JP2011140995A - Control device of continuously variable transmission for vehicle - Google Patents

Abstract

An object of the present invention is to appropriately suppress the occurrence of overshoot in shifting of a continuously variable transmission.
Rotational change of actual input shaft rotational speed N _IN executed by feedback control based on rotational deviation ΔN _IN between target input shaft rotational speed N _IN ^* to be finally realized and actual input shaft rotational speed N _IN And the feed-forward control for setting the actual input shaft rotational speed N _IN to the second target input shaft rotational speed N _IN 2 ^* for starting shifting faster than when the target input shaft rotational speed N _IN ^* is used. by adding the rotation variation of the actual input shaft speed N _iN is, since the shift control to follow the actual input shaft rotational speed N _iN such that the target input shaft rotational speed N _iN ^* is performed, the rotation The shift start is accelerated by the feed forward control based on the second target input shaft rotational speed N _IN 2 ^* with respect to the shift delay due to the feedback control based on the deviation ΔN _IN. In addition, the occurrence of overshoot is appropriately suppressed when the vehicle starts.
[Selection] Figure 12

Description

  The present invention relates to a control device for a continuously variable transmission for a vehicle, and more particularly to a shift control for a continuously variable transmission.

  A control device for a continuously variable transmission for a vehicle that continuously changes a gear ratio by following an actual rotational speed of a rotating element to be subjected to a shift control so as to become a target rotational speed set based on a vehicle state. Well known. For example, the target value (target input rotation speed) of the input rotation speed of the continuously variable transmission is set based on the vehicle state represented by the acceleration request amount such as the accelerator opening and the transmission output rotation speed such as the vehicle speed, The continuously variable transmission is shifted by feedback control so as to eliminate the deviation between the target input rotation speed and the actual input rotation speed (actual input rotation speed). In such feedback control, the shift is executed only after the above-described deviation occurs. Therefore, a shift delay always occurs when the target input rotation speed changes. Therefore, Patent Documents 1 and 2 describe that, in addition to feedback control, feed-forward control that can execute shift based on a target value without waiting for detection of deviation is described.

JP 2006-308060 A JP 2006-308059 A JP 2005-299722 A

  Incidentally, as shown in FIG. 14, even in the shift control using the feedforward control in addition to the feedback control, the feedback control amount (FB control amount) should be sufficiently larger than the feedforward control amount (FF control amount). Thus, when shifting with feedback control as the main component, the above-described deviation occurs and the main gear shifting is executed. Therefore, for example, when the vehicle travels at the maximum speed ratio (lowest speed side speed ratio (maximum Lo)) up to the vehicle speed at the start of the vehicle and then shifts from the maximum speed ratio to the high vehicle speed side, Is generated and the main shift is started, so that an overshoot of the actual input rotation speed may occur with respect to the target input rotation speed. On the other hand, in order to improve the followability of the actual input rotation speed, it is conceivable to increase the feedback control gain (FB control gain) for obtaining the FB control amount for the deviation. However, when the FB control gain is increased, the FB control amount fluctuates excessively, so that hunting of the actual input rotational speed may occur. In addition, assuming the above overshoot, it is conceivable to set a target input rotation speed at which a shift is started early as a target rotation speed to be realized (see FIG. 14). When starting lock-up control, the increase in the actual input rotation speed is suppressed, and the engine rotation speed is pulled in, resulting in a low rotation state. There is a possibility. As described above, when shifting the continuously variable transmission mainly using feedback control, there is a concern that the actual input rotation speed hunting, the engine rotation speed may be low, or a booming noise may occur. It is difficult to reduce the chute appropriately. Such a problem is not yet known.

  The present invention has been made against the background of the above circumstances, and the object of the present invention is to provide a continuously variable transmission for a vehicle that can appropriately suppress the occurrence of overshoot in the shifting of a continuously variable transmission. It is to provide a control device.

  In order to achieve the above object, the gist of the present invention is that (a) the actual rotational speed of the rotary element to be subjected to the shift control is made to follow the target rotational speed set based on the vehicle state. (B) a first target rotational speed to be finally realized and a speed faster than the case where the first target rotational speed is used. Two different target rotational speeds for setting the second target rotational speed for starting the shift are set, and (c) executed by feedback control based on a deviation between the first target rotational speed and the actual rotational speed. By adding the rotational change of the actual rotational speed and the rotational change of the actual rotational speed executed by feedforward control for setting the actual rotational speed to the second target rotational speed, 1 target rotation Is to perform the shift control said to follow the actual rotational speed so that the degree.

  In this way, the rotation change of the actual rotation speed executed by feedback control based on the deviation between the first target rotation speed to be finally realized and the actual rotation speed, and the actual rotation speed are obtained. By adding the rotational change of the actual rotational speed executed by the feedforward control for setting the second target rotational speed for starting the shift earlier than the case of using the first target rotational speed, Since the shift control for following the actual rotation speed is executed so as to be the first target rotation speed, the second target rotation speed is set for the shift delay caused by the feedback control based on the deviation. The shift start is accelerated by executing the feedforward control. Therefore, the occurrence of overshoot is appropriately suppressed in the shifting of the continuously variable transmission. Further, since the actual rotational speed can be made to follow the first target rotational speed without increasing the feedback control gain with respect to the first target rotational speed to be finally realized, shift hunting also occurs. do not do.

  Here, it is preferable that the rotation element to be subjected to the shift control is an input rotation element of the continuously variable transmission for a vehicle, and the second target rotation speed is a speed change to a gear ratio on the high vehicle speed side. The first target rotational speed is set to be lower than the first target rotational speed so as to start earlier than when the first target rotational speed is used. In this case, when the shift start is advanced by executing the feedforward control for setting the second target rotation speed, the actual rotation speed becomes lower than the first target rotation speed to be finally realized. Thus, a deviation occurs in the first target rotational speed. However, since the actual rotational speed is changed by executing the feedback control so as to eliminate the deviation, the actual rotational speed is changed in the continuously variable transmission. The occurrence of overshoot can be appropriately suppressed while suppressing the decrease in speed.

  Preferably, the control is gear shift control at the time of start of the vehicle, and the second target rotation speed is a process of increasing the target rotation speed as the vehicle speed is increased so as to be maintained at a speed ratio on the lowest vehicle speed side. The shift from the minimum vehicle speed side gear ratio to the high vehicle speed side gear ratio is set lower than the first target rotation speed so as to start at a lower vehicle speed than when the first target rotation speed is used. There is in being. In this way, when the vehicle starts, it is possible to appropriately suppress the occurrence of overshoot while suppressing a decrease in actual rotational speed.

  Preferably, the continuously variable transmission is a so-called so-called continuously variable transmission in which a transmission belt functioning as a power transmission member is wound around a pair of variable pulleys whose effective diameter is variable, and the gear ratio is continuously changed steplessly. Belt-type continuously variable transmission, a pair of cones rotated around a common axis and a plurality of rollers that can rotate around the axis, and are rotated between the pair of cones. It is constituted by a so-called traction type continuously variable transmission whose transmission gear ratio is variable by changing the intersection angle between the center and the shaft center.

