JP2018084275A - Control device of power transmission device for vehicle - Google Patents

Abstract

PROBLEM TO BE SOLVED: To suppress an upshift of a belt-type continuously variable transmission during a C-to-C gear change, and to suppress the lowering of the durability of the belt-type continuously variable transmission.SOLUTION: In an electronic control device 100, either of smaller values out of a variation maximum value Tinesmax of estimation input torque which is obtained by dividing a maximum value Tc2max of a variation of a torque capacity Tc2 with respect to the clutch indication pressure PC2dir of a CVT-traveling clutch C2 with a gear change ratio γc of a continuously variable transmission 24, and turbine torque Tt is set to torque capacity calculation estimation input torque Tlmtc, and a variation minimum value Tinesmin of estimation input torque which is obtained by dividing a minimum value Tc2min of the variation of the torque capacity Tc2 with respect to the clutch indication pressure Pc2dir with the gear change ratio γc, or a value which is obtained by subtracting a division value obtained by dividing a prescribed variation range of the torque capacity Tc2 with the gear change ratio γc from the turbine torque Tr is set to thrust ratio calculation estimation input torque Tinc.SELECTED DRAWING: Figure 3

Description

  The present invention relates to a vehicle power transmission device including a first power transmission path via a belt-type continuously variable transmission and a second power transmission path via a gear-type transmission mechanism in parallel. Suppressing the upshift of the belt-type continuously variable transmission during clutch-to-clutch shift (CtoC shift) that selectively selects the first power transmission path and the second power transmission path by switching from one clutch to the other. The present invention relates to a technique for suppressing a decrease in durability of a belt-type continuously variable transmission.

  A first power transmission path and one or a plurality of gear stages passing through a belt-type continuously variable transmission between an input rotating member to which the power of the driving force source is transmitted and an output rotating member for outputting the power to the driving wheels. In parallel with a second power transmission path that passes through a gear-type transmission mechanism, and a first clutch that is engaged to select the first power transmission path is an output side of the belt-type continuously variable transmission In the vehicle power transmission device provided on the downstream side of the torque transmission path with respect to the pulley or the output side pulley, the first power transmission path and the second clutch can be switched from one to the other by switching the first clutch and the second clutch. There is known a control device for a vehicle power transmission device that performs clutch-to-clutch shift (CtoC shift) that selectively selects a second power transmission path. For example, this is the control device for a vehicle power transmission device disclosed in Patent Document 1.

  In the vehicle power transmission device disclosed in Patent Literature 1, the belt-type continuously variable transmission includes a primary pulley that is an input-side pulley having a variable effective diameter provided in the input rotation member, and a coaxial shaft that is coaxial with the output rotation member. A secondary pulley that is an output-side pulley having a variable effective diameter provided on the rotary shaft, and a transmission belt wound around the primary pulley and the secondary pulley are provided. By controlling the thrust applied to the primary pulley and the secondary pulley, the effective diameter of each pulley is changed, and the gear ratio is continuously changed. The first clutch connects / disconnects between the rotating shaft provided with the secondary pulley and the output rotating member. The second clutch connects and disconnects between an input element to which power is input from the driving force source of the differential gear type forward / reverse switching mechanism and an output element indirectly connected to the output rotating member. The belt is disengaged from the second power transmission path via the gear-type transmission mechanism in which one or a plurality of gears are formed by CtoC shift in which the second clutch is released and the first clutch is engaged. It is switched to the first power transmission path via the continuously variable transmission. The first power transmission path is switched to the second power transmission path by a CtoC shift in which the first clutch is released and the second clutch is engaged. During the CtoC shift, the input torque input to the primary pulley of the belt type continuously variable transmission changes according to the torque capacity of the first clutch. Therefore, during the CtoC shift, the input torque to the belt-type continuously variable transmission is estimated based on the torque capacity of the first clutch calculated from the command hydraulic pressure of the first clutch, and the estimated torque is estimated. Based on the input torque, thrust control of the primary pulley and the secondary pulley during the CtoC shift is performed using hydraulic pressure.

Japanese Patent Laid-Open No. 2006-3673

  By the way, the torque capacity of the first clutch calculated from the indicated hydraulic pressure of the first clutch for estimating the input torque of the belt type continuously variable transmission during CtoC shift is, for example, the output side of the first clutch. There may be variations due to various hardware-related specifications such as hydraulic cylinders and friction plates. In order to ensure the torque capacity of the belt-type continuously variable transmission during CtoC shift, the input torque of the belt-type continuously variable transmission estimated based on the maximum value of the torque capacity variation with respect to the command value of the first clutch Estimated input torque for torque capacity calculation used for calculation of thrust for securing the torque capacity of the belt type continuously variable transmission during CtoC shift, and the belt type continuously variable transmission during CtoC shift. It is conceivable to set the estimated input torque for thrust ratio calculation used for calculating the thrust ratio of the machine. However, in this case, the actual input torque to the belt type continuously variable transmission is lower than the estimated input torque set as the estimated input torque for thrust ratio calculation, and the belt type no An unintentional upshift from the lowest gear ratio γmax, which is the target gear ratio during the CtoC gear shift of the stepped transmission and the gear ratio on the lowest vehicle speed side, to the gear ratio according to the actual input torque occurs, and the heat of the clutch And the drivability of the vehicle could be affected.

  Further, in order to maintain the belt type continuously variable transmission at the lowest speed ratio γmax during the CtoC shift regardless of the actual input torque value of the belt type continuously variable transmission, the lowest speed ratio γmax is achieved. For this reason, it is conceivable that the pressure of the input side hydraulic cylinder that applies thrust to the primary pulley is set to the minimum pressure, with the thrust ratio for this purpose being the maximum value. In this case, the torque capacity of the primary pulley is determined by the reaction force of the belt tension due to the thrust to the secondary pulley according to the pressure supplied to the output side hydraulic cylinder. Thus, the lowest speed ratio γmax is maintained regardless of the actual input torque value due to the characteristics of the thrust ratio. However, since the pressure of the input side hydraulic cylinder is lowered more than necessary to the minimum pressure, there is a possibility that the durability of the belt type continuously variable transmission with the optimized physique cannot be ensured.

  The present invention has been made against the background of the above circumstances. The object of the present invention is to suppress the upshift of the belt type continuously variable transmission during the CtoC shift and to improve the durability of the belt type continuously variable transmission. It is providing the control apparatus of the vehicle power transmission device which suppresses a fall of property.

  The gist of the present invention is that a first power passing through a belt-type continuously variable transmission is provided between an input rotating member to which power from a driving force source is transmitted and an output rotating member for outputting the power to driving wheels. A transmission path and a second power transmission path passing through the gear-type transmission mechanism are provided in parallel, and a first clutch engaged to select the first power transmission path is an output side pulley of the belt type continuously variable transmission Alternatively, in the vehicular power transmission device provided downstream of the output side pulley in the torque transmission path, the first power transmission path and the first power transmission path can be changed by switching the first clutch and the second clutch from one to the other. A clutch-to-clutch shift that selectively selects two power transmission paths, and an input torque of the belt-type continuously variable transmission is estimated from a torque capacity of the first clutch; A control device for a vehicle power transmission device for controlling the pressure of an input side hydraulic cylinder provided in an input side pulley of the belt type continuously variable transmission based on a torque capacity of one clutch, wherein the clutch-to-clutch shifting is performed. When calculating the pressure of the input side hydraulic cylinder, the torque capacity calculation for calculating the thrust for securing the torque capacity of the belt type continuously variable transmission as the input torque of the belt type continuously variable transmission Using the estimated input torque and the estimated input torque for calculating the thrust ratio for calculating the thrust ratio of the belt type continuously variable transmission, the estimated input torque for calculating the torque capacity is a torque with respect to a command value to the first clutch. The estimated input torque for thrust ratio calculation is set based on the maximum value of the capacity variation, and the torque capacity variation with respect to the command value to the first clutch is set. It is to set on the basis of the small value.

  According to the present invention, when calculating the pressure of the input side hydraulic cylinder during the clutch-to-clutch shift, the torque capacity of the belt-type continuously variable transmission is ensured as the input torque of the belt-type continuously variable transmission. Using the estimated input torque for calculating the torque capacity for calculating the thrust capacity and the estimated input torque for calculating the thrust ratio for calculating the thrust ratio of the belt-type continuously variable transmission, Is set based on the maximum value of torque capacity variation with respect to the command value to the first clutch, and the estimated input torque for thrust ratio calculation is the minimum value of torque capacity variation with respect to the command value to the first clutch. Set based on. For this reason, the estimated input torque for thrust ratio calculation is set based on the minimum value of the variation in torque capacity with respect to the command value to the first clutch, so that the actual input torque to the belt type continuously variable transmission is It is suppressed from becoming lower than the estimated input torque for thrust ratio calculation, and the upshift of the belt type continuously variable transmission during the CtoC shift is suppressed. Further, for example, in order to achieve the lowest speed ratio during CtoC shift, the input side hydraulic cylinder is compared with the case where the thrust ratio of the belt type continuously variable transmission is maximized and the pressure of the input side hydraulic cylinder is minimized. Therefore, it is possible to suppress a decrease in the durability of the belt type continuously variable transmission.

