torque Tin increases, the belt clamping pressure instruction value Pdi can be decreased.

  FIG. 9 is data representing the distribution of the operating state of the engine 1 defined by the engine torque Te and the engine speed ne during a running test (JC08 mode) performed on a vehicle equipped with the power train of this embodiment. is there. The average value of the plotted points is close to the fuel efficiency optimum line as shown by the solid line graph E, but the conventional example shown by the phantom line (the belt load pressure instruction value Pdi is not included in the pump load fluctuation term) and In comparison, although shown exaggeratedly in the figure, the engine torque Te is generally reduced.

-Other embodiments-
As mentioned above, although the example which applied this invention to the power train of the vehicle carrying a gasoline engine was shown in embodiment described above, this invention is not restricted to this, The vehicle carrying other engines, such as a diesel engine, is shown. It is also applicable to. 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.

  Further, the present invention is not limited to the belt-type CVT 4 as in the above-described embodiment, and can be applied to, for example, a toroidal-type continuously variable transmission. As long as the hydraulic pump is driven by this, it can be applied to a multi-stage automatic transmission.

  In the above embodiment, the load variation of the oil pump 9 is included in the calculation of the input torque Tin during the shift operation of both the upshift and the downshift of the CVT 4. However, the present invention is not limited to this. Instead, for example, the load fluctuation may be included only during the downshift.

  The present invention can be applied to, for example, a belt-type CVT mounted on a vehicle. Particularly, since the fuel consumption can be reduced by reducing the belt clamping pressure during the shifting operation as much as possible, it is high in passenger cars and the like. There is an effect.

1 Engine (Power source for driving)
4. Continuously variable transmission mechanism (CVT: continuously variable transmission)
41 Primary pulley (drive pulley)
42 Secondary pulley (pulley on the driven side)
43 Belt (endless transmission member)
8 ECU
9 Oil pump (hydraulic pump)
20 Hydraulic control circuit 20b Clamping pressure control unit

Claims (1)

An endless transmission member is wound between the driving side and driven side pulleys, and is configured to control the clamping force of the endless transmission member in accordance with the magnitude of torque input to the driving side pulley. In the control device for a continuously variable transmission,
The hydraulic control circuit for controlling the clamping force of the endless transmission member has a hydraulic pump interposed in a power transmission path from a power source for driving the vehicle to the pulley on the driving side,
When a speed change operation in which the transmission ratio of the continuously variable transmission is changed, a fluctuation amount of the driving load of the hydraulic pump for the speed change operation is calculated, and the clamping pressure of the endless transmission member is also calculated in consideration of the fluctuation amount. A control device for a continuously variable transmission, wherein the continuously variable transmission is configured to control.

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