It is a figure explaining the schematic structure of the power transmission path | route which comprises the vehicle to which this invention is applied. It is a block diagram explaining the principal part of the control system provided in the vehicle. It is a hydraulic circuit diagram which shows the principal part regarding the belt clamping pressure control, gear ratio control, etc. of a continuously variable transmission among hydraulic control circuits. It is a figure which shows an example of the speed change map used when calculating | requiring a target input shaft rotational speed in the speed change control of a continuously variable transmission. It is a figure which shows an example of the required hydraulic pressure map which calculates | requires required hydraulic pressure according to gear ratio etc. in the clamping pressure control of a continuously variable transmission. It is a figure which shows an example of the map previously calculated | required experimentally and memorize | stored in advance of engine rotational speed and engine torque by using required load as a parameter. It is a figure which shows an example of the map previously calculated | required experimentally as a predetermined operating characteristic of a torque converter, and was memorize | stored. It is a figure which illustrates the overshoot which occurs at the time of the shift start after starting, when a shift is made only by feedback control. It is a block diagram which shows the flow of the shift control which uses feedforward control in addition to feedback control so that one actual input shaft rotational speed may be implement | achieved using two different target input shaft rotational speeds. It is a figure which shows an example of the shift map used when calculating | requiring two different target input shaft rotational speeds in the shift control of a continuously variable transmission. It is a functional block diagram explaining the principal part of the control function of an electronic controller. It is a flowchart explaining the control operation for suppressing appropriately generation | occurrence | production of the overshoot in the principal part of the control action of an electronic control unit, ie, the speed change of a continuously variable transmission at the time of vehicle start. It is a time chart which shows an example at the time of performing the control action shown to the flowchart of FIG. It is a block diagram which shows the flow of the conventional transmission control which uses feedforward control together with feedback control.

  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 power transmission path from an engine 12 to a drive wheel 24 constituting a vehicle 10 to which the present invention is applied. In FIG. 1, the power generated by the engine 12 includes a torque converter 14 as a fluid transmission device, a forward / reverse switching device 16, a continuously variable transmission for a vehicle (hereinafter referred to as a continuously variable transmission (CVT)) 18, a deceleration. It is transmitted to the left and right drive wheels 24 via the gear device 20, the differential gear device 22, and the like.

  The torque converter 14 includes a pump impeller 14p connected to the crankshaft 13 of the engine 12, a turbine impeller 14t connected to the forward / reverse switching device 16 via a turbine shaft 30 corresponding to an output side member of the torque converter 14, And a stator impeller 14s that is prevented from rotating in one direction by a one-way clutch, and power is transmitted between the pump impeller 14p and the turbine impeller 14t via a fluid. . Further, a lockup clutch 26 is provided between the pump impeller 14p and the turbine impeller 14t, which can be directly connected between them, that is, between the input and output of the torque converter 14. Further, the pump impeller 14p controls the transmission of the continuously variable transmission 18, generates the belt clamping pressure of the continuously variable transmission 18, controls the operation of the lockup clutch 26, or lubricates each part. And a mechanical oil pump 28 that is generated when the engine 12 is rotationally driven by a hydraulic pressure for supplying the oil.

Lockup clutch 26, As is well known, the differential pressure ΔP of the hydraulic control circuit 100 and the hydraulic pressure _{P ON} in the engagement-side oil chamber 14on the hydraulic _{P OFF} of the disengagement-side oil chamber 14off _{(= P ON} -P _OFF ) is a hydraulic friction clutch that is frictionally engaged with the front cover 14c by being controlled. The operating state of the torque converter 14 is, for example, so-called lockup release (lockup off) in which the differential pressure ΔP is negative and the lockup clutch 26 is released, and the differential pressure ΔP is zero or more and the lockup clutch 26 is A so-called lock-up slip state (slip state) half-engaged with slip, and a so-called lock-up state (engaged state, lock-up) in which the differential pressure ΔP is maximized and the lock-up clutch 26 is completely engaged. On). For example, when the lockup clutch 26 is completely engaged (lockup on), the pump impeller 14p and the turbine impeller 14t are integrally rotated, and the power of the engine 12 is directly transmitted to the continuously variable transmission 18 side. The Further, by the differential pressure ΔP to engagement is controlled in a predetermined slip state, for example, input and output rotational speed difference (i.e. slip speed (slip amount) = the engine rotational speed N _E - turbine speed N _T) by N _S is feedback controlled, while to follow rotation of the turbine shaft 30 relative to the crank shaft 13 by a drive (power-on) times a predetermined slip amount of the vehicle 10, the non-driving (power off) of the vehicle sometimes given The crankshaft 13 is rotated following the turbine shaft 30 by the slip amount. In the slip state of the lock-up clutch 26, since the differential pressure ΔP is made zero, the torque sharing of the lock-up clutch 26 is lost, and the torque converter 14 is operated under the same operating conditions as the lock-up off. .

  The forward / reverse switching device 16 is mainly composed of a forward clutch C1 and a reverse brake B1 as a starting clutch, and a double pinion type planetary gear device 16p. The turbine shaft 30 of the torque converter 14 is integrally connected to the sun gear 16s, and the input shaft 32 as an input rotation element of the continuously variable transmission 18 is integrally connected to the carrier 16c, while the carrier 16c and the sun gear 16s are connected to each other. Are selectively connected via a forward clutch C1, and the ring gear 16r is selectively fixed to a housing 34 as a non-rotating member via a reverse brake B1. The forward clutch C1 and the reverse brake B1 correspond to an intermittent device as a predetermined friction engagement device that transmits the power of the engine 12 to the drive wheel 24 side by engagement, and both are frictionally engaged by a hydraulic cylinder. The hydraulic friction engagement device.

  When the forward clutch C1 is engaged and the reverse brake B1 is released, the forward / reverse switching device 16 is brought into an integral rotation state, whereby the turbine shaft 30 is directly connected to the input shaft 32, and the forward drive power is increased. The transmission path is established (achieved), and the driving force in the forward direction is transmitted to the continuously variable transmission 18 side. When the reverse brake B1 is engaged and the forward clutch C1 is released, the forward / reverse switching device 16 establishes (achieves) the reverse power transmission path, and the input shaft 32 is connected to the turbine shaft 30. On the other hand, it is rotated in the opposite direction, and the driving force in the reverse direction is transmitted to the continuously variable transmission 18 side. When both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 16 enters a neutral state (power transmission cut-off state) in which power transmission is cut off.

As the engine 12, for example, a gasoline engine such as an internal combustion engine that generates power by combustion of fuel, a diesel engine, or the like is preferably used, but another prime mover such as an electric motor may be used in combination with the engine. The intake pipe 36 of the engine 12, the electronic throttle valve 40 for electrically controlling the intake air quantity Q _AIR of the engine 12 using the throttle actuator 38 is provided.

  The continuously variable transmission 18 includes an input side member provided on the input shaft 32 and a drive side pulley (primary pulley, primary sheave) 42 having a variable effective diameter, and an output provided on an output shaft 44 as an output rotation element. A driven pulley (secondary pulley, secondary sheave) 46 having a variable effective diameter, which is a side member, and a transmission belt 48 wound around both of these variable pulleys 42, 46. This is a belt-type continuously variable transmission in which power is transmitted via a frictional force between 46 and a transmission belt 48.

Both variable pulleys 42 and 46 are fixed rotating bodies 42 a and 46 a fixed to the input shaft 32 and the output shaft 44, respectively, and are not rotatable relative to the input shaft 32 and the output shaft 44 and movable in the axial direction. And a driving side hydraulic cylinder (primary pulley side hydraulic cylinder) 42c and a driven side hydraulic cylinder (secondary pulley) as hydraulic actuators that apply thrust to change the V groove width between them. Side hydraulic cylinder) 46c. Then, the hydraulic oil supply / discharge flow rate to the drive side hydraulic cylinder 42c is controlled by the hydraulic control circuit 100, whereby the V groove widths of the variable pulleys 42 and 46 change, and the engagement diameter (effective diameter) of the transmission belt 48 changes. ) Is changed, and the gear ratio γ (= input shaft rotational speed N _IN / output shaft rotational speed N _OUT ) is continuously changed. Further, the secondary pulley pressure (hereinafter referred to as belt clamping pressure) Pd, which is the hydraulic pressure of the driven hydraulic cylinder 46c, is regulated by the hydraulic control circuit 100, so that the transmission belt 48 does not slip. Is controlled. As a result of such control, a primary pulley pressure (hereinafter referred to as shift control pressure) Pin, which is the hydraulic pressure of the drive side hydraulic cylinder 42c, is generated.