  Preferably, a first predetermined relationship between a thrust for securing a torque capacity of an output side pulley of the belt type continuously variable transmission and an input torque of the belt type continuously variable transmission, The input torque of the belt-type continuously variable transmission and the thrust ratio of the belt-type continuously variable transmission are calculated using the target gear ratio of the belt-type continuously variable transmission as a parameter for calculating the thrust ratio of the belt-type continuously variable transmission. A predetermined second relationship, and during the clutch-to-clutch shift, the output of the belt-type continuously variable transmission is based on the estimated input torque for torque capacity calculation from the first relationship. The thrust for securing the torque capacity of the side pulley is calculated, and during the clutch-to-clutch shift, the target transmission ratio of the belt-type continuously variable transmission is set as the maximum transmission ratio formed by the belt-type continuously variable transmission. The first Based on the relationship between the maximum transmission ratio and the estimated input torque for thrust ratio calculation, a thrust ratio that achieves the maximum transmission ratio of the belt-type continuously variable transmission is calculated, and the belt-type continuously variable transmission The thrust for securing the torque capacity of the output side pulley is divided by the thrust ratio that achieves the maximum speed ratio of the belt type continuously variable transmission to calculate the target thrust of the input side pulley, The pressure of the input side hydraulic cylinder provided in the input side pulley of the belt type continuously variable transmission is calculated from the target thrust. Therefore, from the second relationship, the thrust ratio calculation estimation set based on the maximum speed ratio of the belt-type continuously variable transmission and the minimum value of the variation in torque capacity with respect to the command value to the first clutch. Based on the input torque, a thrust ratio that achieves the maximum speed ratio of the belt type continuously variable transmission is calculated. Accordingly, it is possible to suppress the upshift of the belt type continuously variable transmission during the CtoC shift, and it is possible to suppress a decrease in durability of the belt type continuously variable transmission.

  Preferably, the second relationship is a relationship in which the thrust ratio for achieving the target gear ratio of the belt type continuously variable transmission increases as the input torque of the belt type continuously variable transmission decreases. is there. Therefore, the thrust ratio calculation estimation input set based on the maximum speed ratio of the belt-type continuously variable transmission and the minimum value of the torque capacity variation with respect to the command value to the first clutch from the second relationship. The thrust ratio calculated based on the torque is calculated based on the maximum transmission ratio of the belt-type continuously variable transmission and the actual input torque of the belt-type continuously variable transmission from the second relationship. Larger than the ratio. As a result, the target gear ratio according to the actual input torque becomes a downshift, and thus the upshift of the belt type continuously variable transmission during the CtoC shift is suppressed.

It is a figure explaining the schematic structure of the vehicle to which the present invention is applied. It is a figure for demonstrating the switching of the running pattern of the power transmission device with which the vehicle of FIG. 1 is equipped. FIG. 2 is a functional block diagram illustrating a main part of a control function and a control system for shift control in the power transmission device of FIG. 1 and a main part of a control function of the electronic control device. In the continuously variable transmission provided in the power transmission device of FIG. 1, a torque shift generation factor that generates a torque shift (upshift) during CtoC shift, and a countermeasure for suppressing the occurrence of torque shift due to the torque shift generation factor FIG. In the continuously variable transmission of FIG. 1, the primary thrust Win determined from each thrust ratio (γmax balance thrust ratio) of the present embodiment and a comparative example for achieving the lowest speed ratio γmax during the CtoC shift is described. is there. A predetermined thrust ratio map of the torque ratio and the thrust ratio τ using the target speed ratio γtgt as a parameter, which is used to calculate the thrust ratio with respect to the target speed ratio of the continuously variable transmission provided in the power transmission device of FIG. It is a figure which shows an example. It is a figure which shows the relationship between the input torque Tin, thrust ratio (tau), and gear ratio (gamma) c of the continuously variable transmission of FIG. 1, Comprising: It is a conceptual diagram explaining a torque shift. It is a flowchart explaining the principal part of the control action of the electronic controller of FIG. It is a figure explaining an example of the control action in the upshift by the CtoC shift of the electronic control unit of FIG. It is a figure explaining an example of the control action in the upshift by the CtoC shift of the electronic control apparatus of the comparative example which has a control function different from the electronic control apparatus of FIG. FIG. 3 is a time chart showing an example of a control operation in an upshift by CtoC shift of the electronic control device of FIG. 3 and an example of a control operation of a comparative electronic control device having a control function different from that of the electronic control device of FIG. It is.

  Hereinafter, an embodiment of a control device for a vehicle power transmission device 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. In FIG. 1, a vehicle 10 includes an engine 12 that functions as a driving force source for traveling, drive wheels 14, and a power transmission device 16 provided between the engine 12 and the drive wheels 14. The power transmission device 16 includes a torque converter 20 connected to the engine 12 in a housing 18 as a non-rotating member, an input shaft 22 provided integrally with a turbine shaft that is an output rotating member of the torque converter 20, an input shaft A belt-type continuously variable transmission 24 (hereinafter referred to as a continuously variable transmission 24) as a continuously variable transmission mechanism connected to 22, a forward / reverse switching device 26, and a forward / reverse switching device 26 also connected to the input shaft 22. A gear mechanism 28 as a transmission mechanism connected to the input shaft 22 and provided in parallel with the continuously variable transmission 24, an output shaft 30 that is a common output rotating member of the continuously variable transmission 24 and the gear mechanism 28, and a counter shaft 32. Further, a reduction gear device 34 comprising a pair of gears which are provided in mesh with each other so as not to rotate relative to the output shaft 30 and the counter shaft 32, respectively, and a gear provided to the counter shaft 32 so as to be relatively non-rotatable. Differential gear 38 connected to 36, and includes an axle 40 or the like of the pair coupled to a differential gear 38. In the power transmission device 16 configured as described above, the power of the engine 12 (the torque and the force are synonymous unless otherwise specified) is transmitted to the torque converter 20, the continuously variable transmission 24 (or the forward / reverse switching device 26 and the gear mechanism 28). ), The reduction gear device 34, the differential gear 38, the axle 40, and the like are sequentially transmitted to the pair of drive wheels 14.

  The power transmission device 16 includes a first power transmission path that transmits the power of the engine 12 from the input shaft 22 that is an input rotation member to the drive wheel 14 side (that is, the output shaft 30) through the continuously variable transmission 24, and the engine 12. A second power transmission path for transmitting power from the input shaft 22 through the gear mechanism 28 to the drive wheel 14 side (that is, the output shaft 30) is provided between the input shaft 22 and the output shaft 30 in parallel. The first power transmission path and the second power transmission path are switched in accordance with the traveling state. Therefore, the power transmission device 16 is a CVT that intermittently transmits power in the first power transmission path as a clutch mechanism that selectively (selectively) switches between the first power transmission path and the second power transmission path. A travel clutch C2 and a forward clutch C1 and a reverse brake B1 for interrupting power transmission in the second power transmission path are provided. The CVT travel clutch C2, the forward clutch C1, and the reverse brake B1 are all hydraulic friction engagement devices (friction clutches) that are frictionally engaged by a hydraulic actuator. The continuously variable transmission 24 corresponds to the belt-type continuously variable transmission mechanism of the present invention, and the forward / reverse switching device 26 and the gear mechanism 28 correspond to the gear-type transmission mechanism of the present invention. The CVT travel clutch C2 corresponds to the first clutch of the present invention, and the forward clutch C1 corresponds to the second clutch of the present invention.

  The torque converter 20 is provided around the input shaft 22 coaxially with the input shaft 22, and includes a pump impeller 20 p connected to the engine 12 and a turbine impeller 20 t connected to the input shaft 22. ing. The pump impeller 20p controls the speed of the continuously variable transmission 24, generates a belt narrow pressure in the continuously variable transmission 24, switches the operation of each clutch mechanism, and transmits power to the power transmission device 16. A mechanical oil pump 68 is connected, which is generated by rotationally driving hydraulic pressure for supplying lubricating oil to each part of the path by the engine 12.

  The forward / reverse switching device 26 is provided coaxially with the input shaft 22 around the input shaft 22 and mainly includes a double pinion planetary gear device 26p, a forward clutch C1, and a reverse brake B1. Has been. The carrier 26c of the planetary gear unit 26p is integrally connected to the input shaft 22, the ring gear 26r of the planetary gear unit 26p is selectively connected to the housing 18 via the reverse brake B1, and the sun gear 26s of the planetary gear unit 26p is A small-diameter gear 42 is provided around the input shaft 22 so as to be rotatable relative to the input shaft 22 coaxially. The carrier 26c and the sun gear 26s are selectively connected via the forward clutch C1.

