JP2008025678A - Control device of vehicular continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To improve controllability, in a control device controlling a continuously variable transmission based on estimated hydraulic pressure in an input-side hydraulic cylinder. <P>SOLUTION: Since a thrust ratio τ is calculated from a driving-time map by a thrust ratio calculation means 164 when it is determined by a fuel cut state determination means 170 that an engine 12 is not in a fuel cut state, and the thrust ratio τ is calculated from a driven-time map when it is determined that the engine is in a fuel cut state, the map is changed over based on the fuel cut executed at delayed timing relative to changeover timing from a driving state to a driven state, and the map is changed over at timing at which actual Pin pressure having response delay relative to the timing is actually lowered to driven-time actual Pin pressure. Therefore estimated Pin pressure calculated based on the thrust pressure τ is prevented from being set lower than the actual Pin pressure, and controllability in controlling the continuously variable transmission 18 based on the estimated Pin pressure is improved. <P>COPYRIGHT: (C)2008,JPO&INPIT

Description

  The present invention relates to a control device for a continuously variable transmission for a vehicle having an input side variable pulley and an output side variable pulley, and a belt wound around both pulleys, and in particular, an input side variable pulley and an output side variable pulley. The present invention relates to a technique for calculating an estimated value of an oil pressure in an input side hydraulic cylinder based on a thrust ratio of each hydraulic cylinder for changing a groove width and an oil pressure in an output side hydraulic cylinder.

  In a control device for a continuously variable transmission for a vehicle having an input-side variable pulley and an output-side variable pulley and a belt wound around both pulleys, an input-side hydraulic cylinder for changing the groove width of the input-side variable pulley It is well known to control a continuously variable transmission based on the acting hydraulic pressure. For example, the line oil pressure, which is the original pressure of the oil pressure, is controlled so that the oil pressure in the input side hydraulic cylinder (hereinafter referred to as the primary oil pressure) is properly obtained, or the speed is changed according to the flow rate of the working oil into the input side hydraulic cylinder. When the ratio is controlled, the flow rate (opening area of the flow control valve) is obtained based on the primary hydraulic pressure. When such a continuously variable transmission is controlled, if an oil pressure sensor for detecting the primary oil pressure is not provided, it is necessary to calculate an estimated value of the oil pressure.

  In this regard, Patent Document 1 discloses a thrust ratio that is a ratio of a thrust by the input-side hydraulic cylinder and a thrust by the output-side hydraulic cylinder for changing the groove width of the output-side variable pulley, A control device for a continuously variable transmission that calculates an estimated value of a primary hydraulic pressure based on a hydraulic pressure (hereinafter referred to as a secondary hydraulic pressure) has been proposed.

JP 2004-125209 A

  By the way, since the thrust ratio is different between when the vehicle is in a driving state and when it is in a driven state, the thrust ratio is calculated in consideration of whether the vehicle is in a driving state or in a driven state. It is necessary to calculate the estimated value.

  However, when the vehicle is switched from the driven state to the driven state, there is a response delay in the decrease in the actual primary hydraulic pressure. Therefore, the primary hydraulic pressure estimated based on the thrust ratio when the vehicle is in the driven state is estimated. The value may be lower than the actual primary hydraulic pressure, which may reduce the controllability of the continuously variable transmission. For example, if the estimated value of the primary oil pressure is lower than the actual primary oil pressure, the line oil pressure required for the primary oil pressure may be insufficient.

  The present invention has been made against the background of the above circumstances, and an object of the present invention is to provide a continuously variable transmission for a vehicle that controls a continuously variable transmission based on an estimated value of oil pressure in an input side hydraulic cylinder. In the control device, the controllability of the continuously variable transmission is to be improved.

  To achieve this object, the gist of the invention according to claim 1 is that (a) a power transmission path between an engine and a drive wheel is wound around an input side variable pulley, an output side variable pulley, and both pulleys. In a vehicle provided with a continuously variable transmission having a belt that is hung, the thrust by the input hydraulic cylinder and the groove width of the output variable pulley are changed to change the groove width of the input variable pulley. Input side hydraulic pressure estimated value calculation that calculates the estimated value of the hydraulic pressure in the input side hydraulic cylinder based on the thrust ratio that is the ratio of the thrust by the output side hydraulic cylinder and the hydraulic pressure in the output side hydraulic cylinder A control device for a continuously variable transmission for a vehicle that controls the continuously variable transmission based on an estimated value of hydraulic pressure in the input side hydraulic cylinder, and (b) when the vehicle is in a driving state Find the thrust ratio Storage means for storing a predetermined driving relationship and a predetermined driving relationship for determining a thrust ratio when the vehicle is in a driven state, and (c) the engine is fuel cut A fuel cut state determining means for determining whether or not the fuel cut state is present, and (d) while determining that the fuel cut state is not in the fuel cut state by the fuel cut state determining means, while obtaining the thrust ratio from the relationship at the time of driving, And a thrust ratio calculating means for determining the thrust ratio from the driven relationship when the fuel cut state determining means determines that the fuel cut state is established.