  FIG. 2 is a block diagram illustrating a main part of a control system provided in the vehicle 10 for controlling the engine 12, the forward / reverse switching device 16, the continuously variable transmission 18, and the like. In FIG. 2, the vehicle 10 is provided with an electronic control device 50 including a control device for a vehicle continuously variable transmission for hydraulic control related to, for example, shift control of the continuously variable transmission 18 and belt clamping pressure control. ing. The electronic control unit 50 includes, for example, a so-called microcomputer having a CPU, a RAM, a ROM, an input / output interface, and the like. 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. For example, the electronic control unit 50 performs output control of the engine 12, shift control of the continuously variable transmission 18, belt clamping pressure control, torque capacity control of the lockup clutch 26, and the like. It is configured separately for engine control, continuously variable transmission 18 and lockup clutch 26 hydraulic control.

The electronic control unit 50, for example, the rotation angle (position) of the crankshaft 13 detected by the crankshaft rotation speed sensor 52 A rotational speed of the _CR and the crankshaft 13 (that is, the engine rotational speed) crankshaft rotation corresponding to N _E A signal representing the speed, a signal representing the rotational speed (turbine rotational speed) _NT of the turbine shaft 30 detected by the turbine rotational speed sensor 54, and an input rotational speed of the continuously variable transmission 18 detected by the input shaft rotational speed sensor 56. _Is a signal representing the rotational speed of the input shaft 32 (input shaft rotational speed) N _IN, and the rotational speed of the output shaft 44 (output shaft) which is the output rotational speed of the continuously variable transmission 18 detected by the output shaft rotational speed sensor 58. rotational _{speed) N OUT} that is, the signal representing the vehicle speed V corresponding to the output shaft speed _{N OUT,} is detected by a throttle sensor 60 engine 12 signal representing the throttle valve opening theta _TH of the electronic throttle valve 40 provided in an intake pipe 36 (see FIG. 1) of a signal representing the cooling water temperature T _W of the engine 12 detected by a coolant temperature sensor 62, CVT oil A signal representing the oil temperature T _CVT of the hydraulic oil in the hydraulic control circuit 100 such as the continuously variable transmission 18 detected by the temperature sensor 64, an accelerator pedal as the driver's acceleration request amount detected by the accelerator opening sensor 66 A signal representing the accelerator opening Acc, which is an operation amount of 68, a signal representing the intake air amount Q _AIR of the engine 12 detected by the intake air amount sensor 70, and a foot brake which is a service brake detected by the foot brake switch 72 signal representing the manipulated brake oN B _oN, operation of the shift lever 76 detected by a lever position sensor 74 Jishon such as a signal representative of the (operating position) P _SH is supplied.

Further, the electronic control unit 50, for example, as an engine output control command signal S _E for the output control of the engine 12, the drive signal and the fuel injection system to a throttle actuator 38 for controlling the opening and closing of the electronic throttle valve 40 78 An injection signal for controlling the amount of fuel injected from the engine, an ignition timing signal for controlling the ignition timing of the engine 12 by the ignition device 80, and the like are output. Further, for driving the solenoid valve DS1 and the solenoid valve DS2 for controlling the flow of hydraulic fluid to the shift control command signal S _T for example drive side hydraulic cylinder 42c for changing the speed ratio γ of the continuously variable transmission 18 hydraulic command signal, oil pressure command signal for driving the squeezing force control command signal S _B for example, a linear solenoid valve pressure belt clamping pressure Pd adjusted SLS for aligning clamping pressure of the transmission belt 48, the engagement of the lock-up clutch 26, release, hydraulic pressure command signal and the lock-up clutch 26 for driving the linear solenoid valve for switching the valve position of the lock-up relay valve of the lockup control command signal S _L for example, a hydraulic control circuit 100 for controlling the slip amount N _S push hydraulic pressure command signal for driving a linear solenoid valve for adjusting the torque capacity T _C of the line pressure P _L tone A hydraulic command signal for driving a linear solenoid valve is output to the hydraulic control circuit 100.

  The shift lever 76 is disposed in the vicinity of the driver's seat, for example, and is one of five operation positions “P”, “R”, “N”, “D”, and “L” that are sequentially positioned. To be manually operated. The “P” position releases the power transmission path of the vehicle 10, that is, a neutral state (neutral state) in which the power transmission of the vehicle 10 is interrupted and mechanically prevents (locks) the rotation of the output shaft 44 by the mechanical parking mechanism. The “R” position is a reverse travel position (position) for reversing the rotation direction of the output shaft 44, and the “N” position is the power transmission of the vehicle 10 is cut off. The “D” position is a forward travel for executing the automatic shift control by establishing the automatic shift mode in the shift range that allows the shift of the continuously variable transmission 18. Position (position), "L" position is the engine brake position (position) for applying strong engine brake . As described above, the “P” position and the “N” position are non-traveling positions that are selected when the power transmission path is in the neutral state and the vehicle is not traveling, and are “R” position, “D” position, and “L”. The position is a travel position that is selected when the vehicle 10 travels with the power transmission path in a power transmission enabled state that enables power transmission through the power transmission path.

  FIG. 3 is a hydraulic circuit diagram showing the main part of the hydraulic control circuit 100 related to belt clamping pressure control, gear ratio control, etc. of the continuously variable transmission 18. In FIG. 3, a hydraulic control circuit 100 includes a transmission ratio control valve UP116 that functions as a transmission control valve that controls the flow rate of hydraulic oil to the drive hydraulic cylinder 42c and a transmission ratio so that the transmission ratio γ can be continuously changed. The control valve DN118, the clamping pressure control valve 120 that regulates the belt clamping pressure Pd, which is the hydraulic pressure of the driven hydraulic cylinder 46c, and the forward clutch C1 and the reverse brake B1 are engaged or released so that the transmission belt 48 does not slip. As described above, a manual valve 122 or the like that mechanically switches the oil passage according to the operation of the shift lever 76 is provided.

Here, the first line oil pressure _PL1 in the oil pressure control circuit 100 is, for example, a relief type primary using the working oil pressure output (generated) from the mechanical oil pump 28 driven to rotate by the engine 12 as a source pressure. become regulator valve as pressure is adjusted to a value corresponding to the input torque T _iN or the like into (the first line pressure regulating valve) 110 by the linear output of the solenoid valve the hydraulic and is based on a control hydraulic CVT 18 Yes. Further, the second line oil pressure P _L2 is, for example, a relief type secondary regulator valve (for example, a relief type secondary regulator valve (for example, using the oil pressure discharged from the primary regulator valve 110 for regulating the first line oil pressure P _L1 by the primary regulator valve 110). The second line hydraulic pressure adjusting valve) 112 adjusts the pressure based on the control hydraulic pressure that is the output hydraulic pressure of the linear solenoid valve. Moreover, modulator pressure P _M is adapted to be pressure regulated to a constant pressure based on the control oil pressure which is the output oil pressure of the linear solenoid valve by the modulator valve 114 as source pressure for example the first line pressure P _L1.