  The gear mechanism 28 includes a small-diameter gear 42 and a large-diameter gear 46 that is provided on the gear mechanism counter shaft 44 so as not to rotate relative to the small-diameter gear 42. Therefore, the gear mechanism 28 is a transmission mechanism in which one gear stage (speed ratio) is formed. An idler gear 48 is provided around the gear mechanism counter shaft 44 so as to be rotatable relative to the gear mechanism counter shaft 44 coaxially. Around the gear mechanism counter shaft 44, a meshing clutch D1 is provided between the gear mechanism counter shaft 44 and the idler gear 48 to selectively connect and disconnect between them. The meshing clutch D1 can be engaged with (engaged with) the first gear 50 formed on the gear mechanism counter shaft 44, the second gear 52 formed on the idler gear 48, and the first gear 50 and the second gear 52. And a hub sleeve 54 formed with inner teeth that can be meshed. In the meshing clutch D <b> 1 configured as described above, the gear mechanism counter shaft 44 and the idler gear 48 are connected by the hub sleeve 54 being engaged with the first gear 50 and the second gear 52. The meshing clutch D1 further includes a known synchromesh mechanism S1 as a synchronizing mechanism that synchronizes rotation when the first gear 50 and the second gear 52 are engaged. The idler gear 48 meshes with an output gear 56 having a larger diameter than the idler gear 48. The output gear 56 is provided around the same rotational axis as the output shaft 30 so as not to rotate relative to the output shaft 30. When one of the forward clutch C1 and the reverse brake B1 is engaged and the meshing clutch D1 is engaged, the power of the engine 12 in the first power transmission path and the second power transmission path is input to the input shaft 22. From the forward / reverse switching device 26, the gear mechanism 28, the idler gear 48, and the output gear 56, the second power transmission path that is transmitted to the output shaft 30 is selected.

  The continuously variable transmission 24 is provided on a power transmission path between the input shaft 22 and the output shaft 30. The continuously variable transmission 24 includes an input side member provided on the input shaft 22 and a primary pulley 58 that is an input side pulley having a variable effective diameter, and an output side provided on a secondary shaft 60 coaxial with the output shaft 30. A secondary pulley 62 that is an output-side pulley having a variable effective diameter, which is a member, and a transmission belt 64 wound between the pair of variable pulleys 58 and 62, and the pair of variable pulleys 58 and 62 and the transmission belt Power is transmitted via a frictional force between 64 and 64. The output shaft 30 is arranged around the secondary shaft 60 so as to be rotatable relative to the secondary shaft 60 coaxially.

  In the primary pulley 58, the hydraulic pressure acting on the primary pulley 58 (that is, the primary pressure Pin supplied to the primary hydraulic cylinder 58c provided in the primary pulley 58) is pressure-controlled by the hydraulic control circuit 66 (see FIG. 3). Thus, the input side thrust (primary thrust) Win (= primary pressure Pin × pressure receiving area) in the primary pulley 58 for changing the V groove width between the sheaves 58a and 58b is controlled. In the secondary pulley 62, the hydraulic pressure acting on the secondary pulley 62 (that is, the secondary pressure Pout supplied to the secondary hydraulic cylinder 62 c provided in the secondary pulley 62) is regulated by the hydraulic control circuit 66. The output side thrust (secondary thrust) Wout (= secondary pressure Pout × pressure receiving area) in the secondary pulley 62 for changing the V groove width between the sheaves 62a and 62b is controlled. By controlling the primary thrust Win and the secondary thrust Wout, the width of the V-groove of each pulley 58, 62 is changed, the engagement diameter (effective diameter) of the transmission belt 64 is changed, and the gear ratio (gear ratio) γ ( = Input shaft rotational speed Nin / Output shaft rotational speed Nout) is continuously changed, and frictional force (belt clamping) between the pulleys 58 and 62 and the transmission belt 64 is prevented so that the transmission belt 64 does not slip. Pressure) is necessary and well controlled. In this way, by controlling the primary thrust Win and the secondary thrust Wout, the actual transmission ratio (actual transmission ratio) γ is set to the target transmission ratio γtgt while preventing the transmission belt 64 from slipping. The primary hydraulic cylinder 58c corresponds to the input hydraulic cylinder of the present invention.

  The CVT travel clutch C2 is a secondary pulley 62 (secondary shaft 60) that is an output side member (output pulley) of the continuously variable transmission 24 (here, the downstream side in the torque transmission path from the secondary pulley 62 (secondary shaft 60)). The output shaft 30 on the drive wheel 14 side also agrees), and connects and disconnects the secondary pulley 62 and the output shaft 30. In other words, the CVT travel clutch C <b> 2 is provided between the secondary pulley 62 and the output shaft 30. Of the first power transmission path and the second power transmission path, the first power transmission path in which the power of the engine 12 is transmitted from the input shaft 22 to the output shaft 30 via the continuously variable transmission 24 is selected. Sometimes, the CVT travel clutch C2 is engaged.

  FIG. 2 is a diagram for explaining the switching of the travel pattern using the engagement table of the engagement elements for each travel pattern of the power transmission device 16. In FIG. 2, C1 corresponds to the operating state of the forward clutch C1, C2 corresponds to the operating state of the CVT traveling clutch C2, B1 corresponds to the operating state of the reverse brake B1, and D1 is the meshing clutch D1. "○" indicates engagement (connection), and "x" indicates release (cutoff).

  For example, from gear traveling (also referred to as a gear mode), which is a traveling pattern in which the power of the engine 12 is transmitted to the output shaft 30 via the gear mechanism 28 (that is, a traveling pattern in which power is transmitted through the second power transmission path), CVT traveling (also referred to as a belt mode) that is a traveling pattern in which the power of the engine 12 is transmitted to the output shaft 30 via the continuously variable transmission 24 (that is, a traveling pattern in which the power is transmitted through the first power transmission path). When switching to (high vehicle speed), a CtoC shift is performed in which the forward clutch C1 is released and the clutch is switched so that the CVT travel clutch C2 is engaged, and the power transmission device 16 is upshifted.

  Further, for example, when switching from CVT travel (high vehicle speed) to gear travel, the gear change is performed such that the CVT travel clutch C2 is released from the CVT travel (medium vehicle speed) state and the forward clutch C1 is engaged. (For example, CtoC shift) is executed, and the power transmission device 16 is downshifted. Therefore, the first power transmission path and the second power transmission path are alternatively selected by executing the CtoC shift in which the CVT travel clutch C2 and the forward clutch C1 are switched from one to the other.

  FIG. 3 is a diagram for explaining a control function for shift control in the power transmission device 16 and a main part of the control system. In FIG. 3, the vehicle 10 functions as a control device of the vehicle power transmission device of the present invention, for example, for switching the traveling pattern of the power transmission device 16 or controlling the continuously variable transmission of the continuously variable transmission 24. Is included. Therefore, FIG. 3 is a diagram illustrating an input / output system of the electronic control device 100, and is also a functional block diagram for explaining a main part of a control function by the electronic control device 100. The electronic control device 100 includes, for example, a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like, and the CPU uses a temporary storage function of the RAM according to a program stored in the ROM in advance. Various controls of the vehicle 10 are executed by performing signal processing. For example, the electronic control unit 100 executes output control of the engine 12, shift control of the continuously variable transmission 24, belt clamping pressure control, control for switching a running pattern, and the like. In addition, it is configured separately for shift control and the like.

  The electronic control device 100 is based on detection signals from various sensors provided in the vehicle 10 (for example, various rotational speed sensors 82, 84, 86, accelerator opening sensor 88, throttle valve opening sensor 90, secondary rotational speed sensor 92, etc.). Corresponding to various actual values (for example, the engine speed Ne (rpm), the input shaft rotational speed Nin (rpm), which is the rotational speed of the primary pulley 58 corresponding to the turbine rotational speed Nt (rpm), and the vehicle speed V (km / h). The output shaft rotation speed Nout (rpm), the accelerator pedal operation amount θacc (%) as the driver's acceleration request amount, the throttle valve opening θth (%), and the secondary pulley 62 rotation speed Rotational speed Nsec, etc.) are respectively supplied. The electronic control unit 100 calculates the gear ratio γc (= input shaft rotational speed Nin / secondary rotational speed Nsec) of the continuously variable transmission 24 based on, for example, the input shaft rotational speed Nin and the secondary rotational speed Nsec.

  The electronic control unit 100 also outputs an engine output control command signal Se for output control of the engine 12, hydraulic control command signals Sin and Sout for hydraulic control related to the shift of the continuously variable transmission 24, and the power transmission device 16. Hydraulic control command signals Sc1, Sb1, Sc2, Sd1, etc. for controlling the forward / reverse switching device 26, the CVT traveling clutch C2, and the meshing clutch D1 related to the switching of the traveling pattern are output. Specifically, as an engine output control command signal Se, a throttle signal for driving the throttle actuator to control the opening and closing of the electronic throttle valve, an injection signal for controlling the amount of fuel injected from the fuel injection device, An ignition timing signal or the like for controlling the ignition timing of the engine 12 by the ignition device is output. Further, as a hydraulic control command signal Sin, a command signal for driving a solenoid valve for regulating the primary pressure Pin supplied to the actuator of the primary pulley 58 and a hydraulic control command signal Sout are supplied to the actuator of the secondary pulley 62. A command signal or the like for driving a solenoid valve that regulates the secondary pressure Pout is output to the hydraulic control circuit 66. Further, as the hydraulic control command signals Sc1, Sb1, Sc2, Sd1, hydraulic pressures acting on the forward clutch C1, the reverse brake B1, the CVT travel clutch C2, and the meshing clutch D1, respectively (that is, the forward clutch C1, the reverse clutch). Command signals for driving the solenoid valves for regulating the clutch pressure Pc1, the clutch pressure Pb1, the clutch pressure Pc2, and the clutch pressure Pd1) supplied to the actuators of the brake B1, the CVT travel clutch C2, and the meshing clutch D1. Is output to the hydraulic control circuit 66.

