JP6984505B2 - Control device for vehicle power transmission device - Google Patents

Description

The present invention relates to a control device for a vehicle power transmission device having a plurality of power transmission paths provided in parallel between a power source and a drive wheel.

To transmit the power from the input rotating member to the output rotating member, which is provided in parallel between the input rotating member to which the power of the power source is transmitted and the output rotating member that outputs the power to the drive wheel. The plurality of power transmission paths are the first power transmission path via a gear mechanism having a gear stage formed by the engagement of the first engagement device, and the first power transmission path. 2 A control device for a vehicle power transmission device, which is a second power transmission path via a stepless speed change mechanism in which a transmission element is wound between a primary pulley and a secondary pulley, which is formed by engaging the engagement device. Is well known. For example, the control device for a vehicle transmission described in Patent Document 1 is that. In this Patent Document 1, in consideration of the slip limit thrust according to the input torque to the stepless speed change mechanism, which is the thrust necessary to prevent the transmission element from slipping with the minimum necessary thrust, the primary pulley It is disclosed to calculate the thrust target value and the thrust target value of the secondary pulley, respectively.

Japanese Unexamined Patent Publication No. 2015-227697

By the way, in a vehicle provided with a power transmission device for a vehicle having a plurality of power transmission paths, the second engaging device may be in an engaged state or the second engaging device may be in an released state. Therefore, the input torque to the continuously variable transmission mechanism used to calculate the thrust of each pulley required to prevent the transmission element from slipping is represented by, for example, the engaged state or the disengaged state of the second engaging device. It is conceivable to calculate according to the operating state. For example, when the second engaging device is in the released state, it is conceivable that the input torque to the continuously variable transmission mechanism is converted into the drag torque of the second engaging device converted onto the input rotating member. On the other hand, if the first engaging device is switched from the released state to the engaged state while the second engaging device is in the released state, the input rotating member rotates during the excessive engagement of the first engaging device. The inertia torque accompanying the change is input to the stepless speed change mechanism. This inertia torque is increased as the angular acceleration of the input rotating member is larger and the gear ratio of the continuously variable transmission mechanism is the gear ratio on the high side. If an inertial shuttlek that is large enough to exceed the drag torque of the second engaging device described above is generated, the required thrust may be insufficient and the transmission element may slip.

The present invention has been made in the background of the above circumstances, and an object thereof is that a large inner shuttle torque is input to the continuously variable transmission mechanism when the second engaging device is in a completely released state. It is an object of the present invention to provide a control device for a power transmission device for a vehicle, which can prevent the occurrence of slippage of a transmission element under such circumstances.

The gist of the first invention is (a) the power provided in parallel between an input rotating member to which the power of the power source is transmitted and an output rotating member to output the power to the drive wheels. Each of the input rotating members has a plurality of power transmission paths that can be transmitted to the output rotating member, and the plurality of power transmission paths form a gear stage formed by engagement of the first engaging device. A first power transmission path via a gear mechanism having a gear mechanism, and a stepless speed change mechanism in which a transmission element is wound between a primary pulley and a secondary pulley formed by engagement of a second engagement device. (2) The primary of the power transmission device for a vehicle, which is a power transmission path, which is (b) necessary for preventing the transmission element from slipping based on the input torque to the stepless speed change mechanism. The thrust of the primary pulley for pinching the transmission element applied by the hydraulic actuator of the pulley and the thrust of the secondary pulley for pinching the transmission element applied by the hydraulic actuator of the secondary pulley are calculated and the second. When the engaging device is in a completely released state, it includes a shift control unit that uses the input torque to the stepless speed change mechanism as the drag torque of the second engaging device converted on the input rotating member (c). ) The shift control unit releases the first engaging device from the released state in accordance with the shift operation by the driver who switches the shift switching device to the traveling operation position when the second engaging device is in the completely released state. The engagement control of the first engaging device that switches to the engaged state is being executed, and the gear ratio of the stepless speed change mechanism is a predetermined high-side limit that enables slip prevention of the transmission element. When it is on the higher side than the gear ratio, the thrust of the primary pulley and the thrust of the secondary pulley, which are necessary for preventing the transmission element from slipping, are calculated using the drag torque of the second engaging device. The thrust is applied to the inner shuttlek that is input to the stepless speed change mechanism according to the rotation change of the input rotating member during the execution of the engagement control of the first engaging device, respectively, to prevent the transmission element from slipping. It is to increase the predetermined thrust that guarantees.

According to the first invention, when the second engaging device is in the completely released state, the engagement control of the first engaging device accompanying the shift operation is being executed, and the stepless speed change mechanism. When the gear ratio is higher than the predetermined high limit gear ratio, it is calculated using the drag torque of the second engaging device, which is necessary for preventing the transmission element from slipping. Since the thrust of the pulley and the thrust of the secondary pulley are each increased by a predetermined thrust that guarantees slip prevention of the transmission element with respect to the inner shuttle torque input to the stepless speed change mechanism, the gear ratio of the stepless speed change mechanism is increased. Since the engagement control of the first engaging device is being executed due to the shift operation in which the angular acceleration of the input rotating member is likely to be increased when the input rotating member is on the relatively high side, the drag torque of the second engaging device is exceeded. The torque capacity of the transmission element required for preventing the transmission element from slipping can be appropriately secured against the tendency of such a large inner shuttlek to occur. Therefore, when the second engaging device is in the completely released state, it is possible to prevent the transmission element from slipping under the condition that a large inertia torque is input to the continuously variable transmission mechanism.

It is a figure explaining the schematic structure of the vehicle to which this invention is applied, and also is the figure explaining the main part of the control function and the control system for various control in a vehicle. It is a figure for demonstrating the structure of a stepless speed change mechanism and the structure of a hydraulic control circuit. It is a figure which shows an example for demonstrating the thrust necessary for shift control. It is a figure which shows an example of the relationship of each thrust at the time of t2 of FIG. It is a block diagram which shows the control structure for achieving both the target shift and the belt slip prevention with the minimum necessary thrust. It is a figure which shows an example of the thrust ratio map for calculating the thrust ratio used for the calculation of the thrust on the secondary pulley side. It is a figure which shows an example of the thrust ratio map for calculating the thrust ratio used for the calculation of the thrust on the primary pulley side. It is a figure which shows an example of the differential thrust map for calculating the secondary differential thrust. It is a figure which shows an example of the differential thrust map for calculating the primary differential thrust. Explains the control operation to prevent the occurrence of belt slip in the situation where a large inertia torque is input to the continuously variable transmission mechanism when the main part of the control operation of the electronic control device, that is, the second clutch is in the completely released state. It is a flowchart to be performed.

In the embodiment of the present invention, the primary pulley which is the pulley on the input side and the secondary pulley which is the pulley on the output side are, for example, the groove width between the fixed sheave and the movable sheave and their fixed sheave and the movable sheave, respectively. It has the hydraulic actuator which applies a thrust for changing the above. The vehicle provided with the vehicle power transmission device includes a hydraulic control circuit that independently controls the pulley hydraulic pressure as the operating hydraulic pressure supplied to the hydraulic actuator. This hydraulic control circuit may be configured to result in pulley hydraulic pressure by, for example, controlling the flow rate of hydraulic oil to the hydraulic actuator. By controlling each thrust (= pulley hydraulic pressure × pressure receiving area) in the primary pulley and the secondary pulley by such a hydraulic pressure control circuit, the target shift can be realized while preventing the transmission element from slipping. The shift control is executed as follows. The transmission element wound between the primary pulley and the secondary pulley is an endless annular hoop and an element which is a thick plate piece-like block connected in a large number in the thickness direction along the hoop. An endless annular compression type transmission belt, or a tension type transmission belt in which the ends of alternately stacked link plates are connected to each other by connecting pins to form an endless annular link chain. The continuously variable transmission mechanism is a known belt-type continuously variable transmission. In a broad sense, the concept of this belt-type continuously variable transmission includes a chain-type continuously variable transmission.

The gear ratio is "rotational speed of the rotating member on the input side / rotational speed of the rotating member on the output side". For example, the gear ratio of the continuously variable transmission mechanism is "rotational speed of primary pulley / rotational speed of secondary pulley". The gear ratio of the vehicle power transmission device is "rotational speed of the input rotating member / rotational speed of the output rotating member". The high side in the gear ratio is the high vehicle speed side on which the gear ratio becomes smaller. The low side in the gear ratio is the low vehicle speed side on which the gear ratio becomes large. For example, the lowest gear ratio is the gear ratio on the lowest vehicle speed side, which is the lowest vehicle speed side, and is the maximum gear ratio at which the gear ratio is the largest.

Further, the power source is, for example, an engine such as a gasoline engine or a diesel engine that generates power by burning fuel. Further, the vehicle may be provided with an electric motor or the like as the power source in addition to or in place of the engine.

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings.

FIG. 1 is a diagram illustrating a schematic configuration of a vehicle 10 to which the present invention is applied, and is a diagram illustrating a main part of a control function and a control system for various controls in the vehicle 10. In FIG. 1, the vehicle 10 includes an engine 12 that functions as a power source, a drive wheel 14, and a vehicle power transmission device 16 provided in a power transmission path between the engine 12 and the drive wheel 14. .. Hereinafter, the power transmission device 16 for a vehicle is referred to as a power transmission device 16.