The transmission ratio control valve UP116 is provided so as to be movable in the axial direction, thereby opening and closing the input / output port 116t and the input / output port 116i, and the spool valve element 116a as an input / output port 116t and an input / output port 116i. Spring 116b as an urging means for urging in a direction in which the input / output port communicates with each other, and a thrust in a direction in which the input / output port 116t and the input / output port 116i communicate with each other are accommodated in the spool 116a. In order to apply an oil chamber 116c that receives a control oil pressure _PS2 that is an output oil pressure of the solenoid valve DS2 that is duty-controlled by the electronic control device 50, and a thrust in a direction to close the input / output port 116i to the spool valve element 116a. Soleno that is duty controlled by the electronic control unit 50 And an oil chamber 116d for receiving a control oil pressure _PS1 which is an output oil pressure of the id valve DS1. The transmission ratio control valve DN118 is provided so as to be movable in the axial direction, and serves as a spool valve element 118a that opens and closes the input / output port 118t, and an urging unit that urges the spool valve element 118a in the valve closing direction. A spring 118b, an oil chamber 118c that accommodates the spring 118b and receives the control hydraulic pressure _{PS1 in order} to give a thrust in the valve closing direction to the spool valve element 118a, and a thrust in the valve opening direction to the spool valve element 118a Therefore, an oil chamber 118d for receiving the control oil pressure _PS2 is provided.

The solenoid valve DS1 is controlled to supply hydraulic oil to the drive side hydraulic cylinder 42c to increase its hydraulic pressure and to reduce the V groove width of the drive side pulley 42 to reduce the speed ratio γ, that is, control to the upshift side. The hydraulic pressure _PS1 is output. Further, the solenoid valve DS2 discharges the hydraulic oil in the drive side hydraulic cylinder 42c, lowers its hydraulic pressure, increases the V groove width of the drive side pulley 42, and controls to the side that increases the gear ratio γ, that is, the downshift side. Control oil pressure _PS2 . Specifically, the first line pressure _{P L1} results is supplied to the drive side hydraulic cylinder 42c via the input and output ports 116t input to the supply port 116s and control hydraulic pressure _{P S1} is output speed ratio control valve UP116 Therefore, the shift control pressure Pin is continuously controlled. When the control oil pressure _PS2 is output, the hydraulic oil in the drive side hydraulic cylinder 42c is discharged from the discharge port 118x via the input / output port 116t, the input / output port 116i, and the input / output port 118t, resulting in the shift control pressure Pin. Are continuously controlled. For example, as shown in FIG. 4, the relationship between the vehicle speed V and the target input shaft rotational speed N _IN ^{* that} is experimentally obtained and stored in advance using the accelerator operation amount Acc corresponding to the driver's acceleration request amount as a parameter (speed change) The deviation ΔN _IN (= N _IN ^* −) of the target input shaft rotational speed N _IN ^* calculated according to the map) so that the actual input shaft rotational speed N _IN (actual input shaft rotational speed N _IN ) matches. N _IN ), for example, the continuously variable transmission 18 is controlled by feedback control, that is, the transmission control pressure Pin is controlled by supplying and discharging hydraulic fluid to the drive side hydraulic cylinder 42c, and the transmission ratio γ continuously changes. Be made. The shift map in FIG. 4 corresponds to the shift conditions, and the target input shaft rotational speed N _IN ^* is set such that the larger the vehicle speed V is and the larger the accelerator opening Acc is, the larger the gear ratio γ is. Further, since the vehicle speed V corresponds to the output shaft rotation speed N _OUT, which is the target value of the input shaft rotational speed N _IN target input shaft rotational speed N _IN ^* corresponds to the target speed ratio, the minimum of the continuously variable transmission 18 It is determined within the range of the gear ratio γmin and the maximum gear ratio γmax.

Returning to FIG. 3, the clamping pressure control valve 120 is provided, for example, so as to be movable in the axial direction, and opens and closes the output port 120 t, and the biasing force that biases the spool valve element 120 a in the valve opening direction. A spring 120b as a means, and a control oil pressure P which is an output oil pressure of a linear solenoid valve SLS duty-controlled by the electronic control unit 50 in order to accommodate the spring 120b and apply a thrust force in the valve opening direction to the spool valve element 120a. _An oil chamber 120c that receives the _SLS and a feedback oil chamber 120d that receives the belt clamping pressure Pd that is output to give the spool valve element 120a thrust in the valve closing direction are provided. The clamping force control valve 120 is adapted to control hydraulic pressure P _SLS from the linear solenoid valve SLS to the first line pressure P _L1 to control continuously regulating control outputs the belt clamping pressure Pd as a pilot pressure . For example, the input torque T _IN is belt slippage does so obtained in advance experimentally caused stored as a parameter the speed ratio γ and the required oil pressure of the continuously variable transmission 18 corresponding to the transmitted torque, as shown in FIG. 5 ( The belt clamping pressure Pd to the driven hydraulic cylinder 46c is regulated in accordance with the relationship (belt clamping pressure map) with Pd ^* (corresponding to the target belt clamping pressure). The frictional force between 42 and 46 and the transmission belt 48 is increased or decreased. The belt clamping pressure Pd in the driven hydraulic cylinder 46c, which is the output hydraulic pressure of the clamping pressure control valve 120, is detected by the hydraulic sensor 120s.

The input torque T _IN of the continuously variable transmission 18 used when adjusting the belt clamping pressure Pd is, for example, the engine torque _TE to the torque ratio t of the torque converter 14 (= the output torque of the torque converter 14 (hereinafter referred to as the turbine torque T). is calculated by the electronic control unit 50 as hereinafter) / input torque of the torque converter 14 (hereinafter referred to as the pump torque _{T P))} multiplied by the torque _{(= T E × t) T} . The engine torque T _E, for example pre-experiments with the engine rotational speed N _E and engine torque T _E of the intake air quantity Q _AIR as the required load of the engine 12 _(or throttle opening theta _TH like equivalent) as a parameter The electronic control unit 50 calculates the estimated engine torque T _E es based on the intake air amount Q _AIR and the engine rotational speed N _E from the relationship (map, engine torque characteristic diagram) shown in FIG. Calculated. Alternatively, the engine torque T _E, for example the actual output torque (actual engine torque) of the engine 12 detected by such a torque sensor such as a T _E may be used. The torque ratio t of the torque converter 14 is expressed as follows: the speed ratio e of the torque converter 14 (= the output rotational speed of the torque converter 14 (hereinafter referred to as turbine rotational speed _NT ) / the input rotational speed of the torque converter 14 (hereinafter referred to as pump rotational speed N). _P (referred to as engine speed N _E ))), for example, the speed ratio e, the torque ratio t, the efficiency η, and the capacity coefficient C, which are experimentally obtained and stored in advance, as shown in FIG. From the relationship (map, predetermined operation characteristic diagram of the torque converter 14), the electronic control unit 50 calculates the actual speed ratio e. The estimated engine torque T _E es is calculated so as to represent the actual engine torque T _E itself, and unless otherwise distinguished from the actual engine torque T _E , the estimated engine torque T _E es is calculated as the actual engine torque T _E es. it is assumed that the handling of as a T _E. Accordingly, the estimated engine torque T _E es includes the actual engine torque T _E.

Returning to Figure 3, the manual valve 122, a constant oil pressure to the pressure-regulated the modulator pressure P _M is provided by the input port 122a for example modulator valve 114. When the shift lever 76 is operated to the "D" position or "L" position, and the reverse brake modulator pressure P _M is supplied to the forward clutch C1 via a forward output port 122f as forward running output pressure The oil passage of the manual valve 122 is switched so that the hydraulic oil in B1 is drained (discharged), for example, to the atmospheric pressure from the reverse output port 122r via the discharge port EX, and the forward clutch C1 is engaged and reversely moved. The brake B1 is released.