  In the hydraulic control circuit 66, the line pressure P1 is regulated by a solenoid valve using, for example, the operating hydraulic pressure output (generated) from the oil pump 68 as a source pressure. In the hydraulic control circuit 66, for example, the primary pressure Pin and the secondary pressure Pout are controlled so as to generate a belt narrow pressure in the pulleys 58 and 62 that does not cause belt slippage and does not increase unnecessarily. Further, the gear ratio γc of the continuously variable transmission 24 is changed by changing the thrust ratio τ (= Wout / Win) of the pulleys 58 and 62 due to the mutual relationship between the primary pressure Pin and the secondary pressure Pout. For example, the gear ratio γc is increased as the thrust ratio τ is increased (that is, the continuously variable transmission 24 is downshifted).

  The electronic control device 100 includes a shift control unit 102. The shift control unit 102 includes a CtoC shift determination unit 104, a belt input torque calculation unit 106, a thrust control unit 110, and a CtoC shift control unit 112.

  The CtoC shift determining unit 104 switches the upshift line and the downshift for switching between the first speed ratio γ1 corresponding to the speed ratio EL in gear traveling and the second speed ratio γ2 corresponding to the lowest speed ratio γmax in CVT traveling. Using the shift line, a shift (switching of the gear ratio) is determined based on the vehicle speed V and the accelerator opening degree θacc, and based on the determination result, it is determined whether or not to switch the travel pattern during vehicle travel. The upshift line and the downshift line are, for example, predetermined shift lines and have a predetermined hysteresis. Note that the lowest speed ratio γmax in CVT traveling corresponds to the maximum speed ratio formed by the belt-type continuously variable transmission of the present invention.

  When the CtoC shift determination unit 104 determines an upshift during gear traveling, the CtoC shift control unit 112 releases the forward clutch C1 and executes an upshift by a CtoC shift that engages the CVT travel clutch C2. Thus, the gear mode is switched to the CVT travel mode (high vehicle speed). Further, from the viewpoint of continuity of the change in the gear ratio γ accompanying the switching from the gear mode to the CVT travel mode, the gear ratio γc of the continuously variable transmission 24 is, for example, the lowest during the upshift by the CtoC gearshift by the thrust control unit 110 Control is performed so that the speed ratio γmax is maintained.

  The CtoC shift control unit 112 releases the CVT travel clutch C2 and engages the forward clutch C1 when the CtoC shift determination unit 104 determines a downshift during CVT travel (high vehicle speed). By executing the downshift, the CVT traveling mode is switched to the gear traveling mode. The downshift by the CtoC shift may be performed simultaneously with the downshift of the continuously variable transmission 24, for example, when a large acceleration operation is performed by the driver.

  During the CtoC shift, the input torque Tin to the continuously variable transmission 24 changes according to the change in the torque capacity Tc2 of the CVT travel clutch C2. Therefore, basically, the torque capacity Tc2 of the CVT travel clutch C2 is reduced. By dividing by the actual speed ratio γc of the continuously variable transmission 24, the input torque Tin to the continuously variable transmission 24 during the CtoC shift is calculated. Hereinafter, the input torque Tin calculated in this way is referred to as an estimated input torque Tines. Here, the torque capacity Tc2 of the CVT travel clutch C2 during the CtoC shift is calculated based on the clutch command pressure Pc2dir output to the hydraulic control circuit 66 and the specifications of the secondary hydraulic cylinder 62c.

  FIG. 4 illustrates a torque shift generation factor that causes a torque shift (upshift) of the continuously variable transmission 24 during CtoC shift, and a torque shift occurrence in the continuously variable transmission 24 due to the torque shift generation factor. It is a figure which shows this correspondence. Here, the torque shift maintains the target gear ratio γtgt due to a torque shift element such as a deviation (variation) between the estimated input torque Tines calculated based on the clutch command pressure Pc2dir and the actual input torque Tin, for example. It is not possible to upshift or downshift from the target gear ratio γtgt.

  As shown in FIG. 4, the torque shift generation factor (torque shift element) corresponding to the horizontal axis (torque) of FIG. 7 described later has a variation in the torque capacity Tc2 of the CVT travel clutch C2 with respect to the clutch command pressure Pc2dir. Are listed. The torque capacity Tc2 of the CVT travel clutch C2 used for calculating the estimated input torque Tines during the CtoC shift is calculated based on the specifications of the clutch command pressure Pc2dir and the secondary hydraulic cylinder 62c. Therefore, for example, the clutch command pressure Pc2dir is caused by the difference between the clutch command pressure Pc2dir and the actual clutch pressure Pc2, the friction coefficient μ of the friction plate of the CVT travel clutch C2, the specifications of the secondary hydraulic cylinder 62c, and the like. The torque capacities Tc2 that are actually obtained with respect to Pc2dir are not uniform and may vary within a predetermined range. Since the estimated input torque Tines is obtained from the calculated value of the torque capacity Tc2 having variations, there is a possibility that a deviation occurs from the actual input torque Tin.

  If the actual input torque Tin is lower than the estimated input torque Tines during the CtoC shift, an upshift may occur from the lowest speed ratio γmax in terms of thrust ratio characteristics. For this reason, during the CtoC shift, in order to maintain the lowest speed ratio γmax regardless of the actual input torque Tin, the target primary thrust Wintgt is set by setting the thrust ratio τ that achieves the lowest speed ratio γmax to the maximum value. It is conceivable that the primary pressure Pin supplied to the primary hydraulic cylinder 58c to be obtained is the minimum pressure. As a result, the lowest speed ratio γmax of the continuously variable transmission 24 is maintained regardless of the actual input torque Tin. However, the primary pressure Pin when the thrust ratio τ that achieves the lowest speed ratio γmax is set to the maximum value is required from the primary pressure Pin that is required to secure the torque capacity according to the actual input torque Tin. Because of the above reduction, the durability of the continuously variable transmission 24 may be reduced. Therefore, during the CtoC shift, the electronic control unit 100 estimates the input torque Tin of the continuously variable transmission 24 during the CtoC shift based on the torque capacity Tc2 of the CVT travel clutch C2, and the estimated continuously variable transmission. Based on the estimated input torque Tines that is the input torque Tin of 24, the primary pressure Pin of the primary hydraulic cylinder 58c provided in the primary pulley 58 is set so that the upshift of the continuously variable transmission 24 does not occur during the CtoC shift. Control. That is, in the present embodiment, the primary pressure Pin is controlled to be smaller than the primary pressure Pin corresponding to the actual input torque Tin, and the thrust ratio characteristic corresponds to the actual input torque Tin at the lowest gear ratio γmax. The torque shift becomes the downshift side, and γmax pressing control (FIG. 4) is performed in which the lowest speed ratio γmax is maintained. In the γmax pressing control of the continuously variable transmission 24, since the primary pressure Pin cannot be lowered to the minimum pressure, it is possible to suppress a decrease in durability of the continuously variable transmission 24. The electronic control unit 100 ensures the torque capacity Tc2 of the continuously variable transmission 24 when calculating the primary pressure Pin of the primary hydraulic cylinder 58c during the CtoC shift in the γmax pressing control of the continuously variable transmission 24 during the CtoC shift. The estimated input torque Tlmtc for calculating the torque capacity and the estimated input torque Tinc for calculating the thrust ratio τ for calculating the thrust ratio τ of the continuously variable transmission 24 are used, but the input of the continuously variable transmission 24 during the CtoC shift is used. As the torque Tin, the following two different estimated input torque Tines are used. Here, the estimated input torque Tlmtc for calculating the torque capacity during the CtoC shift is used to calculate the belt slip guarantee thrust Wlmtc that is a necessary and sufficient thrust for guaranteeing the suppression of the belt slip during the CtoC shift. This is the input torque Tin of the step transmission 24. The estimated input torque Tinc for thrust ratio calculation is the input torque Tin of the continuously variable transmission 24 used to calculate the thrust ratio τ of the continuously variable transmission 24 during the CtoC shift.