The power transmission device 16 is connected to a known torque converter 20 as a fluid transmission device connected to the engine 12, an input shaft 22 connected to the torque converter 20, and an input shaft 22 in the case 18 as a non-rotating member. A gear mechanism connected to the input shaft 22 via a stepless speed change mechanism 24, a forward / backward changeover device 26 also connected to the input shaft 22, and a gear mechanism provided in parallel with the stepless speed change mechanism 24. 28, deceleration consisting of a pair of gears that are provided on the output shaft 30, the counter shaft 32, the output shaft 30 and the counter shaft 32, which are common output rotating members of the stepless speed change mechanism 24 and the gear mechanism 28, and mesh with each other so as not to rotate relative to each other. The gear device 34, the gear 36 provided on the counter shaft 32 so as not to rotate relative to each other, the differential gear 38 connected to the gear 36, and the like are provided. Further, the power transmission device 16 includes left and right axles 40 connected to the differential gear 38. The input shaft 22 is an input rotating member to which the power of the engine 12 is transmitted. The output shaft 30 is an output rotating member that outputs the power of the engine 12 to the drive wheels 14. Unless otherwise specified, the above-mentioned power agrees with torque and force.

In the power transmission device 16 configured in this way, the power output from the engine 12 sequentially passes through the torque converter 20, the forward / backward switching device 26, the gear mechanism 28, the reduction gear device 34, the differential gear 38, the axle 40, and the like. , Is transmitted to the left and right drive wheels 14. Alternatively, in the power transmission device 16, the power output from the engine 12 is sequentially transmitted to the left and right drive wheels 14 via the torque converter 20, the continuously variable transmission mechanism 24, the reduction gear device 34, the differential gear 38, the axle 40, and the like. Will be done.

As described above, the power transmission device 16 includes a gear mechanism 28 and a continuously variable transmission mechanism 24 provided in parallel with the power transmission path PT between the engine 12 and the drive wheels 14. Specifically, the power transmission device 16 includes a gear mechanism 28 and a stepless speed change mechanism 24 provided in parallel with the power transmission path PT between the input shaft 22 and the output shaft 30. That is, the power transmission device 16 is provided in parallel between the input shaft 22 and the output shaft 30, and is capable of transmitting the power of the engine 12 from the input shaft 22 to the output shaft 30. It is equipped with. The plurality of power transmission paths are the first power transmission path PT1 via the gear mechanism 28 and the second power transmission path PT2 via the stepless speed change mechanism 24. That is, the power transmission device 16 includes a plurality of power transmission paths of the first power transmission path PT1 and the second power transmission path PT2 in parallel between the input shaft 22 and the output shaft 30. The first power transmission path PT1 is a power transmission path for transmitting the power of the engine 12 from the input shaft 22 to the drive wheels 14 via the gear mechanism 28. The second power transmission path PT2 is a power transmission path for transmitting the power of the engine 12 from the input shaft 22 to the drive wheels 14 via the continuously variable transmission mechanism 24.

In the power transmission device 16, the power transmission path for transmitting the power of the engine 12 to the drive wheels 14 is switched between the first power transmission path PT1 and the second power transmission path PT2 according to the traveling state of the vehicle 10. Therefore, the power transmission device 16 includes a plurality of engagement devices that selectively form the first power transmission path PT1 and the second power transmission path PT2. The plurality of engaging devices include a first clutch C1, a first brake B1, and a second clutch C2. The first clutch C1 is provided in the first power transmission path PT1 and is an engaging device that selectively connects or disconnects the first power transmission path PT1 and is engaged when moving forward. This is an engaging device that forms the first power transmission path PT1. The first brake B1 is provided in the first power transmission path PT1 and is an engaging device for selectively connecting or disconnecting the first power transmission path PT1 and is engaged when moving backward. This is an engaging device that forms the first power transmission path PT1. The first power transmission path PT1 is formed by engaging the first clutch C1 or the first brake B1. The second clutch C2 is provided in the second power transmission path PT2, and is an engaging device that selectively connects or disconnects the second power transmission path PT2. 2 It is an engaging device that forms the power transmission path PT2. The second power transmission path PT2 is formed by the engagement of the second clutch C2. The first clutch C1, the first brake B1, and the second clutch C2 are all known hydraulic wet friction engagement devices that are frictionally engaged by a hydraulic actuator. The first clutch C1 is a forward first engaging device, the second clutch C2 is a second engaging device, and the first brake B1 is a reverse first engaging device. Each of the first clutch C1 and the first brake B1 is one of the elements constituting the forward / backward switching device 26, as will be described later.

The engine 12 includes an engine control device 42 having various devices necessary for output control of the engine 12, such as an electronic throttle device, a fuel injection device, and an ignition device. The engine 12 is an engine by controlling the engine control device 42 according to the accelerator operation amount θacc, which is the operation amount of the accelerator pedal corresponding to the drive request amount for the vehicle 10 by the driver, by the electronic control device 90 described later. The engine torque Te, which is the output torque of 12, is controlled.

The torque converter 20 includes a pump impeller 20p connected to the engine 12 and a turbine impeller 20t connected to the input shaft 22. The power transmission device 16 includes a mechanical oil pump 44 connected to a pump impeller 20p. The oil pump 44 is rotationally driven by the engine 12 to control the speed change of the continuously variable transmission mechanism 24, generate a belt pinching pressure in the continuously variable transmission mechanism 24, and engage with each of the plurality of engaging devices. The main pressure of the hydraulic pressure for switching the operating state such as setting and releasing is supplied to the hydraulic control circuit 46 provided in the vehicle 10.

The forward / backward switching device 26 includes a double pinion type planetary gear device 26p, a first clutch C1, and a first brake B1. The planetary gear device 26p is a differential mechanism having three rotating elements: a carrier 26c as an input element, a sun gear 26s as an output element, and a ring gear 26r as a reaction force element. The carrier 26c is connected to the input shaft 22. The ring gear 26r is selectively coupled to the case 18 via the first brake B1. The sun gear 26s is connected to a small diameter gear 48 provided around the input shaft 22 so as to be rotatable relative to the input shaft 22 coaxially with the input shaft 22. The carrier 26c and the sun gear 26s are selectively connected via the first clutch C1.

The gear mechanism 28 is provided with a small diameter gear 48, a gear mechanism counter shaft 50, and a large diameter around the gear mechanism counter shaft 50 so as to be coaxially non-rotatable with respect to the gear mechanism counter shaft 50 and mesh with the small diameter gear 48. It is equipped with a gear 52. The large-diameter gear 52 has a larger diameter than the small-diameter gear 48. Further, the gear mechanism 28 is coaxial with the idler gear 54 provided around the gear mechanism counter shaft 50 so as to be rotatable relative to the gear mechanism counter shaft 50 and about the output shaft 30 with respect to the output shaft 30. It is provided with an output gear 56 that is provided in the center so as not to rotate relative to each other and meshes with the idler gear 54. The output gear 56 has a larger diameter than the idler gear 54. Therefore, in the gear mechanism 28, one gear stage is formed in the power transmission path PT between the input shaft 22 and the output shaft 30. The gear mechanism 28 is a gear mechanism having a gear stage. The gear mechanism 28 is further provided around the gear mechanism counter shaft 50 between the large-diameter gear 52 and the idler gear 54, and meshes with which the power transmission path between them is selectively connected or disconnected. It is equipped with a formula clutch D1. The meshing clutch D1 is an engaging device that selectively connects or disconnects the first power transmission path PT1 and forms the first power transmission path PT1 by being engaged. be. The meshing clutch D1 is an engaging device that forms a first power transmission path PT1 by being engaged with the first clutch C1 or the first brake B1, and is included in the plurality of engaging devices. The operating state of the meshing clutch D1 is switched by the operation of a hydraulic actuator (not shown) provided in the power transmission device 16.

The first power transmission path PT1 is formed by engaging both the meshing clutch D1 and the first clutch C1 or the first brake B1 provided on the input shaft 22 side of the meshing clutch D1. .. The engagement of the first clutch C1 forms a power transmission path for forward movement, while the engagement of the first brake B1 forms a power transmission path for reverse movement. In the power transmission device 16, when the first power transmission path PT1 is formed, the power of the engine 12 can be transmitted from the input shaft 22 to the output shaft 30 via the gear mechanism 28. .. On the other hand, the first power transmission path PT1 is put into a neutral state in which power transmission is impossible when both the first clutch C1 and the first brake B1 are released or when the meshing clutch D1 is released.

FIG. 2 is a diagram for explaining the configuration of the continuously variable transmission mechanism 24. In FIGS. 1 and 2, in the stepless speed change mechanism 24, the primary shaft 58 provided coaxially with the input shaft 22 and integrally connected to the input shaft 22, and the effective diameter connected to the primary shaft 58 are variable. Is wound between the primary pulley 60, the secondary shaft 62 provided coaxially with the output shaft 30, the secondary pulley 64 with a variable effective diameter connected to the secondary shaft 62, and the pulleys 60 and 64, respectively. It is provided with a transmission belt 66 as a transmission element. The continuously variable transmission mechanism 24 is a known belt-type continuously variable transmission in which power is transmitted via frictional force between the pulleys 60 and 64 and the transmission belt 66, and the power of the engine 12 is driven by the drive wheels 14. Communicate to the side. The frictional force agrees with the pinching pressure, and is also referred to as a belt pinching pressure. This belt pinching pressure is the belt torque capacity Tcvt, which is the torque capacity of the transmission belt 66 in the continuously variable transmission mechanism 24.