Further, when the shift lever 76 is operated to the "R" position, modulator pressure P _M is the hydraulic fluid in the fed and the forward clutch C1 to the reverse brake B1 via the reverse output port 122r as reverse running output pressure Is switched from the forward output port 122f via the discharge port EX to the atmospheric pressure, for example, so that the oil passage of the manual valve 122 is switched, the reverse brake B1 is engaged, and the forward clutch C1 is released. Be made.

  Further, when the shift lever 76 is operated to the “P” position or the “N” position, the oil path from the input port 122a to the forward output port 122f and the oil path from the input port 122a to the reverse output port 122r are both And the oil path of the manual valve 122 is switched so that the hydraulic oil in the forward clutch C1 and the reverse brake B1 is drained from the manual valve 122, and both the forward clutch C1 and the reverse brake B1 are connected. Be released.

Here, in the continuously variable transmission 18 of the present embodiment, feedback is performed based on the rotational deviation ΔN _IN (= N _IN ^* −N _IN ) between the target input shaft rotational speed N _IN ^* and the actual input shaft rotational speed N _IN. Since the continuously variable transmission 18 is controlled to be shifted by the control, that is, the shift control is started only after the rotation deviation ΔN _IN occurs, for example, as shown in FIG. overshoot of the actual input shaft rotation speed _{N iN} is always generated for N _iN ^*. Further, in a continuously variable transmission that uses a hardware minimum LOW as the maximum gear ratio γmax (minimum vehicle speed side gear ratio (lowest LOW)), for example, between the fixed rotating body 42a and the movable rotating body 42b in the variable pulley 42. In the continuously variable transmission in which the position of the movable rotating body 42b for forming the maximum gear ratio γmax with the V groove width as the maximum is mechanically positioned, the accelerator opening Acc as shown in FIG. The vehicle travels at the gear ratio γ that is the lowest in the hardware until the corresponding shift start vehicle speed V ′. When the shift from γmax to the high vehicle speed side gear ratio starts, the engagement position of the transmission belt 48 in both variable pulleys 42 and 46 is displaced. In this case, the shift control pressure Pin required for the displacement, a predetermined thrust ratio corresponding to the gear ratio gamma tau (= secondary pulley side hydraulic cylinder thrust W _{OUT /} primary pulley side hydraulic cylinder thrust W _IN; W _OUT is receiving area of the belt clamping pressure Pd × secondary pulley side hydraulic cylinder 46c, W _iN is to allow τ large thrust ratio than the pressure receiving area) of the shift control pressure Pin × primary pulley side hydraulic cylinder 42c, calculated from the thrust ratio τ Since a hydraulic pressure higher than the required shift control pressure is required, a delay occurs until the required shift control pressure is reached, and this causes overshoot. Note that, as shown in FIG. 14, even in the shift control used in combination with the feedback control and the feedforward control, by making the FB control amount sufficiently larger than the FF control amount (that is, the FB control gain is controlled by the FF control). When shifting with feedback control as the main component (by making it sufficiently larger than the gain), overshoot occurs as in the case of shifting with feedback control alone. Note that throughout the present specification, the case of shifting with feedback control as a main component and the case of shifting with only feedback control are treated as equivalent.

For such an overshoot, in order to improve followability of the actual input shaft rotational speed N _IN, it is conceivable to increase the FB control gain for determining the FB control amount with respect to the rotation deviation .DELTA.N _IN. However, increasing the FB control gain, since the FB control amount is to vary sensitively, there is a possibility that hunting of the actual input shaft speed N _IN is generated. Further, assuming the overshoot, the shift is started so as to start the shift earlier than the shift start vehicle speed V ′ at the original target input shaft rotation speed N _IN ^* obtained from the shift map as shown in FIG. It is conceivable to set the target input shaft rotation speed. However, for example, upon starting of the vehicle 10 (vehicle start), in order to suppress the fuel consumption by suppressing racing of the engine rotational speed N _E, the engagement of the lock-up clutch 26 from a relatively low vehicle speed when the vehicle starts When the known start-up lock-up slip control for slip engagement is executed, the engine rotational speed _NE is drawn into the input shaft rotational speed N _IN that is restrained from rising, and the engine is in a low rotational state. There is a possibility that sound is generated or the rotation of the engine 12 itself is unstable. Thus, when executing the speed change of the continuously variable transmission 18 feedback control alone or feedback control mainly the actual input shaft rotational speed N _IN of the hunting, a low rotation state of the engine rotational speed N _E, or muffled sound It is difficult to appropriately reduce the overshoot.

Therefore, in this embodiment, the original target input shaft rotational speed N _IN ^* obtained from the shift map as shown in FIG. 4 as the first target rotational speed to be finally realized, and the target input shaft rotational speed N _IN Two different target rotational speeds are set with the second target input shaft rotational speed N _IN 2 ^* as the second target rotational speed for starting the shift earlier than when ^* is used. As shown in FIG. 9, the rotational change of the actual input shaft rotational speed N _IN executed by feedback control based on the rotational deviation ΔN _IN between the target input shaft rotational speed N _IN ^* and the actual input shaft rotational speed N _IN The target input is calculated by adding the actual input shaft rotational speed N _IN to the second target input shaft rotational speed N _IN 2 ^* and the rotational change of the actual input shaft rotational speed N _IN executed by feedforward control. Shift control is performed at the start of the vehicle that causes the actual input shaft rotational speed N _IN to follow the shaft rotational speed N _IN ^* .

FIG. 10 shows a shift map used when two different target rotational speeds of the target input shaft rotational speed N _IN ^* and the second target input shaft rotational speed N _IN 2 ^* are obtained in the shift control of the continuously variable transmission 18. It is a figure which shows an example, Comprising: It is a figure corresponded to the shift map of FIG. In FIG. 10, the target rotational speed indicated by the solid line is the target input shaft rotational speed N _IN ^*, which is the same as the target input shaft rotational speed N _IN ^* shown in the shift map of FIG. Further, the target rotational speed indicated by the two-dot chain line is the second target input shaft rotational speed N _IN 2 ^* , and only the two-dot chain line portion is different from the target input shaft rotational speed N _IN ^* . As is apparent from FIG. 10, the second target input shaft rotational speed N _IN 2 ^* is, for example, greater than that when the target input shaft rotational speed N _IN ^* is used for shifting from the maximum gear ratio γmax to the high vehicle speed side gear ratio. In order to start earlier, the two-dot chain line portion is set lower than the target input shaft rotational speed N _IN ^* . That is, the second target input shaft rotational speed N _IN 2 ^* is increased from the maximum speed ratio γmax to the high speed side shifting in the process of increasing the target speed as the vehicle speed V increases so as to be maintained at the maximum speed ratio γmax. For example, the vehicle speed for starting the shift from the maximum gear ratio γmax to the higher vehicle speed side gear ratio is the target input shaft rotation so that the shift to the ratio starts at a lower vehicle speed than when the target input shaft rotation speed N _IN ^* is used. as a low speed gear shift start vehicle speed V "than the shift start vehicle speed V 'in the rate N _iN ^*, is set lower than its target input shaft rotational speed N _iN ^* in the two-dot chain line portion. for example, shift start vehicle speed V "suppresses overshoot, it is set shift control command signal obtained in advance experimentally to put before the response delay of the shift control pressure Pin changes to S _T. The shift start vehicle speed V "determined by the second target input shaft rotational speed _NIN2 ^* is the shift start vehicle speed in the shift control, that is, the control start vehicle speed, and the shift start determined by the target input shaft rotational speed _NIN ^*. The vehicle speed V ′ is also a shift start vehicle speed that is actually desired to be realized at the actual input shaft rotational speed N _{IN In} the embodiment of FIG. 10, the target input shaft rotational speed N _IN ^* and the second target input shaft rotational speed N _IN 2 ^* is set on the same map, but may be set on a single shift map.