  When the CtoC shift determination unit 104 determines that the CtoC shift is being performed, the belt input torque calculation unit 106 calculates the estimated input torque Tines from the specifications of the clutch command pressure Pc2dir and the secondary hydraulic cylinder 62c. Further, from the predetermined relationship between the clutch command pressure Pc2dir and the torque capacity Tc2 of the CVT travel clutch C2 and its variation, the variation of the torque capacity Tc2 with respect to the clutch command pressure Pc2dir is estimated based on the actual clutch command pressure Pc2dir. The maximum value Tc2max of the variation in the torque capacity Tc2 and the minimum value Tc2min of the variation in the torque capacity Tc2 are estimated. Here, the predetermined variation range of the torque capacity Tc2 obtained by subtracting the minimum value Tc2min of the variation in the torque capacity Tc2 from the maximum value Tc2max of the variation in the torque capacity Tc2 is a substantially constant value, and the clutch command pressure Pc2dir and the actual clutch pressure It is experimentally set in advance from the deviation width from Pc2, the friction coefficient μ of the friction plate of the CVT travel clutch C2, the specifications of the secondary hydraulic cylinder 62c, and the like. The belt input torque calculator 106 is the smaller value of the estimated input torque variation maximum value Tinesmax obtained by dividing the maximum variation Tc2max of the turbine torque Tt and the torque capacity Tc2 by the transmission ratio γc of the continuously variable transmission 24. (Minimum select value) is set to the estimated input torque Tlmtc for torque capacity calculation during the CtoC shift. That is, the belt input torque calculation unit 106 determines that the maximum value Tc2max of the variation in the torque capacity Tc2 with respect to the clutch command pressure Pc2dir to the CVT travel clutch C2 when the maximum variation value Tinesmax of the estimated input torque is smaller than the turbine torque Tt. Based on the above, the estimated input torque Tinesmax for torque capacity calculation is set.

  The belt input torque calculation unit 106 sets the thrust ratio calculation estimated input torque Tinc based on the minimum variation Tc2min of the torque capacity Tc2. Specifically, the estimated input torque variation minimum value Tinesmin obtained by dividing the minimum variation value Tc2min of the torque capacity Tc2 by the speed ratio γc of the continuously variable transmission 24 is set as the estimated input torque Tinc for thrust ratio calculation. Alternatively, when the belt input torque calculation unit 106 sets the estimated input torque variation maximum value Tinesmax to the torque capacity calculation estimated input torque Tlmtc by the minimum selection of the turbine torque Tt and the estimated input torque variation maximum value Tinesmax. Then, the estimated input torque variation minimum value Tinesmin is set to the thrust ratio calculation estimated input torque Tinc, and the turbine torque Tt is estimated by the minimum selection of the turbine torque Tt and the estimated input torque variation maximum value Tinesmax. In the case of setting to Tlmtc, a value obtained by subtracting a division value obtained by dividing the predetermined variation range of the torque capacity Tc2 by the speed ratio γc from the turbine torque Tt may be set as the estimated input torque Tinc for thrust ratio calculation. That is, the belt input torque calculation unit 106 subtracts a value obtained by subtracting a value obtained by dividing the predetermined variation range of the torque capacity Tc2 by the speed ratio γc from the estimated input torque Tlmtc for torque capacity calculation. May be set. This suppresses the actual input torque Tin from being lower than the estimated input torque Tinc for thrust ratio calculation during the CtoC shift.

  The thrust control unit 110 calculates a primary pressure Pin supplied to the primary hydraulic cylinder 58c of the primary pulley 58 for obtaining the primary thrust Win during the CtoC shift. FIG. 5 is a diagram for explaining the primary thrust Win determined from each thrust ratio (γmax balance thrust ratio) of the present embodiment and the comparative example for achieving the lowest speed ratio γmax during the CtoC shift. When calculating the primary pressure Pin, the thrust control unit 110 uses a belt slip guarantee thrust that is necessary and sufficient for securing the torque capacity of the continuously variable transmission 24 as the input torque Tin of the continuously variable transmission 24. In order to calculate Wlmt, the torque capacity calculation estimated input torque Tlmtc set to the smaller value of the turbine torque Tt and the estimated input torque variation maximum value Tinesmax is used, and the thrust ratio τ of the continuously variable transmission 24 is calculated. Is set to a value obtained by subtracting a value obtained by dividing the predetermined variation range of the torque capacity Tc2 by the speed ratio γc of the continuously variable transmission 24 from the minimum value Tinesmin of the estimated input torque or the turbine torque Tt. The estimated input torque Tinc for thrust ratio calculation is used.

The thrust control unit 110 calculates the primary side belt slip guarantee thrust Wlmtin and the secondary side belt slip guarantee thrust Wlmtout during the CtoC shift using the torque capacity calculation estimated input torque Tlmtc. The thrust control unit 110 calculates an estimated input torque Tlmtc for torque capacity calculation set by the belt input torque calculation unit 106 as an input torque of the primary pulley 58 during the CtoC shift from the following equations (1) and (2), respectively. , The output torque Tout (= γ × Tlmtc) of the continuously variable transmission 24 as the input torque of the secondary pulley 62, the sheave angle (cone surface angle) α of each pulley 58, 62, a predetermined belt element-sheave friction coefficient μ The belt engagement diameter Rin on the primary pulley 58 side calculated uniquely from the actual transmission ratio γ, and the belt engagement diameter Rout on the secondary pulley 62 side calculated uniquely from the actual transmission ratio γ (refer to FIG. 1 above). Based on this, the primary side belt slip guarantee thrust Wlmtin and the secondary side belt slip guarantee thrust Wlmtout during the CtoC shift are calculated. The following equation (2) corresponds to the first relationship of the present invention.
Wlmtin = (Tlmtc × cosα) / (2 × μ × Rin) (1)
Wlmtout = (Tout × cosα) / (2 × μ × Rout) (2)

  FIG. 6 is a diagram illustrating an example of a predetermined thrust ratio map of the torque ratio and the thrust ratio τ using the target speed ratio γtgt as a parameter. The thrust control unit 110 calculates the thrust ratio τ of the continuously variable transmission 24 during the CtoC shift using the thrust ratio calculation estimated input torque Tinc from the thrust ratio map shown in FIG. The thrust ratio map uses the target speed ratio γtgt of the continuously variable transmission 24 for calculating the thrust ratio τ of the continuously variable transmission 24 as a parameter and the torque ratio obtained from the input torque Tin of the continuously variable transmission 24. This is a predetermined relationship with the thrust ratio τ of the step transmission 24. The thrust control unit 110 calculates a target input shaft rotation speed Nitgt based on the accelerator opening θacc and the vehicle speed V from a predetermined relationship (for example, CVT shift map), and the target shift based on the target input shaft rotation speed Nitgt. Although the ratio γtgt (= Nitgt / Nout) is calculated, the target speed ratio γtgt is the lowest speed ratio γmax during the CtoC shift. In addition, the thrust control unit 110 calculates a torque ratio used for calculating the thrust ratio τ. The torque ratio during CtoC shift is the estimated input torque Tinc for thrust ratio calculation set by the belt input torque calculation unit 106 and the guaranteed input torque Tlmtin that is the limit torque that can be input to the continuously variable transmission 24. (= Tinc / Tlmtin). The thrust control unit 110 determines the lowest speed ratio γmax, which is the target speed ratio γtgt, based on the target speed ratio γtgt and the torque ratio during the CtoC speed change based on a predetermined relationship (thrust ratio map) as shown in FIG. The thrust ratio τ (γmax balance thrust ratio) for maintaining the above constant is calculated. In the thrust ratio map, the torque ratio is small in a driving state where the driving wheel 14 is driven by the driving force of the engine 12 having a torque ratio of 0 to 1 except in a region where the torque ratio is close to 1 and the torque ratio is relatively high. In other words, as the estimated input torque Tinc for thrust ratio calculation as the input torque of the continuously variable transmission 24 during CtoC shift becomes smaller, the thrust ratio τ (γmax balance thrust ratio) that achieves the lowest speed ratio γmax increases. Have. The thrust ratio map corresponds to the second relationship of the present invention.

  As shown in FIG. 5, the thrust ratio τ (γmax balance thrust ratio) during the CtoC shift in this embodiment is the CVT running clutch with respect to the clutch command pressure Pc2dir that is taken into account for setting the torque capacity calculation estimated input torque Tlmtc. The maximum value Tc2max of the variation in the torque capacity Tc2 of C2 is set in consideration of the minimum value Tc2min of the variation in the torque capacity Tc2 with respect to the clutch command pressure Pc2dir, where the difference (variation) in the torque capacity Tc2 is the maximum (MAX). Is calculated based on the estimated input torque Tinc for thrust ratio calculation. On the other hand, the thrust ratio calculation estimated input torque Tinc for calculating the thrust ratio τ (γmax balance thrust ratio) for achieving the lowest speed ratio γmax during the CtoC shift in the comparative example of FIG. 5 is the torque capacity calculation estimated input. Similar to the torque Tlmtc, the maximum value Tinesmax of the estimated input torque obtained by dividing the maximum value Tc2max of the torque capacity Tc2 by the speed ratio γc and the smaller value of the turbine torque Tt. That is, the thrust ratio (γmax balance thrust ratio) during the CtoC shift in the comparative example is relative to the maximum value Tc2max of the variation in the torque capacity Tc2 with respect to the clutch command pressure Pc2dir that is taken into account for setting the torque capacity calculation estimated input torque Tlmtc. The torque capacity Tc2 is calculated based on the thrust ratio calculating estimated input torque Tinc set to the same value as the torque capacity calculating estimated input torque Tlmtc without considering the variation of the torque capacity Tc2 (C2 torque capacity nominal). For this reason, the thrust ratio τ (γmax balance thrust ratio) during the CtoC shift of this embodiment is larger than the thrust ratio τ (γmax balance thrust ratio) during the CtoC shift of the comparative example.