The primary pulley 60 includes a fixed sheave 60a connected to the primary shaft 58, a movable sheave 60b provided so as to be non-rotatable relative to the axis of the primary shaft 58 and movable in the axial direction with respect to the fixed sheave 60a. It is provided with a hydraulic actuator 60c that applies a primary thrust Win to the movable sheave 60b. The primary thrust Win is the thrust (= primary pressure Pin × pressure receiving area) of the primary pulley 60 for changing the V-groove width between the fixed sheave 60a and the movable sheave 60b. That is, the primary thrust Win is the thrust of the primary pulley 60 that clamps the transmission belt 66 applied by the hydraulic actuator 60c. The primary pressure Pin is the hydraulic pressure supplied to the hydraulic actuator 60c by the hydraulic control circuit 46, and is the pulley hydraulic pressure that produces the primary thrust Win. Further, the secondary pulley 64 has a fixed sheave 64a connected to the secondary shaft 62 and a movable sheave 64b provided so as to be non-rotatable and movable in the axial direction of the secondary shaft 62 with respect to the fixed sheave 64a. And a hydraulic actuator 64c that applies a secondary thrust Wout to the movable sheave 64b. The secondary thrust Wout is the thrust (= secondary pressure Pout × pressure receiving area) of the secondary pulley 64 for changing the V-groove width between the fixed sheave 64a and the movable sheave 64b. That is, the secondary thrust Wout is the thrust of the secondary pulley 64 that clamps the transmission belt 66 applied by the hydraulic actuator 64c. The secondary pressure Pout is the hydraulic pressure supplied to the hydraulic actuator 64c by the hydraulic control circuit 46, and is the pulley hydraulic pressure that causes the secondary thrust Wout.

In the continuously variable transmission mechanism 24, the primary pressure Pin and the secondary pressure Pout are respectively controlled by the hydraulic control circuit 46 driven by the electronic control device 90 described later, so that the primary thrust Win and the secondary thrust Wout are controlled respectively. To. As a result, in the continuously variable transmission mechanism 24, the V-groove width of each of the pulleys 60 and 64 is changed, the hooking diameter (= effective diameter) of the transmission belt 66 is changed, and the gear ratio γcvt (= primary rotation speed Npri / secondary rotation) is changed. The speed Nsec) is changed, and the belt pinching pressure is controlled so that the transmission belt 66 does not slip. That is, by controlling the primary thrust Win and the secondary thrust Wout respectively, the shift ratio γcvt of the continuously variable transmission mechanism 24 is set as the target shift ratio γcvttgt while preventing the belt slip, which is the slip of the transmission belt 66. The primary rotation speed Npri is the rotation speed of the primary shaft 58, and the secondary rotation speed Nsec is the rotation speed of the secondary shaft 62.

In the continuously variable transmission mechanism 24, when the primary pressure Pin is increased, the V-groove width of the primary pulley 60 is narrowed and the gear ratio γcvt is reduced. The reduction of the gear ratio γcvt means that the continuously variable transmission mechanism 24 is upshifted. In the continuously variable transmission mechanism 24, the highest gear ratio γmin is formed where the V-groove width of the primary pulley 60 is minimized. The highest gear ratio γmin is the gear ratio γcvt on the highest vehicle speed side, which is the highest vehicle speed side in the range of the gear ratio γcvt that can be formed by the continuously variable transmission mechanism 24, and the gear ratio γcvt is the smallest value. The minimum gear ratio. On the other hand, in the continuously variable transmission mechanism 24, when the primary pressure Pin is lowered, the V-groove width of the primary pulley 60 is widened and the gear ratio γcvt is increased. Increasing the gear ratio γcvt means that the continuously variable transmission mechanism 24 is downshifted. In the continuously variable transmission mechanism 24, the lowest gear ratio γmax is formed where the V-groove width of the primary pulley 60 is maximized. The lowest gear ratio γmax is the gear ratio γcvt on the lowest vehicle speed side, which is the lowest vehicle speed side in the range of the gear ratio γcvt that can be formed by the continuously variable transmission mechanism 24, and the gear ratio γcvt is the largest value. Maximum gear ratio. In the continuously variable transmission mechanism 24, the belt slip is prevented by the primary thrust Win and the secondary thrust Wout, and the target gear ratio γcvttgt is realized by the mutual relationship between the primary thrust Win and the secondary thrust Wout. The target shift cannot be achieved with only one thrust. As will be described later, in the mutual relationship between the primary pressure Pin and the secondary pressure Pout, the thrust ratio τ (= Wout / Win), which is the value of the ratio between the primary thrust Win and the secondary thrust Wout, is changed to cause stepless speed change. The gear ratio γcvt of the mechanism 24 is changed. The thrust ratio τ is the value of the ratio of the secondary thrust Wout to the primary thrust Win. For example, as the thrust ratio τ increases, the gear ratio γcvt increases, that is, the stepless speed change mechanism 24 is downshifted.

The output shaft 30 is arranged so as to be rotatable relative to the secondary shaft 62 in a coaxial center. The second clutch C2 is provided in the power transmission path between the secondary pulley 64 and the output shaft 30. The second power transmission path PT2 is formed by engaging the second clutch C2. In the power transmission device 16, when the second power transmission path PT2 is formed, the power of the engine 12 can be transmitted from the input shaft 22 to the output shaft 30 via the stepless speed change mechanism 24. Will be done. On the other hand, the second power transmission path PT2 is put into the neutral state when the second clutch C2 is released. The gear ratio γcvt of the continuously variable transmission mechanism 24 corresponds to the gear ratio in the second power transmission path PT2.

In the power transmission device 16, the gear ratio EL of the gear mechanism 28, which is the shift ratio γ gear (= input shaft rotation speed Nin / output shaft rotation speed Nout) in the first power transmission path PT1, is the maximum shift in the second power transmission path PT2. The value is set to be larger than the lowest speed change ratio γmax of the stepless speed change mechanism 24, which is a ratio. That is, the gear ratio EL is set to the gear ratio on the low side of the gear ratio γmax on the lowest side. The gear ratio EL of the gear mechanism 28 corresponds to the first speed gear ratio γ1 in the power transmission device 16, and the lowest gear ratio γmax of the continuously variable transmission mechanism 24 corresponds to the second speed gear ratio γ2 in the power transmission device 16. Equivalent to. In this way, the second power transmission path PT2 is formed with a gear ratio on the higher side than the first power transmission path PT1. The input shaft rotation speed Nin is the rotation speed of the input shaft 22, and the output shaft rotation speed Nout is the rotation speed of the output shaft 30.

In the vehicle 10, it is possible to selectively perform traveling in the gear traveling mode and traveling in the belt traveling mode. The gear traveling mode is a traveling mode in which traveling can be performed using the first power transmission path PT1, and is a traveling mode in which the first power transmission path PT1 is formed in the power transmission device 16. The belt traveling mode is a traveling mode in which traveling can be performed using the second power transmission path PT2, and is a traveling mode in which the second power transmission path PT2 is formed in the power transmission device 16. In the gear traveling mode, the first clutch C1 and the meshing clutch D1 are engaged and the second clutch C2 and the first brake B1 are released when the forward traveling is possible. In the gear traveling mode, the first brake B1 and the meshing clutch D1 are engaged and the second clutch C2 and the first clutch C1 are released when the reverse traveling is possible. In the belt traveling mode, the second clutch C2 is engaged and the first clutch C1 and the first brake B1 are released. In this belt running mode, forward running is possible.

The gear travel mode is selected in a relatively low vehicle speed region, including when the vehicle is stopped. The belt travel mode is selected in a relatively high vehicle speed region including a medium vehicle speed region. The meshing clutch D1 is engaged in the belt travel mode in the medium vehicle speed region of the belt travel mode, while the mesh clutch D1 is released in the belt travel mode in the high vehicle speed region of the belt travel mode. .. The meshing clutch D1 is released in the belt traveling mode in the high vehicle speed region, for example, by eliminating the drag of the gear mechanism 28 or the like during traveling in the belt traveling mode, and at high vehicle speed, the gear mechanism 28 or the planetary gear device. This is to prevent the 26p component, for example, a pinion, from rotating at a high speed.

The vehicle 10 includes an electronic control device 90 as a controller including a control device for the power transmission device 16. The electronic control device 90 includes, for example, a so-called microcomputer provided with a CPU, RAM, ROM, an input / output interface, etc., and the CPU uses a temporary storage function of the RAM and follows a program stored in the ROM in advance. Various controls of the vehicle 10 are executed by performing signal processing. The electronic control device 90 is an output control of the engine 12, a shift control of the continuously variable transmission mechanism 24, a belt pinching pressure control, and a hydraulic control for switching the operating state of each of the plurality of engaging devices (C1, B1, C2, D1). And so on. The electronic control device 90 is separately configured for engine control, hydraulic control, and the like, if necessary.

The electronic control device 90 includes various sensors provided in the vehicle 10 (for example, various rotational speed sensors 70, 72, 74, 76, accelerator operation amount sensor 78, throttle opening sensor 80, shift position sensor 82, etc.). Detection signals, etc. (for example, engine rotation speed Ne, primary rotation speed Npri equal to input shaft rotation speed Nin, secondary rotation speed Nsec, output shaft rotation speed Nout corresponding to vehicle speed V, magnitude of driver's acceleration operation The accelerator operation amount θacc, the throttle opening tap, the operation position POSsh of the shift lever 84 as the shift switching device provided in the vehicle 10, etc.) are supplied respectively. Further, from the electronic control device 90, various command signals (for example, engine control command signal Se for controlling the engine 12) are transmitted to each device (for example, engine control device 42, hydraulic control circuit 46, etc.) provided in the vehicle 10. The hydraulic control command signal Scvt for controlling the speed change of the speed change mechanism 24, the belt pinching pressure, etc., the hydraulic control command signal Scbd for controlling the operating state of each of the plurality of engaging devices, etc.) are output, respectively. Ru. The input shaft rotation speed Nin (= primary rotation speed Npri) is also the turbine rotation speed, the primary rotation speed Npri is also the rotation speed of the primary pulley 60, and the secondary rotation speed Nsec is the rotation speed of the secondary pulley 64. But it is also. Further, the electronic control device 90 calculates the actual gear ratio γcvt (= Npri / Nsec), which is the actual gear ratio γcvt of the continuously variable transmission mechanism 24, based on the primary rotation speed Npri and the secondary rotation speed Nsec.