Hereinafter, the control function by the electronic control unit 50 will be described more specifically. FIG. 11 is a functional block diagram for explaining a main part of the control function by the electronic control unit 50. In FIG. 11, an engine output control unit, that is, an engine output control means 82 sends an engine output control command signal S _E , for example, a throttle signal, an injection signal, an ignition timing signal, etc., for the output control of the engine 12, respectively. Output to the injection device 78 and the ignition device 80. For example, the engine output control means 82 sets the target throttle valve opening θ _TH ^* as the throttle opening θ _TH for obtaining the target engine torque T _E ^* corresponding to the accelerator opening Acc, and the target engine torque T _E ^*. In addition to controlling the opening and closing of the electronic throttle valve 40 by the throttle actuator 38, the fuel injection amount is controlled by the fuel injection device 78, and the ignition timing is controlled by the ignition device 80.

The belt clamping pressure control unit, that is, the belt clamping pressure control means 84, for example, determines the input torque T _IN (= engine torque T _E × torque ratio t: T _E of the continuously variable transmission 18 from the belt clamping pressure map as shown in FIG. For example, the target belt clamping pressure Pd ^* is set based on the vehicle state indicated by the estimated engine torque T _E es) and the actual gear ratio γ (= N _IN / N _OUT ). Then, the belt clamping pressure control means 84, the target belt clamping pressure Pd ^* is squeezing force control command signal S _B for pressure regulates the belt clamping pressure Pd of the driven-side hydraulic cylinder 46c so as to obtain the hydraulic pressure control circuit 100 Output. The hydraulic control circuit 100 actuates the linear solenoid valve SLS so the belt clamping pressure Pd is increased or decreased pressure of the belt clamping pressure Pd adjusted according squeezing force control command signal S _B from the belt clamping pressure control unit 84. Thus, the belt clamping pressure control means 84, by which actuates the linear solenoid valve SLS to control the belt clamping pressure Pd according to the input torque T _IN of the continuously variable transmission 18, to the extent that the belt slippage does not occur The belt clamping pressure is controlled to be as low as possible to improve fuel efficiency.

The lock-up clutch control unit, that is, the lock-up clutch control means 86 is, for example, a lock-up release (lock-up off) region, a slip control region (lock-up slip control operation region), a lock using the throttle valve opening θ _TH and the vehicle speed V as variables. Based on a vehicle state indicated by an actual throttle valve opening θ _TH and a vehicle speed V from a previously stored relationship (not shown) (map, lockup area diagram) having an up control operation area (lockup on) area The switching of the operating state of the up clutch 26 is controlled. For example, the lock-up clutch control means 86 is operated in any one of the lock-up release region, the lock-up slip control operation region, and the lock-up control operation region of the lock-up clutch 26 based on the actual vehicle state from the lock-up region diagram. determines whether to output the lock-up control command signal S _L for switching to the switching or lock-up slip control operation to lock-up control operation of the lock-up of the lockup clutch 26 to the hydraulic control circuit 100. The lock-up clutch control unit 86 determines that a lock-up slip control execution region, and sequentially calculates an actual slip amount N _S of lockup clutch 26, the actual slip amount N _S is the target slip amount N to output _S ^* to become as a lock-up control command signal S _L to control the differential pressure ΔP to the hydraulic control circuit 100. Further, the lock-up clutch control means 86 is a lock for controlling the lock-up clutch 26 to be engaged while slip-engaging so that the engine rotational speed _NE is suppressed, for example, when the vehicle starts with the accelerator on. It functions as a start-up lock-up slip control means for executing a start-up lock-up slip control that outputs the up-control command signal _SL to the hydraulic control circuit 100. For example, the lock-up clutch control means 86 sets a target engine speed N _E ^* for achieving both fuel efficiency and power performance in accordance with the accelerator opening Acc, and sets the engine speed higher than the target engine speed N _E ^*. as target engine speed N _E ^* to the engine rotational speed N _E is maintained (convergence) suppresses the N _E blows up the controls toward the engaged while slip-engaged lockup clutch 26 . That is, by controlling the slip amount N _S of the input shaft rotational speed varies with target engine speed N _E ^* and the vehicle speed V N _{IN (turbine} rotational speed N _T), suppresses the racing of the engine rotational speed N _E At the same time, the target engine speed _NE ^* is maintained.

The hydraulic control circuit 100 activates the lock-up control command signal S released and slipping state to engage with the switching solenoid valve to be switched in the lock-up clutch 26 in accordance with _L from the lock-up clutch control unit 86. The hydraulic control circuit 100, the linear solenoid slip control such that the torque capacity T _C is increased or decreased at the slip state to the engagement of the lock-up control instruction signal lockup clutch 26 according to S _L from the lock-up clutch control unit 86 actuates the valve to control the slip amount N _S of lockup clutch 26 locks or engages the up clutch 26. For example, in a relatively high vehicle speed region, the slip-up loss (internal loss) of the torque converter 14 is achieved by locking up (completely engaging) the lock-up clutch 26 and directly connecting the pump impeller 14p and the turbine impeller 14t. To improve fuel economy. In a relatively low / medium speed region, a slip control (lock-up slip control) is performed by applying a predetermined slight slip between the pump impeller 14p and the turbine impeller 14t, thereby engaging the lock. The up operation area is expanded, the transmission efficiency of the torque converter 14 is improved, and the fuel efficiency is improved.

The shift control unit, that is, the shift control means 88 is based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening Acc from a shift map (two-dot chain line) as shown in FIG. The second target input shaft rotational speed N _IN 2 ^* as the FF target rotational speed used for feedforward control is set. Then, the shift control means 88 is based on, for example, a predetermined feed-forward control formula as shown in the following formula (1) for setting the actual input shaft rotational speed N _IN to the second target input shaft rotational speed N _IN 2 ^*. Based on the second target input shaft rotational speed N _IN 2 ^* , the hydraulic oil for the drive side hydraulic cylinder 42c for making the FF control primary pulley pressure Pin _FF (or the FF control primary pulley pressure Pin _FF ) as the FF control amount The hydraulic pressure command signal to the solenoid valve DS1 and the solenoid valve DS2, etc. corresponding to the flow rate is calculated. In this equation (1), _KFF is a feedforward constant (FF gain).
Pin _FF = K _FF × (f (N _IN 2 ^* )) (1)

Further, the shift control means 88 is a target as an FB target rotational speed used for feedback control based on a vehicle state indicated by an actual vehicle speed V and an accelerator opening Acc from a shift map (solid line) as shown in FIG. 10, for example. Set the input shaft rotation speed N _IN ^* . Then, the speed change control means 88 uses, for example, a target input shaft from a predetermined feedback control formula as shown in the following formula (2) for making the actual input shaft rotational speed N _IN coincide with the target input shaft rotational speed N _IN ^*. Based on the rotational speed N _IN ^* , the FB control primary pulley pressure Pin _FB as the FB control amount (or the solenoid corresponding to the flow rate of the hydraulic oil to the drive side hydraulic cylinder 42c for the FB control primary pulley pressure Pin _FB The hydraulic pressure command signal to the valve DS1 and the solenoid valve DS2) is calculated. In this equation (2), e is a rotational deviation ΔN _IN (= N _IN ^* −N _IN ) between the target input shaft rotational speed N _IN ^* and the actual input shaft rotational speed N _IN, and K _P is a feedback proportional constant (FB proportional). gain), a _{K I} is an integration constant (FB integral gain).
Pin _FB = (K _P × e + K _I × (∫edt)) (2)