  In FIG. 5, the thrust control unit 110 obtains the primary side slip guarantee thrust Wlmtin calculated based on the torque capacity calculation estimated input torque Tlmtc from the equation (1) and the thrust ratio τ (γmax balance) that achieves the lowest gear ratio γmax. The secondary side shift control thrust Woutsh is calculated from the thrust ratio). The thrust control unit 110 is larger of the secondary-side slip guarantee thrust Wlmtout calculated based on the torque capacity calculation estimated input torque Tlmtc from the equation (2) and the secondary-side shift control thrust Woutsh calculated as described above. Is set as the target secondary thrust Wouttgt (secondary thrust indicated by a solid bar graph in FIG. 5). The thrust control unit 110 divides the target secondary thrust Wouttgt by the thrust ratio τ (γmax balance thrust ratio) to calculate a target primary thrust Wintgt (primary thrust indicated by a solid bar graph in FIG. 5). The thrust control unit 110 converts the target secondary thrust Wouttgt and the target primary thrust Wintgt into a target secondary pressure Pouttgt (= Wouttgt / 62c pressure receiving area) and a target primary pressure Pintgt (=) based on the pressure receiving areas of the hydraulic cylinders 62c and 58c. Each pressure is converted into a pressure receiving area of Wintgt / 58c). The thrust control unit 110 obtains the primary command pressure Pindir as the hydraulic control command signal Sin, the secondary command pressure Poutdir as the hydraulic control command signal Sout, and the hydraulic control circuit 66 so that the target primary pressure Pintgt and the target secondary pressure Pouttgt are obtained. Output to. The hydraulic control circuit 66 adjusts the primary pressure Pin and the secondary pressure Pout by operating each solenoid valve according to the hydraulic control command signals Sin and Sout.

  The target primary thrust Wintgt and target secondary thrust Wouttgt in the comparative example are the maximum value Tc2max of the variation in the torque capacity Tc2 with respect to the turbine torque Tt and the clutch command pressure Pc2dir based on the formulas (1), (2) and the thrust ratio map. It is calculated based on the estimated input torque Tlmtc for torque capacity calculation and the estimated input torque Tinc for thrust ratio calculation set to the smaller one of the maximum variations Tinesmax of the estimated input torque divided by the speed ratio γc. In FIG. 5, the target primary thrust Wintgt of the comparative example is a thrust corresponding to the predetermined variation range of the torque capacity Tc2 with respect to the clutch command pressure Pc2dir indicated by a broken line rather than the target primary thrust Wintgt of the embodiment indicated by the solid bar graph. Bigger than that. That is, the thrust ratio τ (γmax balance thrust ratio) during the CtoC shift of the present embodiment is larger than the thrust ratio τ (γmax balance thrust ratio) during the CtoC shift of the comparative example, and thus the target primary thrust in this embodiment. Wintgt is set to the primary thrust Win that is subtracted (lowered) by the thrust corresponding to the predetermined variation range of the torque capacity Tc2 from the target primary thrust Wintgt in the comparative example. The primary pressure for obtaining the target primary thrust Wintgt in the comparative example of FIG. 5 of the primary pressure Pin supplied to the primary hydraulic cylinder 58c for obtaining the target primary thrust Win in the present embodiment set as described above. The amount of decrease from Pin is, for example, the amount of decrease from the primary pressure Pin in the comparative example of FIG. 5 of the primary pressure Pin set to the minimum pressure with the thrust ratio τ that achieves the lowest gear ratio γmax being maximized. Is also small.

  FIG. 7 is a diagram showing a relationship among the input torque Tin, the thrust ratio τ, and the speed ratio γc of the continuously variable transmission 24, and is a conceptual diagram for explaining the torque shift. In FIG. 7, a series of points indicating the same gear ratio γc is illustrated by a solid line. The white circle position in FIG. 7 is calculated based on the target gear ratio γtgt at the white circle position and the estimated input torque Tines that is the input torque Tin (control recognition value) of the control continuously variable transmission 24 indicated by the chain line. The target speed ratio γtgt is maintained by the balance thrust obtained from the thrust ratio τ. At this time, if the actual input torque Tin (actual value) of the continuously variable transmission 24 indicated by the one-dot chain line is lower than the estimated input torque Tines (control recognition value), the thrust ratio τ is calculated from the white circle position. There is a possibility that the speed ratio γc at the higher vehicle speed side than the speed ratio γc at the white circle position corresponding to the actual (actual) input torque Tin moved in the arrow direction toward the low torque side while being constant may be upshifted. On the contrary, when the actual input torque Tin (actual value) of the continuously variable transmission 24 is higher than the estimated input torque Tines (control recognition value), the high torque is maintained with the thrust ratio τ kept constant from the white circle position. There is a possibility of downshifting to a gear ratio γc on the lower vehicle speed side than the gear ratio γc at the white circle position corresponding to the actual (actual) input torque Tin moved in the arrow direction.

  In this embodiment, the thrust ratio calculation estimated input torque Tinc for calculating the thrust ratio τ (γmax balance thrust ratio) for maintaining the lowest speed ratio γmax during the CtoC shift is the torque capacity with respect to the clutch command pressure Pc2dir. The estimated input torque variation minimum value Tinesmin obtained by dividing the Tc2 variation minimum value Tc2min by the transmission gear ratio γc, or the gradual value obtained by dividing the predetermined variation range of the torque capacity Tc2 by the transmission gear ratio γc from the turbine torque Tt. Set to a value. For this reason, since the actual input torque Tin is suppressed to be lower than the estimated input torque Tinc for thrust ratio calculation during the CtoC shift, the upshift from the lowest speed ratio γmax of the continuously variable transmission 24 is prevented. It is suppressed. On the other hand, the estimated input torque Tinc for thrust ratio calculation of the comparative example is set in consideration of the maximum variation Tc2max of the torque capacity Tc2, so that the actual input torque Tin is lower than the estimated input torque Tinc for thrust ratio calculation. There is a possibility that an upshift from the lowest speed ratio γmax of the continuously variable transmission 24 occurs during the CtoC shift. That is, the primary pressure Pin supplied to the primary hydraulic cylinder 58c of the present embodiment is shown in FIG. 5 in order to suppress an upshift from the lowest gear ratio γmax during the CtoC shift due to the variation in the torque capacity Tc2. The required minimum pressure is lowered from the primary pressure Pin of the comparative example. As a result, in the γmax pressing control of this embodiment, the upshift from the lowest gear ratio γmax during the CtoC shift due to the variation in the torque capacity Tc2 with respect to the clutch command pressure Pc2dir is suppressed, and the durability of the continuously variable transmission 24 is improved. An optimal target primary pressure Pintgt for setting the target primary thrust Wintgt that can achieve both reduction suppression is set.

  As shown in FIG. 4, during the CtoC shift, there is a deviation in the thrust ratio (vertical axis) of FIG. 7 (the actual thrust ratio relative to the indicated thrust ratio) due to variations in the actual primary pressure Pin with respect to the primary indicated pressure Pindir. There is a possibility that a torque shift (upshift) from the lowest gear ratio γmax occurs. For this reason, for example, an actual secondary pressure Pout from the hydraulic sensor, for example, from a predetermined map of the primary pressure Pin and the amount of decrease in the primary pressure Pin that allows for hydraulic pressure variation to maintain the lowest speed ratio γmax. Based on the primary pressure Pin calculated from the thrust ratio τ, a reduction amount of the primary pressure Pin during the CtoC shift is set (Pin variation downshift guarantee control). The γmax pressing control of the present embodiment may be performed in parallel with the Pin variation downshift guarantee control. In this case, the amount of decrease in the primary pressure Pin during the CtoC shift is obtained by adding the amount of decrease in the γmax pressing control and the amount of decrease in the Pin variation downshift guarantee control.

  FIG. 8 is a flowchart for explaining a main part of the control operation of the electronic control device 100. FIG. 9 is a diagram for explaining an example of the control operation of the electronic control unit 100 of the present embodiment in the upshift by the CtoC shift. FIG. 10 is a diagram for explaining an example of a control operation of a comparative electronic control device having a control function different from that of the electronic control device 100 in an upshift by CtoC shift. FIG. 11 shows an example of the control operation of the electronic control device 100 of the present embodiment and an example of the control operation of the comparative electronic control device having a control function different from that of the electronic control device 100 in the upshift by the CtoC shift. These are time charts shown respectively.