The operation position POSsh of the shift lever 84 is, for example, a P, R, N, D operation position. The P operation position is a parking operation position for selecting the P position of the power transmission device 16 in which the power transmission device 16 is in the neutral state and the output shaft 30 is mechanically fixed so as not to rotate. The neutral state of the power transmission device 16 is realized, for example, by releasing the first clutch C1, the first brake B1, and the second clutch C2. That is, the neutral state of the power transmission device 16 is a state in which neither the first power transmission path PT1 nor the second power transmission path PT2 is formed. The R operation position is a reverse travel operation position that selects the R position of the power transmission device 16 that enables reverse travel in the gear travel mode. The N operation position is a neutral operation position that selects the N position of the power transmission device 16 in which the power transmission device 16 is in the neutral state. The D operation position is the D of the power transmission device 16 that enables forward traveling in the gear traveling mode or by executing automatic transmission control of the continuously variable transmission mechanism 24 in the belt traveling mode. It is a forward running operation position that selects a position. Therefore, the D operation position and the R operation position are each running operation positions for setting the power transmission device 16 into a power transmission capable state.

The electronic control device 90 includes an engine control means, that is, an engine control unit 92, and a shift control means, that is, a shift control unit 94, in order to realize various controls in the vehicle 10.

The engine control unit 92 applies the accelerator operation amount θacc and the vehicle speed V to, for example, a driving force map, which is a relationship that is experimentally or designedly obtained and stored, that is, a predetermined relationship, so that the target driving force Fwtgt Is calculated. The engine control unit 92 sets a target engine torque Tetgt from which the target driving force Fwtgt can be obtained, and outputs an engine control command signal Se that controls the engine 12 so that the target engine torque Tetgt can be obtained to the engine control device 42. ..

When the operation position POSsh is the P operation position or the N operation position while the vehicle is stopped, the shift control unit 94 is a hydraulic control command signal that engages the meshing clutch D1 in preparation for the transition to the gear traveling mode. Scbd is output to the hydraulic control circuit 46. When the operation position POSsh is changed from the P operation position or the N operation position to the D operation position while the vehicle is stopped, the shift control unit 94 sends the hydraulic control command signal Scbd that engages the first clutch C1 to the hydraulic control circuit 46. Output. As a result, the traveling mode is shifted to the gear traveling mode that enables forward traveling. When the operation position POSsh is changed from the P operation position or the N operation position to the R operation position while the vehicle is stopped, the shift control unit 94 sends the hydraulic control command signal Scbd that engages the first brake B1 to the hydraulic control circuit 46. Output. As a result, the traveling mode is shifted to the gear traveling mode that enables reverse traveling.

When the operation position POSsh is the D operation position, the shift control unit 94 executes switching control for switching between the gear travel mode and the belt travel mode. Specifically, the shift control unit 94 corresponds to the first speed shift stage corresponding to the gear ratio EL of the gear mechanism 28 in the gear travel mode and the lowest gear ratio γmax of the continuously variable transmission mechanism 24 in the belt travel mode. The vehicle speed V and the accelerator operation amount θacc are applied to the upshift line and the downshift line as the stepped shift map, which has a predetermined relationship and has a predetermined hysteresis for switching to the second gear. By doing so, the necessity of shifting is determined, and the driving mode is switched based on the determination result.

When the shift control unit 94 determines upshift while traveling in the gear traveling mode and switches to the belt traveling mode, the clutch that releases the first clutch C1 and re-grasps the clutch so as to engage the second clutch C2. The hydraulic control command signal Scbd for shifting the clutch is output to the hydraulic control circuit 46. As a result, the power transmission path PT in the power transmission device 16 is switched from the first power transmission path PT1 to the second power transmission path PT2. As described above, the shift control unit 94 changes from the gear traveling mode in which the first power transmission path PT1 is formed by the stepped shift control by the release of the first clutch C1 and the engagement of the second clutch C2. 2 The upshift of the power transmission device 16 for switching to the belt traveling mode in which the power transmission path PT2 is formed is executed. In this embodiment, the upshift of the power transmission device 16 for switching from the gear traveling mode to the belt traveling mode is referred to as a stepped upshift.

When the shift control unit 94 determines the downshift while traveling in the belt traveling mode and switches to the gear traveling mode, the clutch that releases the second clutch C2 and re-grasps the clutch so as to engage the first clutch C1. The hydraulic control command signal Scbd that performs the two-clutch shift is output to the hydraulic control circuit 46. As a result, the power transmission path PT in the power transmission device 16 is switched from the second power transmission path PT2 to the first power transmission path PT1. As described above, the shift control unit 94 changes from the belt traveling mode in which the second power transmission path PT2 is formed by the stepped shift control by the release of the second clutch C2 and the engagement of the first clutch C1. 1 The downshift of the power transmission device 16 for switching to the gear traveling mode in which the power transmission path PT1 is formed is executed. In this embodiment, the downshift of the power transmission device 16 that switches from the belt traveling mode to the gear traveling mode is referred to as a stepped downshift.

In the switching control for switching between the gear traveling mode and the belt traveling mode, the torque is transferred by the clutch-to-clutch shift by passing through the state of the belt traveling mode in the medium vehicle speed region in which the meshing clutch D1 is engaged. Since the first power transmission path PT1 and the second power transmission path PT2 are switched only by this, the switching shock is suppressed.

In the belt traveling mode, the shift control unit 94 sets the primary pressure Pin and the secondary pressure Pout so as to achieve the target shift ratio γcvttgt of the continuously variable transmission mechanism 24 while preventing the belt slip of the continuously variable transmission mechanism 24. The hydraulic pressure control command signal Scvt that controls the above is output to the hydraulic pressure control circuit 46, and the continuously variable transmission mechanism 24 is changed. The hydraulic pressure control command signal Scvt is a primary instruction pressure Spin for setting the primary pressure Pin as the target primary pressure Pintgt, and a secondary instruction pressure Spout for setting the secondary pressure Pout as the target secondary pressure Pouttgt.

The target primary pressure Pintgt is a target value of the primary pressure Pin that causes the target thrust of the primary pulley 60, that is, the target value of the primary thrust Win, which is the target value of the primary target thrust Wintgt. The target secondary pressure Pouttgt is a target value of the secondary pressure Pout that causes a secondary target thrust Wouttgt, which is a target value of the target thrust of the secondary pulley 64, that is, the secondary thrust Wout. In the calculation of the primary target thrust Wintgt and the secondary target thrust Wouttgt, the required thrust, which is the thrust required to prevent the belt slip of the continuously variable transmission mechanism 24 with the minimum necessary thrust, is taken into consideration. This required thrust is the belt slip limit thrust Wlmt, which is the thrust immediately before the belt slip of the continuously variable transmission mechanism 24 occurs. In this embodiment, the belt slip limit thrust Wlmt is referred to as a slip limit thrust Wlmt.

Specifically, the shift control unit 94 calculates the primary target thrust Wintgt and the secondary target thrust Wouttgt, respectively. The shift control unit 94 has a secondary thrust Wout calculated based on the primary side slip limit thrust Winlmt, which is the slip limit thrust Wlmt in the primary pulley 60, and a secondary side, which is the slip limit thrust Wlmt in the secondary pulley 64, as the secondary target thrust Wouttgt. Select the larger thrust of the slip limit thrust Woutlmt. The secondary thrust Wout calculated based on the primary side slip limit thrust Winlmt is the secondary side shift control thrust Woutsh, which is the thrust required for shift control on the secondary pulley 64 side, as will be described later.

The shift control unit 94 sets the primary thrust Win calculated based on the secondary target thrust Wouttgt as the primary target thrust Wintgt. The primary thrust Win calculated based on the secondary target thrust Wouttgt is the primary side shift control thrust Winsh, which is the thrust required for shift control on the primary pulley 60 side, as will be described later. Further, as will be described later, the shift control unit 94 controls the primary shift control thrust by feedback control of the primary thrust Win based on the shift ratio deviation Δγcvt (= γcvttgt−γcvt) between the target shift ratio γcvttgt and the actual shift ratio γcvt. Correct Winsh, that is, correct the primary target thrust Wintgt.

In the correction of the shift control thrust Winsh on the primary side described above, the deviation between the target value and the actual value in the parameter corresponding to the shift ratio γcvt and 1: 1 may be used instead of the shift ratio deviation Δγcvt. For example, in the correction of the shift control thrust Winsh on the primary side, the deviation ΔXin (= Xintgt-Xin) between the target pulley position Xintgt and the actual pulley position Xin (see FIG. 2) in the primary pulley 60 and the target pulley position Xouttgt in the secondary pulley 64. Deviation ΔXout (= Xouttgt-Xout) from the actual pulley position Xout (see FIG. 2), deviation ΔRin (= Rintgt-Rin) between the target belt hanging diameter Rintgt and the actual belt hanging diameter Rin (see FIG. 2) in the primary pulley 60. , Deviation ΔRout (= Routtgt-Rout) between the target belt hanging diameter Routtgt and the actual belt hanging diameter Rout (see FIG. 2) in the secondary pulley 64, Deviation ΔNpri (= Npritgt) between the target primary rotation speed Npritgt and the actual primary rotation speed Npri. -Npri) and the like can be used.