Further, the shift control means 88, by adding the primary for FF control pulley pressure Pin _FF and FB control for the primary pulley pressure Pin _FB, controlling the primary pulley pressure as a final control amount _{Pinc (= Pin FF + Pin FB} ) Is calculated. The speed change control means 88 controls the flow rate of the hydraulic oil to the drive side hydraulic cylinder 42c based on the control primary pulley pressure Pinc, for example, thereby changing the V groove widths of the variable pulleys 42 and 46. shift control command signal for changing the (hydraulic pressure command) S _T and outputs to the hydraulic control circuit 100 continuously changes the speed ratio γ with. The hydraulic control circuit 100 actuates the solenoid valve DS1 and the solenoid valve DS2 so shifting of the continuously variable transmission 18 is executed in accordance with the shift control command signal S _T from the shift control unit 88 to the drive side hydraulic cylinder 42c The shift control pressure Pin is adjusted by supplying and discharging hydraulic oil. In this way, the FB control amount and the FF control amount are calculated from two different target values (target input shaft rotational speed N _IN ^* , second target input shaft rotational speed N _IN 2 ^* ), respectively, and one actual rotational speed N Control that suppresses overshoot at start by realizing _IN is referred to as overshoot countermeasure control as start-up control.

On the other hand, the shift control means 88 is, for example, a vehicle indicated by an actual vehicle speed V and an accelerator opening Acc from a shift map (solid line) as shown in FIG. Based on the state, the target input shaft rotation speed N _IN ^* is set. Then, Then, the shift control means 88, so that the actual input shaft speed _{N IN} coincides with the target input shaft rotational speed _{N IN} ^*, for example, the actual input shaft rotational speed _{N IN} and the target input shaft rotational speed _{N IN} ^* Based on the rotation deviation ΔN _IN (= N _IN ^* −N _IN ), the shift of the continuously variable transmission 18 is executed by, for example, feedback control. In other words, for example, during normal vehicle travel, the shift control means 88 sets the primary pulley pressure for control _FF to the primary pulley pressure Pin _FB for FB control by setting the primary pulley pressure Pin _FF for FF control to zero, and for the FB control. continuously changing the speed ratio γ and outputs a shift control command signal S _T based on the primary pulley pressure Pin _FB to the hydraulic control circuit 100.

  The vehicle state determination unit, that is, the vehicle state determination unit 90 is, for example, whether or not the vehicle is started when the accelerator pedal 68 is depressed from the vehicle stop state based on the accelerator opening degree Acc and the vehicle speed V, that is, the vehicle starts when the accelerator is turned on. Determine if it is time.

The start-time control execution determination unit, that is, the start-time control execution determination unit 92 is, for example, a second set by the shift control unit 88 when the vehicle state determination unit 90 determines that the vehicle starts when the accelerator is on. Whether to start the start-time control (overshoot countermeasure control) by the shift control means 88 based on whether or not the actual vehicle speed V has reached the shift start vehicle speed V "determined from the target input shaft rotational speed N _IN 2 ^* When the start time control execution determination unit 92 determines to start the overshoot countermeasure control by the shift control unit 88, the start control execution determination unit 92 turns the overshoot countermeasure control flag from OFF to ON (OFF → ON). When the overshoot countermeasure control flag is on, the shift control means 88 performs the primary control for FF control according to the above equation (1). -Ery pressure Pin _FF (FF control amount) is calculated, and overshoot countermeasure control is executed.

Further, the start-time control execution determination unit 92 determines whether the second target input shaft rotational speed N _IN 2 ^* coincides with the target input shaft rotational speed N _IN ^* set by the shift control unit 88 (that is, the target input). Based on the difference rotational speed (N _IN ^* −N _IN 2 ^* ) between the shaft rotational speed N _IN ^* and the second target input shaft rotational speed N _IN 2 ^* or not, a shift control means It is determined whether or not the overshoot countermeasure control by 88 is ended. Further, when it is determined that the start time control execution determination unit 92 ends the overshoot countermeasure control by the shift control unit 88, the overshoot countermeasure control flag is changed from ON to OFF (ON → OFF). The shift control means 88 sets the FF control primary pulley pressure Pin _FF to zero (FF control amount = 0) when the overshoot countermeasure control flag is OFF.

  FIG. 12 is a flowchart for explaining a control operation for appropriately suppressing the occurrence of the overshoot in the main part of the control operation of the electronic control unit 50, that is, the shift of the continuously variable transmission 18 at the start of the vehicle. It is repeatedly executed with a very short cycle time of about several tens of milliseconds. FIG. 13 is a time chart showing an example when the control operation shown in the flowchart of FIG. 12 is executed.

In FIG. 12, first, in step (hereinafter, step is omitted) S10 corresponding to the vehicle state determination means 90, it is determined whether or not the vehicle is starting when the accelerator is on based on the accelerator opening Acc and the vehicle speed V, for example. Determined. If the determination in S10 is negative, this routine is terminated, but if the determination is affirmative (time t1 in FIG. 13), in S20 corresponding to the start-time control execution determination means 92, for example, a shift as shown in FIG. Shift start vehicle speed V determined from the second target input shaft rotational speed N _IN 2 ^* for feedforward control set based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening Acc from the map (two-dot chain line) It is determined whether or not the start time control (overshoot countermeasure control) is started based on whether or not the actual vehicle speed V has reached “.” When the determination in S20 is negative (t2 in FIG. 13). For example, before the time point), in S30 corresponding to the start-time control execution determination means 92 and the shift control means 88, for example, the overshoot countermeasure control flag is kept off and F Controlling the primary pulley pressure Pin _FF is zero (FF control amount = 0). On the other hand, if the determination at S20 is affirmative (t2 time in FIG. 13) is the starting time control execution determination means 92 and the shift control means In S40 corresponding to 88, for example, the overshoot countermeasure control flag is changed from ON to OFF (ON → OFF), and the actual vehicle speed V and accelerator opening are determined from a shift map (two-dot chain line) as shown in FIG. The second target input shaft rotational speed N _IN 2 ^* for feedforward control is set based on the vehicle state indicated by Acc Next, in S50 corresponding to the shift control means 88, for example, the above equation (1) is used to Based on the second target input shaft rotational speed N _IN 2 ^* , the FF control primary pulley pressure Pin _FF (FF control amount) is calculated. 50, in S60 corresponding to the shift control means 88, for example, a target input for feedback control based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening Acc from a shift map (solid line) as shown in FIG. the shaft speed _{N iN} ^* is set. then, the shift control in unit 88 S70 corresponding to, for example, the equation (2) the target input shaft rotational speed from _{N iN} ^* primary pulley pressure for FB control based on Pin _FB Next, in S80 corresponding to the shift control means 88, for example, the FF control primary pulley pressure Pin _FF and the FB control primary pulley pressure Pin _FB are added. In S90 corresponding to the means 88, for example, the control primary pulley pressure Pinc (= Pin _FF + Pin) as the final control amount, for example. _FB ) is calculated.

As shown in FIG. 13, feedforward control based on the second target input shaft rotational speed N _IN 2 ^* is first started from time t2, and then feedback control based on the rotational deviation ΔN _IN is executed in addition to the feedforward control. Is done. When the second target input shaft rotational speed N _IN 2 ^* coincides with the target input shaft rotational speed N _IN ^* , the overshoot countermeasure control flag is turned from on to off (ON → OFF), and the primary pulley pressure Pin _FF for FF control is zero. (FF control amount = 0). After time t4, the actual input shaft rotational speed N _IN is caused to follow the target input shaft rotational speed N _IN ^* by feedback control. By the overshoot countermeasure control is performed, as shown in solid line with an overshoot of the actual input shaft speed N _IN is suppressed, lowering of the input shaft rotational speed N _IN is prevented. Note that the actual input shaft speed N _IN shown in dashed lines is an example of a case where the main feedback control, overshoot of the actual input shaft speed N _IN in the shift starts t3 after the time has occurred.