  In FIG. 8, in a step (hereinafter, “step” is omitted) S1 corresponding to the function of the CtoC shift determination unit 104, it is determined whether or not the CtoC shift is being performed. If the determination in S1 is negative, the flowchart is terminated. If the determination in S1 is affirmative (from time t1 to time t3 in the present embodiment in FIG. 9 and from time t1 to time t3 in FIG. 11), in S2 corresponding to the function of the belt input torque calculation unit 106, the clutch instruction The maximum variation Tinesmax of the estimated input torque obtained by dividing the maximum value Tc2max of the torque capacity Tc2 of the CVT travel clutch C2 with respect to the pressure Pc2dir (C2 command hydraulic pressure) by the speed ratio γc of the continuously variable transmission 24 and the turbine torque Tt A small value is set as the estimated input torque Tlmtc for torque capacity calculation. 9 and FIG. 11, from the time t1 to the time t2, the estimated input torque variation maximum corresponding to the belt input torque indicated by the broken line in FIG. 9 and the maximum C2 torque capacity variation / transmission ratio indicated by the solid line in FIG. The value Tinesmax is set to the torque capacity calculation estimated input torque Tlmtc (narrow pressure calculation estimated input torque), and from the time t2 to the time t3 in FIG. 9 and FIG. The torque Tt is set to the estimated input torque Tlmtc for torque capacity calculation. In FIG. 11, for the sake of clarity, the estimated input torque Tlmtc for torque capacity calculation indicated by a broken line is slightly shifted from the estimated input torque variation maximum value Tinesmax or turbine torque Tt. In S3 corresponding to the function of the belt input torque calculator 106, the torque is calculated from the minimum value Tinesmin of the estimated input torque obtained by dividing the minimum value Tc2min of the torque capacity Tc2 with respect to the clutch command pressure Pc2dir by the speed ratio γc, or the turbine torque Tt. A value obtained by subtracting the division value obtained by dividing the predetermined variation range of the capacity Tc2 by the speed ratio γc is set as the estimated input torque Tinc for thrust ratio calculation. From time t1 to time t3 in FIG. 9, the estimated input torque variation minimum value Tinesmin corresponding to the belt input torque indicated by the two-dot chain line is set as the thrust ratio calculation estimated input torque Tinc. From time t1 to time t2 in FIG. 11, the estimated input torque variation minimum value Tinesmin is set to the estimated input torque Tinc for thrust ratio calculation indicated by the solid line, and from time t2 to time t3 in FIG. The value obtained by subtracting the value obtained by dividing the predetermined variation range of the capacity Tc2 (the difference between the maximum value Tc2max of the variation in the torque capacity Tc2 and the minimum value Tc2min of the variation with respect to the clutch command pressure Pc2dir) by the gear ratio γc is calculated. The estimated input torque Tinc is set. For this reason, as shown in FIG. 9, the actual input torque Tin (actual belt input torque) during the CtoC shift is suppressed from being lower than the estimated input torque Tinc for thrust ratio calculation. As a result, the torque shift according to the actual input torque Tin becomes the downshift side at the lowest gear ratio γmax, and the actual gear ratio (actual ratio) is maintained at the lowest gear ratio γmax, which is the target gear ratio (target ratio) γtgt. Is done.

  In S4 corresponding to the function of the thrust control unit 110, the belt slip guarantee thrust Wlmt is calculated based on the torque capacity calculation estimated input torque Tlmtc from the equations (1) and (2). Further, the thrust ratio τ that achieves the lowest gear ratio γmax is calculated from the thrust ratio map shown in FIG. 4 based on the estimated input torque Tinc for thrust ratio calculation. In FIG. 11, the thrust ratio τ (γmax balance thrust ratio (C2 capacity variation lower limit)) that achieves the lowest speed ratio γmax indicated by the solid line is determined from the torque capacity calculation estimated input torque Tlmtc indicated by the broken line. Is calculated based on a thrust ratio calculation estimated input torque Tinc indicated by a solid line set to a value obtained by subtracting a division value obtained by dividing the predetermined variation range of the torque capacity Tc2 by the gear ratio γc. In S5 corresponding to the function of the thrust control unit 110, a target secondary thrust Wouttgt and a target primary thrust Wintgt are set from the belt slip guarantee thrust Wlmtc and the thrust ratio τ. The target primary thrust Wintgt is set to a value obtained by dividing the target secondary thrust Wouttgt by the thrust ratio τ. As shown in FIG. 11, the thrust ratio τ indicated by the solid line in this embodiment is set in consideration of the predetermined variation range of the torque capacity Tc2 with respect to the clutch command pressure Pc2dir. The primary pressure Pin indicated by the solid line can be made higher than when the achieved thrust ratio τ is the maximum value and the primary pressure Pin is the minimum pressure. For this reason, the primary pressure Pin supplied to the primary hydraulic cylinder 58c of the primary pulley 58 is prevented from being upshifted from the lowest speed ratio γmax during the CtoC shift, and the durability of the continuously variable transmission 24 is reduced. An optimal primary pressure can be set for suppression. In S6 corresponding to the function of the thrust control unit 110, thrust control of the continuously variable transmission 24 is performed based on the target secondary thrust Wouttgt and the target primary thrust Wintgt during the CtoC shift. After execution of S6, this flowchart is terminated.

On the other hand, from time t1 to time t2 in FIGS. 10 and 11, the belt input torque indicated by the broken line in FIG. 10 and the estimated input torque corresponding to the maximum C2 torque capacity variation / transmission ratio indicated by the solid line in FIG. Is set as the estimated input torque Tinc for thrust ratio calculation, and the turbine torque Tt indicated by the alternate long and short dash line is set as the estimated input torque Tinc for thrust ratio calculation from time t2 to time t3 in FIGS. Has been. That is, in the comparative example of FIG. 10 and FIG. 11, the same value as the estimated input torque Tlmtc for torque capacity calculation is set as the estimated input torque Tinc for thrust ratio calculation. As shown in FIG. 10, in the comparative example, during the CtoC shift, the actual input torque Tin becomes a value lower than the estimated input torque Tinc for thrust ratio calculation, and a torque shift toward the upshift occurs, resulting in the target shift. The actual gear ratio (actual ratio) γc is smaller than the lowest gear ratio γmax, which is the ratio (target ratio) γtgt. In addition, as shown in FIG. 11, the predetermined variation range of the torque capacity Tc2 is taken into consideration for the thrust ratio (γmax balance thrust ratio) for maintaining the lowest speed ratio γmax indicated by the broken line in the comparative example. Since there is no (C2 torque capacity nominal), it is smaller than the thrust ratio τ shown by the solid line in this embodiment. Therefore, as shown in FIG. 11, the primary pressure Pin (C2 torque capacity nominal) indicated by the broken line in the comparative example is more than the predetermined variation in the torque capacity Tc2 than the primary pressure Pin indicated by the solid line in the embodiment. It is larger by the hydraulic pressure of the thrust corresponding to the range. 9 and 10, the clutch command pressure Pc2dir for quick apply control for supplying predetermined hydraulic oil to the hydraulic cylinder of the CVT travel clutch C2 is indicated by a dotted line during a predetermined time from the time t1 in FIGS. In FIG. 11, the turbine rotational speed Nt (input shaft rotational speed Nin) is indicated by a solid line, and the basic target rotational speed Ninbs of the continuously variable transmission 24 is indicated by a broken line. Further, the secondary safety factor reciprocal number SFout ⁻¹ (= Wlmtout / Wout) is indicated by a dotted line, and the secondary thrust ratio reciprocal number is indicated by a solid line.

  As described above, according to the electronic control unit 100 of the present embodiment, when calculating the primary pressure Pin supplied to the primary hydraulic cylinder 58c for obtaining the target primary thrust Wintgt during the CtoC shift, the continuously variable transmission As the input torque Tin of 24, an estimated input torque Tlmtc for calculating the torque capacity for calculating the belt slip guarantee thrust Wlmtc for securing the torque capacity of the continuously variable transmission 24 and the thrust ratio τ of the continuously variable transmission 24 are obtained. Using the estimated input torque Tinc for thrust ratio calculation for calculation, the estimated input obtained by dividing the maximum value Tc2max of the variation in the torque capacity Tc2 of the CVT travel clutch C2 with respect to the clutch command pressure Pc2dir by the speed ratio γc of the continuously variable transmission 24. The smaller one of the maximum torque variation Tinesmax and the turbine torque Tt is set as the estimated input torque Tlmtc for torque capacity calculation. Torque capacity Tc2 from the minimum value Tinesmin of estimated input torque obtained by dividing the minimum value Tc2min of the torque capacity Tc2 of the CVT travel clutch C2 with respect to the clutch command pressure Pc2dir by the speed ratio γc of the continuously variable transmission 24, or from the turbine torque Tt A value obtained by subtracting a division value obtained by dividing the predetermined variation range by the speed ratio γc of the continuously variable transmission 24 is set as the estimated input torque Tinc for thrust ratio calculation. For this reason, the estimated input torque Tinc for thrust ratio calculation is set in consideration of the minimum value Tc2min of the variation in the torque capacity Tc2 of the CVT travel clutch C2 with respect to the clutch command pressure Pc2dir. Is suppressed from becoming lower than the estimated input torque Tinc for thrust ratio calculation, and the upshift of the continuously variable transmission 24 during the CtoC shift is suppressed. Further, the primary pressure Pin of the primary hydraulic cylinder 58c provided in the primary pulley 58 during the CtoC shift is set based on the maximum value Tc2max of variation in the torque capacity Tc2 of the CVT travel clutch C2 with respect to, for example, the clutch command pressure Pc2dir. The amount of decrease from the primary pressure Pin of the primary hydraulic cylinder 58c provided in the primary pulley 58 of the comparative example in which the thrust ratio τ of the continuously variable transmission 24 is calculated from the estimated input torque Tinc for thrust ratio calculation is CVT running The hydraulic pressure corresponding to the estimated variation range of the torque capacity Tc2 of the clutch C2 is suppressed. Therefore, for example, the primary hydraulic cylinder 58c is compared with the case where the thrust ratio τ of the continuously variable transmission 24 that achieves the lowest speed ratio γmax is set to the maximum value and the primary pressure Pin of the primary hydraulic cylinder 58c is set to the minimum pressure. Since the amount of decrease in the primary pressure Pin can be reduced, it is possible to suppress a decrease in durability of the continuously variable transmission 24.