The thrust required for the shift control described above is the thrust required to realize the target shift, and is the thrust required to achieve the target gear ratio γcvttgt and the target shift speed dγtgt. The shift speed dγ is, for example, the amount of change (= dγcvt / dt) of the shift ratio γcvt per unit time. In this embodiment, the shift speed dγ is defined as the pulley position movement amount (= dX / dNelm) per element of the transmission belt 66. “DX” is the amount of axial displacement of the pulley [mm / ms] per unit time, and “dNelm” is the number of elements [pieces / ms] that bite into the pulley per unit time. The shift speed dγ is represented by a primary shift speed dγin (= dXin / dNelmin) and a secondary shift speed dγout (= dXout / dNelmout).

Specifically, the thrust of each of the pulleys 60 and 64 in a steady state in which the gear ratio γcvt is constant is referred to as a balanced thrust Wbl. Balanced thrust Wbl is also a steady thrust. The balanced thrust Wbl of the primary pulley 60 is the primary balanced thrust Winbl, the balanced thrust Wbl of the secondary pulley 64 is the secondary balanced thrust Woutbl, and these ratios are the thrust ratio τ (= Woutbl / Winbl). On the other hand, when a certain thrust is added or subtracted to any of the thrusts of the pulleys 60 and 64 in the steady state, the steady state collapses and the gear ratio γcvt changes, and the magnitude of the added or subtracted thrust is obtained. The shift speed dγ corresponding to the above is generated. This added or subtracted thrust is referred to as a shift difference thrust ΔW. Hereinafter, the shift difference thrust ΔW is referred to as a differential thrust ΔW. The differential thrust ΔW is also a transient thrust. The differential thrust ΔW when the target shift is realized on the primary pulley 60 side is the primary differential thrust ΔWin as the differential thrust ΔW converted on the primary pulley 60 side. The differential thrust ΔW when the target shift is realized on the secondary pulley 64 side is the secondary differential thrust ΔWout as the differential thrust ΔW converted to the secondary pulley 64 side.

The thrust required for the above-mentioned shift control is to realize the target shift ratio γcvttgt corresponding to one thrust based on the thrust ratio τ for maintaining the target shift ratio γcvttgt when one thrust is set. It is the sum of the other balanced thrust Wbl and the differential thrust ΔW for realizing the target shift speed dγtgt when the target shift ratio γcvttgt is changed. The target shift speed dγtgt is represented by a primary target shift speed dγintgt and a secondary target shift speed dγouttgt. The primary differential thrust ΔWin has a positive value exceeding zero, that is, “ΔWin> 0” in the upshift state, and a negative value less than zero, that is, “ΔWin <0” in the downshift state, and is in a steady state with a constant gear ratio. If so, it becomes zero, that is, “ΔWin = 0”. Further, the secondary difference thrust ΔWout has a negative value less than zero, that is, “ΔWout <0” in the upshift state, and a positive value exceeding zero, that is, “ΔWout> 0” in the downshift state, and the gear ratio is constant. In the steady state, it becomes zero, that is, "ΔWout = 0".

FIG. 3 is a diagram for explaining the thrust required for the shift control described above. FIG. 4 is a diagram showing an example of the relationship between each thrust at the time point t2 in FIG. 3 and 4 show the primary set when the target upshift is realized on the primary pulley 60 side when the secondary thrust Wout is set so as to prevent the belt slip on the secondary pulley 64 side, for example. An example of thrust Win is shown. In FIG. 3, since the target gear ratio γcvttgt is in a constant steady state and ΔWin = 0 before the time t1 or after the time t3, the primary thrust Win is only the primary balance thrust Winbl (= Wout / τ). At the time point t1 − time point 3 t3, the target gear ratio γcvttgt is in an upshift state, so that the primary thrust Win is the sum of the primary balance thrust Winbl and the primary differential thrust ΔWin as shown in FIG. The shaded portion of each thrust shown in FIG. 4 corresponds to each balanced thrust Wbl for maintaining the target gear ratio γcvttgt at t2 in FIG.

FIG. 5 is a block diagram showing a control structure for achieving both target shifting and belt slip prevention with the minimum necessary thrust, and is a diagram illustrating hydraulic control, that is, CVT hydraulic control in the continuously variable transmission mechanism 24. be.

In FIG. 5, the shift control unit 94 calculates the target gear ratio γcvttgt. Specifically, the shift control unit 94 calculates the target primary rotation speed Npritgt by applying the accelerator operation amount θacc and the vehicle speed V to, for example, the CVT shift map, which has a predetermined relationship. The shift control unit 94 calculates the post-shift target shift ratio γcvttgtl (= Npritgt / Nsec), which is the shift ratio γcvt to be achieved after shifting of the continuously variable transmission mechanism 24, based on the target primary rotation speed Npritgt. The shift control unit 94 is shifting based on the shift ratio γcvt before the start of the shift, the target shift ratio γcvttgtl after the shift, and their differences, for example, in a predetermined relationship so as to realize a quick and smooth shift. The target gear ratio γcvttgt is determined as the target value of the transient gear ratio γcvt. For example, the shift control unit 94 determines the target shift ratio γcvttgt to be changed during the shift as a function of the elapsed time that changes along a smooth curve that changes from the start of the shift toward the target shift ratio γcvttgtl after the shift. This smooth curve is, for example, a first-order lag curve or a second-order lag curve. When the shift control unit 94 determines the target shift ratio γcvttgt, the shift control unit 94 calculates the target shift speed dγtgt during the shift based on the target shift ratio γcvttgt. When the shift is completed and the target shift ratio γcvttgt becomes a constant steady state, the target shift speed dγtgt is set to zero.

The shift control unit 94 calculates the input torque to the continuously variable transmission mechanism 24 used for calculating the primary target thrust Wintgt and the secondary target thrust Wouttgt. The input torque to the stepless speed change mechanism 24 is the belt input torque Tb1 for calculating the thrust ratio as the first input torque used for calculating the thrust ratio τ that realizes the target speed change ratio γcvttgt of the stepless speed change mechanism 24, and It is a belt input torque Tb2 for preventing belt slip as a second input torque used for each calculation of the primary side slip limit thrust Winlmt and the secondary side slip limit thrust Woutlmt. In this embodiment, the belt input torque Tb1 for calculating the thrust ratio is referred to as the input torque Tb1 for calculating the thrust ratio, and the belt input torque Tb2 for preventing the belt slip is referred to as the input torque Tb2 for preventing the belt slip.

Specifically, the shift control unit 94 calculates the estimated value of the engine torque Te by applying the throttle opening tap and the engine rotation speed Ne to, for example, the engine torque map, which is a predetermined relationship. The shift control unit 94 calculates the turbine torque Tt based on the estimated value of the engine torque Te and, for example, the characteristics of the torque converter 20 which is a predetermined relationship. This turbine torque Tt is an estimated value of the input torque to the continuously variable transmission mechanism 24. The shift control unit 94 uses this turbine torque Tt as the input torque Tb1 for calculating the thrust ratio.

As the input torque Tb2 for belt slip prevention, basically, the input torque Tb1 for calculating the thrust ratio may be used. However, it is not preferable that the slip limit thrust Wlmt is set to zero when the input torque Tb1 for calculating the thrust ratio is zero, in consideration of variations and the like. Therefore, as the belt slip prevention input torque Tb2, a torque to which a lower limit guard process is applied to the absolute value of the thrust ratio calculation input torque Tb1 is used. The shift control unit 94 selects, as the belt slip prevention input torque Tb2, the larger of the absolute value of the thrust ratio calculation input torque Tb1 and the minimum guaranteed torque Tblim. This minimum guaranteed torque Tblim is, for example, a predetermined lower limit torque for increasing the belt slip prevention input torque Tb2 to the safe side in consideration of variation in order to prevent belt slip, and is a positive torque. .. When the input torque Tb1 for calculating the thrust ratio is a negative value, considering that the torque accuracy is low, a predetermined torque corresponding to the input torque Tb1 for calculating the thrust ratio is used as the input torque Tb2 for preventing belt slippage. May be done. This predetermined torque is, for example, a torque having a positive value predetermined as a value larger than the absolute value of the input torque Tb1 for calculating the thrust ratio. As described above, the belt slip prevention input torque Tb2 is a torque based on the thrust ratio calculation input torque Tb1.

In the block B1 and the block B2 of FIG. 5, the shift control unit 94 calculates the slip limit thrust Wlmt based on the actual gear ratio γcvt and the belt slip prevention input torque Tb2. Specifically, the shift control unit 94 calculates the secondary side slip limit thrust Woutlmt using the following equation (1). The shift control unit 94 calculates the primary side slip limit thrust Winlmt using the following equation (2). In the following equations (1) and (2), "Tb2" is the belt slip prevention input torque Tb2, and "Tout" is the belt slip prevention input torque Tb2 converted to the secondary pulley 64 side (= γcvt × Tb2). = (Rout / Rin) × Tb2), “α” is the sheave angle of each pulley 60, 64, “μin” is the coefficient of friction between a predetermined element and the pulley in the primary pulley 60, and “μout” is the predetermined coefficient in the secondary pulley 64. The coefficient of friction between the element and the pulley, "Rin" is the belt hook diameter in the primary pulley 60 uniquely calculated from the actual gear ratio γcvt, and "Rout" is the belt in the secondary pulley 64 uniquely calculated from the actual gear ratio γcvt. The coefficient of friction (see FIG. 2).