As described above, according to the present embodiment, the actual input executed by feedback control based on the rotational deviation ΔN _IN between the target input shaft rotational speed N _IN ^* to be finally realized and the actual input shaft rotational speed N _IN. rotation change and axial rotation speed N _iN, the actual input shaft rotational speed N _iN target input shaft rotational speed N _iN ^* second target input shaft rotational speed for starting the earlier transmission than with N _iN 2 ^* The FF control primary pulley pressure Pin _FF and the FB control primary pulley pressure Pin _FB are added to each other by adding the rotational change of the actual input shaft rotational speed N _IN executed by the feedforward control Accordingly, since the shift control to follow the actual input shaft speed N _iN such that the target input shaft rotational speed N _iN ^* is executed, based on the rotation deviation .DELTA.N _iN Fi Relative delay in transmission due to execution of Dobakku control, the shift start is advanced by the execution of the feed-forward control for the second target input shaft rotational speed N _IN 2 ^*. Therefore, the occurrence of overshoot is appropriately suppressed in the shifting of the continuously variable transmission 18. Further, the actual input shaft rotational speed N _IN is set to the target input shaft rotational speed without increasing the feedback control gain (FB proportional gain, FB integral gain) with respect to the target input shaft rotational speed N _IN ^* to be finally realized. Since it is possible to follow N _IN ^* , shift hunting does not occur.

Further, according to the present embodiment, the rotation element to be subjected to the shift control is the input rotation element (for example, the input shaft 32) of the continuously variable transmission 18, and the second target input shaft rotation speed N _IN 2 ^* is Since the shift (upshift) to the gear ratio on the high vehicle speed side is set lower than the target input shaft rotational speed N _IN ^* so as to start earlier than when the target input shaft rotational speed N _IN ^* is used, when the shift start is advanced by the execution of the feed-forward control for the second target input shaft rotational speed N _iN 2 ^*, the actual input shaft rotational speed N _iN is ultimately to the target input shaft to achieve rotational speed N _iN ^* smaller than the target input shaft rotational speed N _iN ^* and the rotational deviation .DELTA.N _iN becomes that occurs, the rotational deviation .DELTA.N actual input shaft rotational speed by executing the feedback control as to eliminate _iN N _I Because There is rotated changes can in the shift of the continuously variable transmission 18, to appropriately suppress the occurrence of overshoot while suppressing a decrease in the actual input shaft speed N _IN.

Further, according to the present embodiment, the overshoot countermeasure control is a shift control when the vehicle 10 is started, and the second target input shaft rotation speed N _IN 2 ^* is maintained at the minimum vehicle speed side speed ratio γmax. In the process of increasing the target input shaft rotational speed N _IN ^* as the vehicle speed V increases, the shift (upshift) from the minimum vehicle speed side gear ratio γmax to the high vehicle speed side gear ratio is changed to the target input shaft rotational speed N. Since it is set to be lower than the target input shaft rotational speed N _IN ^* so as to start at a lower vehicle speed than when using _IN ^* , a decrease in the actual input shaft rotational speed N _IN is suppressed when the vehicle starts. However, the occurrence of overshoot can be appropriately suppressed.

In other words, the second sets the target input shaft rotational speed N _IN 2 ^*, FF control amount required to achieve the second target input shaft rotational speed N _IN 2 ^* to shift start early assuming overshoot The calculation and output (FF control) is effective for the response delay of the shift control pressure (primary pulley pressure) Pin necessary for displacing the engagement position of the transmission belt 48 from the lowest hardware level, for example. At this time, a rotational deviation ΔN _IN occurs between the target input shaft rotational speed N _IN ^* and the actual input shaft rotational speed N _IN to be realized by speeding up the shift start (N _IN <N _IN ^* ). At this time, in order to calculate and output (FB control) the FB control amount necessary to eliminate the rotational deviation ΔN _IN , the engine rotational speed _NE is low due to the decrease in the actual input shaft rotational speed N _IN (ie, This is effective for the concern that the rotation of the engine 12 itself becomes unstable. In addition, two different types of the use of specific techniques of this embodiment that the target rotational speed (shift map) to realize one actual input shaft speed N _IN, be actually realized the target input shaft rotational speed N _Since it is possible to follow _IN ^* without increasing the FB control gain, occurrence of shift hunting is also suppressed.

  As mentioned above, although the Example of this invention was described in detail based on drawing, this invention is applied also in another aspect.

For example, in the above-described embodiment, the start-time control execution determination unit 92 (step S20 in FIG. 12) has reached the actual vehicle speed V at the shift start vehicle speed V ″ determined from the second target input shaft rotational speed N _IN 2 ^*. Whether or not the overshoot countermeasure control is to be started is determined based on whether or not it is performed, but is not necessarily limited to this. For example, the shift start vehicle speed V ″ may be a constant value. Further, it may be determined whether or not the overshoot countermeasure control is to be started based on whether or not a predetermined time has elapsed since the vehicle started with the accelerator on. In addition, the predetermined time may be a predetermined value set in advance in consideration of, for example, a delay in hydraulic response, or may be changed according to the accelerator opening Acc, for example, similarly to the shift start vehicle speed V ″. Various modes are conceivable, such as changing the shift start vehicle speed V ″ and the predetermined time in accordance with the changing speed of the vehicle speed V.

  In the above-described embodiment, the continuously variable transmission 18 using the hardware lowest LOW as the maximum gear ratio γmax (the gear ratio (lowest LOW) on the lowest vehicle speed side) is illustrated. The present invention can be applied even to a continuously variable transmission in which the maximum speed ratio γmax is maintained by the hydraulic pressure command signal.

  The above description is only an embodiment, and the present invention can be implemented in variously modified and improved forms based on the knowledge of those skilled in the art.

18: continuously variable transmission for vehicle 32: input shaft (input rotation element of continuously variable transmission for vehicle)
50: Electronic control device (control device)

Claims (3)

A control device for a continuously variable transmission for a vehicle that continuously changes a gear ratio by following an actual rotation speed of a rotation element to be subjected to a shift control so that a target rotation speed set based on a vehicle state is obtained. There,
Two different target rotational speeds, a first target rotational speed to be finally realized and a second target rotational speed for starting a shift earlier than when using the first target rotational speed, are set,
Rotational change of the actual rotational speed executed by feedback control based on a deviation between the first target rotational speed and the actual rotational speed, and feed for setting the actual rotational speed to the second target rotational speed A speed change control for causing the actual rotation speed to follow the first target rotation speed by adding the rotation change of the actual rotation speed executed by the forward control is performed. Control device for continuously variable transmission. The rotation element to be subject to the shift control is an input rotation element of the vehicle continuously variable transmission,
The second target rotation speed is set to be lower than the first target rotation speed so that the shift to the gear ratio on the high vehicle speed side is started earlier than the case where the first target rotation speed is used. The control device for a continuously variable transmission for a vehicle according to claim 1. Shift control when starting the vehicle,
In the process of increasing the target rotational speed as the vehicle speed increases so that the second target rotational speed is maintained at the speed ratio on the minimum vehicle speed side, the speed ratio on the minimum vehicle speed side is changed to the speed ratio on the high vehicle speed side. 3. The vehicle-use absence according to claim 2, wherein the vehicle speed is set to be lower than the first target rotation speed so that the first shift is started at a lower vehicle speed than when the first target rotation speed is used. Control device for step transmission.

Images