  Further, according to the present embodiment, the electronic control unit 100 includes the primary side slip guarantee thrust Wlmtin that is necessary and sufficient for securing the torque capacity of the primary pulley 58 of the continuously variable transmission 24, and the continuously variable transmission. The above-mentioned predetermined equation (1) with the input torque Tin of 24, the secondary side slip guarantee thrust Wlmtout which is a necessary and sufficient thrust to secure the torque capacity of the secondary pulley 62 of the continuously variable transmission 24, and no The above equation (2) for the input torque Tin of the continuously variable transmission 24 and the target speed ratio γtgt of the continuously variable transmission 24 for calculating the thrust ratio τ of the continuously variable transmission 24 are not used as parameters. A predetermined thrust ratio map of the input torque Tin of the stepped transmission 24 and the thrust ratio τ of the continuously variable transmission 24 is provided. Further, during the CtoC shift, the thrust control unit 110 performs the primary side slip of the primary pulley 58 of the continuously variable transmission 24 based on the estimated input torque Tlmtc for calculating the torque capacity based on the equation (1) or the equation (2). The guaranteed thrust Wlmtin and the secondary side slip guaranteed thrust Wlmtout of the secondary pulley 62 are calculated, and the target transmission ratio γtgt of the continuously variable transmission 24 is set to the lowest speed transmission ratio γmax formed by the continuously variable transmission 24, and the thrust Based on the ratio map, the thrust ratio τ that achieves the lowest speed ratio γmax of the continuously variable transmission 24 is calculated based on the lowest speed ratio γmax and the estimated input torque Tinc for thrust ratio calculation. Then, the thrust control unit 110 calculates a larger value of the secondary side shift control thrust Woutsh and the secondary side slip guarantee thrust Wlmtout calculated from the primary side slip guarantee thrust Wlmtin and the thrust ratio τ that achieves the lowest speed ratio γmax. The target secondary thrust Wouttgt is set, the target secondary thrust Wouttgt is divided by the thrust ratio τ that achieves the lowest speed ratio γmax of the continuously variable transmission 24, and the target primary thrust Wintgt of the primary pulley 58 is calculated. From Wintgt, the primary command pressure Pindir of the primary hydraulic cylinder 58c provided on the primary pulley 58 of the continuously variable transmission 24 is calculated. Therefore, the thrust set in consideration of the predetermined variation range of the torque capacity Tc2 with respect to the lowest speed ratio γmax of the continuously variable transmission 24 and the clutch command pressure Pc2dir to the CVT running clutch C2 from the thrust ratio map. Based on the ratio calculation estimated input torque Tinc, the thrust ratio τ that achieves the lowest speed ratio γmax of the continuously variable transmission 24 is calculated. As a result, an upshift of the continuously variable transmission 24 during the CtoC shift is suppressed, and a decrease in durability of the continuously variable transmission 24 is suppressed.

  According to the present embodiment, the thrust ratio map is such that the thrust ratio τ for achieving the target speed ratio γtgt of the continuously variable transmission 24 increases as the input torque Tin of the continuously variable transmission 24 decreases. It is. Therefore, the thrust ratio set in consideration of the predetermined variation range of the torque capacity Tc2 with respect to the lowest gear ratio γmax of the continuously variable transmission 24 and the clutch command pressure Pc2dir to the CVT travel clutch C2 from the thrust ratio map. The thrust ratio τ calculated based on the estimated input torque Tinc for calculation is based on the lowest speed ratio γmax of the continuously variable transmission 24 and the actual input torque Tin of the continuously variable transmission 24 from the thrust ratio map. It becomes larger than the calculated thrust ratio τ. As a result, the torque shift according to the actual input torque Tin at the lowest gear ratio γmax becomes the downshift side, and the actual gear ratio is maintained at the lowest gear ratio γmax, which is the target gear ratio γtgt.

  As mentioned above, although this invention was demonstrated in detail with reference to the table | surface and drawing, this invention can be implemented in another aspect, and can be variously changed in the range which does not deviate from the main point.

  For example, in the electronic control unit 100 of the above-described embodiment, the estimated input torque obtained by dividing the maximum variation Tc2max or the minimum variation Tc2min of the torque capacity Tc2 of the CVT travel clutch C2 by the speed ratio γc of the continuously variable transmission 24. The maximum variation value Tinesmax and the minimum variation value Tinesmin of the estimated input torque are set to the estimated input torque Tlmtc for torque capacity calculation or the estimated input torque Tinc for thrust ratio calculation. However, the present invention is not limited to this. For example, the actual engine torque Te and the CVT traveling are calculated from the equation of motion or the relation map which is obtained and stored in advance, which is composed of the engine torque Te output from the engine 12 and the torque capacity Tc2 of the CVT traveling clutch C2. Based on the maximum value Tc2max of the variation of the torque capacity Tc2 with respect to the clutch command pressure Pc2dir to the clutch C2, a maximum value Tinesmax of the estimated input torque is calculated, and the variation of the torque capacity Tc2 with respect to the clutch command pressure Pc2dir to the CVT travel clutch C2 Based on the minimum value Tc2min, the estimated input torque variation minimum value Tinesmin may be calculated.

  In the above-described embodiment, the gear mechanism 28 is a transmission mechanism in which one gear stage is formed. However, the present invention is not limited to this. For example, the gear mechanism 28 may be a transmission mechanism in which a plurality of gear stages having different gear ratios γ are formed. That is, the gear mechanism 28 may be a stepped transmission that is shifted to two or more stages.

  It should be noted that the above description is merely an embodiment, and other examples are not illustrated. However, the present invention should be implemented in variously modified and improved modes based on the knowledge of those skilled in the art without departing from the gist of the present invention. Can do.

12: Engine (drive power source)
16: Vehicle power transmission device 22: Input shaft (input rotation member)
24: continuously variable transmission (belt type continuously variable transmission)
28: Gear mechanism (gear-type transmission mechanism)
30: Output shaft (output rotating member)
58: Primary pulley (input pulley)
58c: Primary hydraulic cylinder (input hydraulic cylinder)
62: Secondary pulley (output pulley)
100: Electronic control device (control device for vehicle power transmission device)
C1: Forward clutch (first clutch)
C2: CVT travel clutch (second clutch)

Claims (1)

A first power transmission path that passes through the belt-type continuously variable transmission and a gear-type transmission mechanism are passed between an input rotating member that transmits power from the driving force source and an output rotating member that outputs the power to the driving wheels. A first clutch that is provided in parallel with a second power transmission path, and is engaged to select the first power transmission path, transmits torque more than the output-side pulley or the output-side pulley of the belt-type continuously variable transmission. In the vehicle power transmission device provided on the downstream side in the path,
A clutch-to-clutch shift that selectively selects the first power transmission path and the second power transmission path by switching the first clutch and the second clutch from one to the other is performed, and the belt type The input torque of the step transmission is estimated from the torque capacity of the first clutch, and the pressure of the input side hydraulic cylinder provided in the input side pulley of the belt-type continuously variable transmission is determined based on the torque capacity of the first clutch. A control device for a vehicle power transmission device to be controlled,
When calculating the pressure of the input side hydraulic cylinder during the clutch-to-clutch shift, a thrust for ensuring the torque capacity of the belt-type continuously variable transmission is calculated as the input torque of the belt-type continuously variable transmission. Using an estimated input torque for torque capacity calculation and an estimated input torque for calculating a thrust ratio for calculating a thrust ratio of the belt type continuously variable transmission,
The estimated input torque for torque capacity calculation is set based on the maximum value of torque capacity variation with respect to the command value to the first clutch,
The control device for a vehicle power transmission device, wherein the estimated input torque for thrust ratio calculation is set based on a minimum value of a variation in torque capacity with respect to a command value to the first clutch.

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