Woutlmt = (Tout × cosα) / (2 × μout × Rout)
= (Tb2 × cosα) / (2 × μout × Rin)… (1)
Winlmt = (Tb2 × cosα) / (2 × μin × Rin)… (2)

As the slip limit thrust Wlmt, for example, a value to which the lower limit guard process is applied to the calculated slip limit thrust Wlmt may be used. The shift control unit 94 has, for example, as the primary side slip limit thrust Winlmt used in the block B3 of FIG. 5, the larger of the primary side slip limit thrust Winlmt calculated by using the above equation (2) and the primary side minimum thrust Winmin. Select one thrust. The primary-side minimum thrust Winmin is the hard limit minimum thrust of the primary pulley 60, including the thrust generated by the control variation of the primary pressure Pin and the thrust generated by the centrifugal hydraulic pressure in the hydraulic actuator 60c. The control variation of the primary pressure Pin is a predetermined primary that may be supplied from the hydraulic control circuit 46 to the hydraulic actuator 60c even if the primary instruction pressure Spin having the primary pressure Pin of zero is output. This is the maximum value of the pressure Pin. The same applies to the secondary side slip limit thrust Woutlmt.

In the block B3 and the block B6 of FIG. 5, the shift control unit 94 calculates the balance thrust Wbl. That is, the shift control unit 94 calculates the secondary balance thrust Woutbl for the primary side slip limit thrust Winlmt and the primary balance thrust Winbl for the secondary target thrust Wouttgt, respectively.

^{Specifically, the shift control unit 94 applies the target gear ratio γcvttgt and the inverse number SFin -1} of the primary safety factor SFin to the thrust ratio map map (τin) as shown in FIG. 6, for example, to achieve the target gear ratio. Calculate the thrust ratio τin that realizes γcvttgt. The thrust ratio map map (τin) is a diagram showing an example of the relationship between the ^{reciprocal SFin -1} of the primary safety factor and the thrust ratio τin, which are predetermined with the target gear ratio γcvttgt as a parameter. The thrust ratio τin is a thrust ratio for calculating the secondary thrust as a thrust ratio used when calculating the thrust on the secondary pulley 64 side based on the thrust on the primary pulley 60 side. The shift control unit 94 calculates the secondary balance thrust Woutbl based on the primary side slip limit thrust Winlmt and the thrust ratio τin using the following equation (3). The primary safety factor SFin is, for example, "Win / Winlmt" or "Tb2 / Tb1", and the reciprocal SFin ^-1 of the primary safety factor is, for example, "Winlmt / Win" or "Tb1 / Tb2". be. Further, the shift control unit 94 realizes the target shift ratio γcvttgt by applying the ^{target shift ratio γcvttgt and the inverse number SFout -1} of the secondary safety factor SFout to the thrust ratio map map (τout) as shown in FIG. 7, for example. Calculate the thrust ratio τout. The thrust ratio map map (τout) is a diagram showing an example of the relationship between the ^{reciprocal SFout -1} of the secondary safety factor and the thrust ratio τout, which are predetermined with the target gear ratio γcvttgt as a parameter. The thrust ratio τout is a thrust ratio for calculating the primary thrust as a thrust ratio used when calculating the thrust on the primary pulley 60 side based on the thrust on the secondary pulley 64 side. The shift control unit 94 calculates the primary balance thrust Winbl based on the secondary target thrust Wouttgt and the thrust ratio τout using the following equation (4). The secondary safety factor SFout is, for example, "Wout / Woutlmt" or "Tb2 / Tb1", and the reciprocal SFout ^-1 of the secondary safety factor is, for example, "Woutlmt / Wout" or "Tb1 / Tb2". be. ^{Since the input torque Tb2 for belt slip prevention is always a positive value, the reciprocals SFin -1} and SFout of the above safety factors are used when the vehicle 10 such that the input torque Tb1 for thrust ratio calculation has a positive value is in the driving state. ^{Since -1 is} also a positive value, the value in the drive region is used as the thrust ratio τ. On the other hand, when the vehicle 10 in which the input torque Tb1 for calculating the thrust ratio has a negative value is in the driven state, the reciprocals SFin ^-1 and SFout ^-1 of the above safety factors also have negative values, so that the thrust ratio τ is The value of the driven region is used. The reciprocals SFin ^-1 and SFout ^-1 may be calculated each time the balance thrust Wbl is calculated, but if the safety factors SFin -1 and SFout are set to predetermined values (for example, about 1-1.5), respectively. If so, the reciprocal of it may be set.

Woutbl ＝ Winlmt × τin… (3)
Winbl = Wouttgt / τout… (4)

As described above, the slip limit thrust Winlmt and Woutlmt are calculated based on the belt slip prevention input torque Tb2 based on the thrust ratio calculation input torque Tb1. ^{The reciprocals SFin -1} and SFout ^-1 of the above safety factors, which are the basis for calculating the thrust ratios τin and τout, are values based on the input torque Tb1 for calculating the thrust ratio. Therefore, the shift control unit 94 calculates the thrust ratio τ that realizes the target shift ratio γcvttgt of the continuously variable transmission mechanism 24 based on the input torque Tb1 for calculating the thrust ratio.

In the block B4 and the block B7 of FIG. 5, the shift control unit 94 calculates the differential thrust ΔW. That is, the shift control unit 94 calculates the secondary differential thrust ΔWout and the primary differential thrust ΔWin.

Specifically, the shift control unit 94 calculates the secondary differential thrust ΔWout by applying the secondary target shift speed dγouttgt to the differential thrust map (ΔWout) as shown in FIG. 8, for example. The differential thrust map map (ΔWout) is a diagram showing an example of the relationship between the predetermined secondary shift speed dγout and the secondary differential thrust ΔWout. The shift control unit 94 calculates the secondary shift control thrust Woutsh (= Woutbl + ΔWout) by adding the secondary differential thrust ΔWout to the secondary balance thrust Woutbl as the secondary thrust required to prevent the belt slip on the primary pulley 60 side. Further, the shift control unit 94 calculates the primary differential thrust ΔWin by applying the primary target shift speed dγintgt to the differential thrust map (ΔWin) as shown in FIG. 9, for example. The differential thrust map map (ΔWin) is a diagram showing an example of the relationship between the predetermined primary shift speed dγin and the primary differential thrust ΔWin. The shift control unit 94 calculates the primary shift control thrust Winsh (= Winbl + ΔWin) by adding the primary differential thrust ΔWin to the primary balance thrust Winbl.

In the calculation in the blocks B3 and B4, a predetermined physical characteristic diagram such as a thrust ratio map map (τin) as shown in FIG. 6 and a differential thrust map map (ΔWout) as shown in FIG. 8 is used. Therefore, there are variations in the physical characteristics in the calculation results of the secondary balance thrust Woutbl and the secondary differential thrust ΔWout due to individual differences in the hydraulic control circuit 46 and the like. Therefore, when considering such variations in physical characteristics, the shift control unit 94 may add the control margin Wmgn to the primary side slip limit thrust Winlmt. The control margin Wmgn is a predetermined predetermined thrust corresponding to the variation in the physical characteristics related to the calculation of the secondary balance thrust Woutbl and the secondary differential thrust ΔWout. When considering the variation in the physical characteristics as described above, the shift control unit 94 uses the equation “Woutbl ＝ (Winlmt + Wmgn) × τin” shown in FIG. 5 instead of the equation (3) to be secondary. Calculate the balance thrust Woutbl. The variation in the physical characteristics is different from the variation in the actual pulley hydraulic pressure with respect to the hydraulic control command signal Scvt. The variation in the pulley hydraulic pressure is a relatively large value depending on the hard unit such as the hydraulic control circuit 46, but the variation in the physical characteristics is an extremely small value as compared with the variation in the pulley hydraulic pressure.

In block B5 of FIG. 5, the shift control unit 94 selects the larger thrust of the secondary side slip limit thrust Woutlmt and the secondary side shift control thrust Woutsh as the secondary target thrust Wouttgt.

In the block B8 of FIG. 5, the shift control unit 94 calculates the feedback control amount Winfb. Specifically, the shift control unit 94 uses a feedback control formula as shown in the following equation (5), for example, to match the actual gear ratio γcvt with the target gear ratio γcvttgt (feedback control amount). = FB control amount) Calculate Winfb. In the following equation (5), "Δγcvt" is a gear ratio deviation Δγcvt, "Kp" is a predetermined proportionality constant, "Ki" is a predetermined integral constant, and "Kd" is a predetermined differential constant. The shift control unit 94 calculates the value (= Winsh + Winbb) after correcting the primary shift control thrust Winsh by feedback control by adding the feedback control amount Winfb to the primary shift control thrust Winsh as the primary target thrust Wintgt. ..

Winfb ＝ Kp × Δγcvt ＋ Ki × (∫Δγcvtdt) ＋ Kd × (dΔγcvt / dt)… (5)

In the block B9 and the block B10 of FIG. 5, the shift control unit 94 converts the target thrust into the target pulley pressure. Specifically, the shift control unit 94 sets the secondary target thrust Wouttgt and the primary target thrust Wintgt as the target secondary pressure Pouttgt (= Wouttgt / pressure receiving area) and the target, respectively, based on the pressure receiving areas of the hydraulic actuators 60c and 64c, respectively. It is converted to the primary pressure Pintgt (= Wintgt / pressure receiving area) respectively. The shift control unit 94 sets the target secondary pressure Pouttgt and the target primary pressure Pintgt as the secondary instruction pressure Spout and the primary instruction pressure Spin, respectively.

The shift control unit 94 outputs the primary instruction pressure Spin and the secondary instruction pressure Spout as the hydraulic control command signal Scvt to the hydraulic control circuit 46 so that the target primary pressure Pintgt and the target secondary pressure Pouttgt can be obtained. The hydraulic pressure control circuit 46 regulates the primary pressure Pin and the secondary pressure Pout according to the hydraulic pressure control command signal Scvt.

Here, in the power transmission device 16, the operating state of the second clutch C2 represented by the four states of complete release, complete engagement, release transient, and engagement transient differs depending on the traveling mode and the like. For example, in the belt traveling mode, the second clutch C2 is in a fully engaged state, while in the gear traveling mode, the second clutch C2 is in a completely released state. Further, in the switching control for switching between the gear traveling mode and the belt traveling mode, the second clutch C2 is temporarily in the release transient state or the engagement transient state. Further, in the belt traveling mode, when the garage operation in which the shift lever 84 is operated between the N operation position and the D operation position is performed, the second clutch C2 is temporarily in the release transient state or the engagement transient state. It is said that. If the operating state of the second clutch C2 is different, the input torque to the continuously variable transmission mechanism 24 is different. That is, the input torque to the continuously variable transmission mechanism 24 is a torque according to the operating state of the second clutch C2.

In addition to the belt traveling mode, it is desirable to realize the target gear ratio γcvttgt of the continuously variable transmission mechanism 24 while preventing the belt slip of the continuously variable transmission mechanism 24 as in the belt traveling mode. Therefore, the shift control unit 94 calculates the input torque to the continuously variable transmission mechanism 24 used for calculating the primary target thrust Wintgt and the secondary target thrust Wouttgt according to the operating state of the second clutch C2. That is, the shift control unit 94 calculates the thrust ratio calculation input torque Tb1 and the belt slip prevention input torque Tb2, respectively, according to the operating state of the second clutch C2.

The shift control unit 94 is like the calculation method of the primary target thrust Wintgt and the secondary target thrust Wouttgt described by exemplifying the case where the vehicle 10 is in the belt traveling mode when the second clutch C2 is in the fully engaged state. The input torque Tb1 for calculating the thrust ratio is the turbine torque Tt, and the input torque Tb2 for preventing belt slippage is the turbine torque Tt in consideration of the minimum guaranteed torque Tblim.

When the second clutch C2 is in the completely released state, the shift control unit 94 sets the thrust ratio calculation input torque Tb1 to zero, sets the belt slip prevention input torque Tb2, and sets the drag torque of the second clutch C2 to the primary shaft 58. The torque value converted above, that is, the drag torque of the second clutch C2 converted on the input shaft 22. The drag torque of the second clutch C2 is, for example, a predetermined torque when the second clutch C2 is in a completely released state.

When the second clutch C2 is in the engagement transient state, the shift control unit 94 puts the input torque Tb1 for calculating the thrust ratio on the torque value obtained by converting the C2 clutch torque Tcltc2 on the primary shaft 58, that is, on the input shaft 22. Let it be the converted C2 clutch torque Tcltc2. When the second clutch C2 is in the engagement transient state, the shift control unit 94 selects the larger torque of the thrust ratio calculation input torque Tb1 and the minimum guaranteed torque Tblim as the belt slip prevention input torque Tb2. That is, the input torque Tb2 for preventing belt slippage is the C2 clutch torque Tcltc2 converted on the input shaft 22 in consideration of the minimum guaranteed torque Tblim. The shift control unit 94 calculates the C2 clutch torque Tcltc2 by applying the C2 instruction pressure to, for example, a predetermined C2 clutch torque map. The minimum guaranteed torque Tblim here may be the same value as when the second clutch C2 is in the fully engaged state, or a different value may be used.

When the second clutch C2 is in the release transient state, the shift control unit 94 determines the smaller torque of the turbine torque Tt and the C2 clutch torque Tcltc2 converted on the input shaft 22 as the input torque Tb1 for calculating the thrust ratio. Select. When the second clutch C2 is in the release transient state, the shift control unit 94 converts the input torque Tb2 for belt slip prevention into the input torque Tb1 for calculating the thrust ratio, the turbine torque Tt, and the input shaft 22 as the second clutch C2. The larger of the drag torques of the clutch C2 is selected.

By the way, when the second clutch C2 is in the completely released state, as shown in the equation of motion of the following equation (6), the rotation change of the primary shaft 58 and the rotation change of the secondary shaft 62 due to the rotation change of the input shaft 22. Along with this, an inner shuttle torque Tbint is generated. This inertia shuttle Tbint is a torque converted on the primary shaft 58, and corresponds to an inertia shuttle that is input to the continuously variable transmission mechanism 24 as the rotation of the input shaft 22 changes. This inertia shuttle Tbint is the input torque to the continuously variable transmission mechanism 24. In the following equation (6), "Ipri" is the moment of inertia of the primary pulley 60, "Isec" is the moment of inertia of the secondary pulley 64, "dNpri / dt" is the time derivative of the primary rotation speed Npri, that is, the time change rate, "dNsec /". "dt" is the time derivative of the secondary rotation speed Nsec, that is, the time change rate. The primary rotation speed Npri corresponds to the angular velocity of the primary pulley 60, and "dNpri / dt" corresponds to the angular acceleration of the primary pulley 60. The secondary rotation speed Nsec corresponds to the angular velocity of the secondary pulley 64, and "dNsec / dt" corresponds to the angular acceleration of the secondary pulley 64.

Tbint = Ipri × (dNpri / dt) ＋ (1 / γcvt) × Isec × (dNsec / dt)… (6)

As shown in the above equation (6), the inertia torque Tbint is determined based on the angular acceleration of each of the pulleys 60 and 64 and the gear ratio γcvt of the continuously variable transmission mechanism 24. On the other hand, as described above, when the second clutch C2 is in the completely released state, the control belt slip prevention input torque Tb2 is regarded as the drag torque of the second clutch C2 converted on the input shaft 22. The slip limit thrust Winlmt and Woutlmt are calculated. When the inner shuttlek Tbint exceeds the drag torque of the second clutch C2 converted on the input shaft 22, the belt pinching pressure, that is, the belt torque capacity Tcvt is insufficient, and the belt slip of the continuously variable transmission mechanism 24 may occur. The inertia torque Tbint is increased as the angular acceleration of the pulleys 60 and 64 is larger and the gear ratio γcvt is the gear ratio on the high side. Under the condition that the angular acceleration of the pulleys 60 and 64 is large and the gear ratio γcvt is the gear ratio on the high side, the belt slip as described above is likely to occur. The large angular acceleration of the pulleys 60 and 64 means that the angular acceleration of the input shaft 22 is large.

There is no situation where the angular acceleration of the pulleys 60 and 64 is large and the gear ratio γcvt is the gear ratio on the high side, that is, there is no large inner shuttlek Tbint that exceeds the drag torque of the second clutch C2 converted on the input shaft 22. The situation that may be input to the speed change mechanism 24 is, for example, when the speed change ratio γcvt of the stepless speed change mechanism 24 is not the lowest side speed change ratio γmax and the belt return failure occurs, for example, the speed change ratio γcvt is on the high side. When the gear ratio on the highest side is near γmin, it is assumed that the first clutch C1 or the first brake B1 is suddenly engaged.

As a scene where sudden engagement of the first clutch C1 or the first brake B1 occurs, for example, in the gear traveling mode in which the second clutch C2 is completely released, the first clutch C1 accompanies the N → D garage operation. The first brake B1 is being executed during the C1 garage engagement control, which is the engagement control of the first clutch C1 that switches from the released state to the engaged state, that is, during the C1 garage engagement transient control, or with the N → R garage operation. It is assumed that the B1 garage engagement control, which is the engagement control of the first clutch B1 that switches the clutch from the released state to the engaged state, is being executed, that is, the B1 garage engagement transient control is being performed. The N → D garage operation and the N → R garage operation are shift operations by the driver that switches the shift lever 84 from the N operation position to the traveling operation position, respectively.

As the belt return failure in which the belt slip is likely to occur as described above, for example, it is assumed that the gear ratio γcvt of the continuously variable transmission mechanism 24 is higher than the predetermined gear ratio. This predetermined gear ratio is a predetermined high-side limit gear ratio that enables belt slip prevention even during C1 garage engagement transient control or B1 garage engagement transient control, for example.

In response to the possibility of belt slippage due to the input of the large inner shuttle clutch Tbint to the continuously variable transmission mechanism 24 as described above, the shift control unit 94 converts the input torque Tb2 for preventing belt slippage onto the input shaft 22. By increasing the slip limit thrusts Winlmt and Woutlmt calculated using the drag torque of the second clutch C2 by a predetermined thrust, belt slip occurs when a large inner shuttle torque Tbint is input to the continuously variable transmission mechanism 24. To prevent.

The electronic control device 90 further has a state determination means, that is, a state determination unit, in order to realize a control function of preventing the occurrence of belt slip in a situation where a large inertia torque Tbint is input to the above-mentioned continuously variable transmission mechanism 24. It is equipped with 106.

The state determination unit 106 has either condition 1 that the second clutch C2 is in the completely released state and that the C1 garage engagement transient control is in progress, or condition 2 that the B1 garage engagement transient control is in progress. Is determined whether or not is established. Further, the state determination unit 106 determines whether or not the gear ratio γcvt of the continuously variable transmission mechanism 24 is on the higher side than the predetermined gear ratio.

The shift control unit 94 determines that any one of the condition 1 and the condition 2 is satisfied when the second clutch C2 is in the completely released state by the state determination unit 106, and the stepless speed change mechanism 24 When it is determined that the gear ratio γcvt is higher than the predetermined gear ratio, the primary side slip limit thrust Winlmt and the secondary side are calculated using the drag torque of the second clutch C2 converted on the input shaft 22. The slip limit thrust Woutlmt is increased by a predetermined thrust. As a result, the secondary target thrust Wouttgt and the primary target thrust Wintgt are increased, so that the primary pressure Pin and the secondary pressure Pout are increased, and the inner shuttle torque Tbint at the time of sudden engagement of the first clutch C1 or the first brake B1 is increased. The belt torque capacity Tcvt is secured. The predetermined thrust is, for example, a predetermined thrust increase amount that guarantees belt slip prevention against the inertia torque Tbint input to the continuously variable transmission mechanism 24 during C1 garage engagement transient control or B1 garage engagement transient control. Is.

The shift control unit 94 converts the guaranteed input torque, which is the inertia shuttle Tbint that should guarantee the belt slip prevention, into a predetermined thrust. For example, the shift control unit 94 sets the value of "Woutlmt" calculated by putting the guaranteed input torque into "Tb2" of the above equation (1) as a predetermined thrust to be added to the secondary side slip limit thrust Woutlmt. The shift control unit 94 sets the value of "Winlmt" calculated by putting the guaranteed input torque into "Tb2" of the above equation (2) as a predetermined thrust to be added to the primary side slip limit thrust Winlmt. The guaranteed input torque may be, for example, a predetermined torque having a predetermined maximum value that may occur as an inertia shuttle Tbint, or a predetermined torque that is increased as the gear ratio γcvt is the gear ratio on the high side. It may be a predetermined torque that becomes the specified value. Even if the predetermined thrust applied to the secondary side slip limit thrust Woutlmt and the predetermined thrust applied to the primary side slip limit thrust Winlmt are the same values using any of the above equations (1) and (2). good.

FIG. 10 shows the occurrence of belt slip in a situation where a large inertia torque Tbint is input to the continuously variable transmission mechanism 24 when the main part of the control operation of the electronic control device 90, that is, the second clutch C2 is in a completely released state. It is a flowchart explaining the control operation for prevention, and is executed repeatedly, for example.

In FIG. 10, first, in step S10 corresponding to the function of the state determination unit 106 (hereinafter, step is omitted), when the second clutch C2 is in the completely released state, the C1 garage engagement transient control is being performed. It is determined whether or not any of the condition 1 and the condition 2 that the B1 garage engagement transient control is in progress is satisfied. If the judgment of S10 is denied, this routine is terminated. If this determination in S10 is affirmed, in S20 corresponding to the function of the state determination unit 106, whether or not the gear ratio γcvt of the continuously variable transmission mechanism 24 is higher than the predetermined gear ratio, that is, the gear ratio γcvt is determined. It is determined whether or not the gear ratio is smaller than the predetermined gear ratio. If the judgment of S20 is denied, this routine is terminated. If the determination in S20 is affirmed, the guaranteed input torque is set to a predetermined torque in S30 corresponding to the function of the shift control unit 94, the guaranteed input torque is converted into a predetermined thrust, and the thrust increase amount is calculated. .. Then, the slip limit thrusts Winlmt and Woutlmt calculated by using the drag torque of the second clutch C2 converted on the input shaft 22 are increased by this thrust increase amount.

As described above, according to the present embodiment, when the second clutch C2 is in the completely released state, the C1 garage engagement transient control or the B1 garage engagement transient control is in progress, and the stepless speed change mechanism. When the gear ratio γcvt of 24 is on the higher side than the predetermined gear ratio, the primary side slip limit thrust Winlmt and the secondary side slip limit calculated by using the drag torque of the second clutch C2 converted on the input shaft 22. Since the thrust Woutlmt is increased by a predetermined thrust, when the gear ratio γcvt is relatively high, the angular acceleration of the input shaft 22 is likely to be increased during C1 garage engagement transient control or B1 garage engagement transient control. Because it is inside, a large inner shuttle torque Tbint that exceeds the drag torque of the second clutch C2 converted on the input shaft 22 is likely to occur, whereas the belt torque capacity Tcvt required to prevent belt slippage is Can be properly secured. Therefore, when the second clutch C2 is in the completely released state, it is possible to prevent the occurrence of belt slip under the condition that a large inertia torque Tbint is input to the continuously variable transmission mechanism 24.

Although the embodiments of the present invention have been described in detail with reference to the drawings, the present invention is also applicable to other aspects.

For example, in the above-described embodiment, in order to prevent the occurrence of belt slip in a situation where a large inertia torque Tbint is input to the continuously variable transmission mechanism 24, the slip limit thrusts Winlmt and Woutlmt required to prevent the belt slip. Was increased by a predetermined thrust, but the present invention is not limited to this mode. For example, since the target thrusts Wintgt and Wouttgt are thrusts in consideration of the slip limit thrusts Winlmt and Woutlmt, the target thrusts Wintgt and Wouttgt can also be regarded as the thrusts Win and Wout required for belt slip prevention. Therefore, in order to prevent the occurrence of belt slip under the condition that a large inertia torque Tbint is input to the continuously variable transmission mechanism 24, the target thrusts Wintgt and Wouttgt may be increased by a predetermined thrust.

Further, in the above-described embodiment, the second clutch C2 is provided in the power transmission path between the secondary pulley 64 and the output shaft 30, but the present invention is not limited to this mode. For example, the secondary shaft 62 may be integrally connected to the output shaft 30, and the primary shaft 58 may be connected to the input shaft 22 via the second clutch C2. That is, the second clutch C2 may be provided in the power transmission path between the primary pulley 60 and the input shaft 22.

Further, in the above-described embodiment, the gear mechanism 28 is a gear mechanism in which one gear stage having a gear ratio on the lower side than the lowest gear ratio γmax of the continuously variable transmission mechanism 24 is formed. It is not limited to the embodiment. For example, the gear mechanism 28 may be a gear mechanism in which a plurality of gear stages having different gear ratios are formed. That is, the gear mechanism 28 may be a stepped transmission that shifts to two or more gears. Alternatively, the gear mechanism 28 is a gear mechanism that forms a gear ratio on the higher side than the highest gear ratio γmin of the continuously variable transmission mechanism 24 and / or a gear ratio on the lower side than the lowest gear ratio γmax. There may be.

Further, in the above-described embodiment, the traveling mode of the power transmission device 16 is switched by using a predetermined upshift line and downshift line, but the present invention is not limited to this mode. For example, the traveling mode of the power transmission device 16 may be switched by calculating the target driving force Fwtgt based on the vehicle speed V and the accelerator operation amount θacc and setting a gear ratio that can satisfy the target driving force Fwtgt. ..

Further, in the above-described embodiment, the torque converter 20 is used as the fluid type transmission device, but the present invention is not limited to this mode. For example, instead of the torque converter 20, another fluid type transmission device such as a fluid coupling having no torque amplification action may be used. Alternatively, this fluid transmission device does not necessarily have to be provided. Further, although the meshing clutch D1 is provided in the first power transmission path PT1 via the gear mechanism 28, the meshing clutch D1 does not necessarily have to be provided in carrying out the present invention.

It should be noted that the above is only one embodiment, and the present invention can be carried out in a mode in which various changes and improvements are made based on the knowledge of those skilled in the art.

12: Engine (power source)
14: Drive wheel 16: Vehicle power transmission device 22: Input shaft (input rotating member)
24: Continuously variable transmission mechanism 28: Gear mechanism 30: Output shaft (output rotating member)
60: Primary pulley 60c: Hydraulic actuator 64: Secondary pulley 64c: Hydraulic actuator 66: Transmission belt (transmission element)
84: Shift lever (shift switching device)
90: Electronic control device (control device)
94: Shift control unit B1: 1st brake (1st engaging device)
C1: 1st clutch (1st engaging device)
C2: 2nd clutch (2nd engaging device)
PT1: 1st power transmission path PT2: 2nd power transmission path

Claims (1)

To transmit the power from the input rotating member to the output rotating member, which is provided in parallel between the input rotating member to which the power of the power source is transmitted and the output rotating member that outputs the power to the drive wheel. The plurality of power transmission paths are the first power transmission path via a gear mechanism having a gear stage formed by the engagement of the first engagement device, and the first power transmission path. Control of a vehicle power transmission device, which is a second power transmission path via a stepless speed change mechanism in which a transmission element is wound between a primary pulley and a secondary pulley, which is formed by engaging the two engagement devices. It ’s a device,
Based on the input torque to the continuously variable transmission mechanism, the thrust of the primary pulley and the secondary force for pinching the transmission element applied by the hydraulic actuator of the primary pulley, which is necessary for preventing the transmission element from slipping, and the secondary. The thrust of the secondary pulley that presses the transmission element applied by the hydraulic actuator of the pulley is calculated, and when the second engaging device is in the completely released state, the input torque to the continuously variable transmission mechanism is calculated. , A shift control unit for the drag torque of the second engaging device converted on the input rotating member.
When the second engaging device is in the completely released state, the shift control unit engages the first engaging device from the released state in accordance with the shift operation by the driver who switches the shift switching device to the traveling operation position. The engagement control of the first engaging device that switches to the matching state is being executed, and the gear ratio of the stepless speed change mechanism is a predetermined high-side limit speed change that enables slip prevention of the transmission element. When it is on the higher side than the ratio, the thrust of the primary pulley and the thrust of the secondary pulley, which are calculated by using the drag torque of the second engaging device and are necessary for preventing the transmission element from slipping. And, respectively, to prevent the transmission element from slipping against the inner shuttle torque input to the stepless speed change mechanism according to the rotation change of the input rotating member during the execution of the engagement control of the first engaging device. A control device for a power transmission device for a vehicle, which is characterized by increasing a predetermined thrust to be guaranteed.

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