  In this way, when the thrust ratio calculating means determines that the fuel cut state determining means is not in the fuel cut state, the thrust ratio when the vehicle stored in the storage means is in the driving state is obtained in advance. While the thrust ratio is obtained from the obtained driving relationship, the thrust ratio when the vehicle stored in the storage means is in the driven state when the fuel cut state judging means determines that the fuel cut state is established. Based on the fuel cut of the engine executed at a timing delayed from the timing of switching from the driving state of the vehicle to the driven state The relationship (map) determined in advance for determining the thrust ratio will be switched, and the drive state will change to the driven state. Hydraulic pressure in the switching of the input-side hydraulic cylinder there is a response delay than the timing of the map is switched actually in reduced timing. Therefore, it is avoided that the estimated value of the hydraulic pressure in the input side hydraulic cylinder is lower than the actual hydraulic pressure, and the control when the continuously variable transmission is controlled based on the estimated value of the hydraulic pressure in the input side hydraulic cylinder. Improves.

  The invention according to claim 2 is the control device for a continuously variable transmission for a vehicle according to claim 1, wherein the line hydraulic pressure of the continuously variable transmission is based on the estimated value of the hydraulic pressure in the input hydraulic cylinder. Is further provided with a line hydraulic pressure control means for controlling. In this way, it is avoided that the line hydraulic pressure necessary for the hydraulic pressure in the input side hydraulic cylinder is insufficient.

  Here, preferably, in the normal transmission control of the continuously variable transmission, for example, the target transmission gear ratio is obtained according to a predetermined transmission condition, and the actual transmission gear ratio is set to the target transmission gear ratio. By changing the groove width of the input side variable pulley by controlling the flow rate of hydraulic fluid to the engine, feedback control of the gear ratio is performed, or the input side (engine) is changed according to vehicle speed, output rotation speed (drive wheel rotation speed), etc. Side), and change the groove width of the input-side variable pulley by controlling the flow rate of hydraulic oil to the input-side hydraulic cylinder so that the actual input rotation speed becomes the target rotation speed. Various modes such as feedback control of the ratio can be employed.

  The above-mentioned predetermined speed change conditions are set by a map or an arithmetic expression using, for example, driving conditions such as a driver output request amount (acceleration request amount) such as an accelerator operation amount and a vehicle speed (corresponding to an output rotation speed) as parameters. Is done.

  Further, the gear ratio of the continuously variable transmission is set to a predetermined value as a state in which hydraulic oil is confined in the input side hydraulic cylinder that is executed in the vehicle state when the feedback control is difficult such as when traveling at a low vehicle speed below a predetermined vehicle speed. The hydraulic control for the transmission ratio is performed by independently controlling the hydraulic pressure of the input side hydraulic cylinder and the hydraulic pressure of the output side hydraulic cylinder, and the predetermined thrust ratio τ (= the hydraulic pressure in the output side hydraulic cylinder × the pressure receiving area of the output side hydraulic cylinder) (The hydraulic pressure in the input side hydraulic cylinder × the pressure receiving area of the input side hydraulic cylinder) may be used to control the hydraulic pressure of the input side hydraulic cylinder. The hydraulic pressure of the input side hydraulic cylinder is controlled based on the control pressure output from the thrust ratio control valve. It makes it is desirable to configure such that a predetermined thrust ratio tau.

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

  FIG. 1 is a skeleton diagram illustrating the configuration of a vehicle drive device 10 to which the present invention is applied. This vehicle drive device 10 is a horizontal automatic transmission, which is suitably employed in an FF (front engine / front drive) type vehicle, and includes an engine 12 as a driving power source. The output of the engine 12 composed of an internal combustion engine is the crankshaft of the engine 12, the torque converter 14 as a fluid transmission device, the forward / reverse switching device 16, the belt type continuously variable transmission (CVT) 18, the reduction gear. It is transmitted to the differential gear device 22 via the device 20 and distributed to the left and right drive wheels 24L, 24R.

  The torque converter 14 includes a pump impeller 14p connected to the crankshaft of the engine 12 and a turbine impeller 14t connected to the forward / reverse switching device 16 via a turbine shaft 34 corresponding to an output side member of the torque converter 14. And power transmission is performed via a fluid. A lock-up clutch 26 is provided between the pump impeller 14p and the turbine impeller 14t, and a lock-up control valve (L not shown) in the hydraulic control circuit 100 (see FIGS. 2 and 3) is provided. The hydraulic pressure supply to the engagement side oil chamber and the release side oil chamber is switched by the (C / C control valve) or the like, thereby being engaged or released. The turbine impeller 14t is rotated integrally. The pump impeller 14p has a hydraulic pressure for controlling the transmission of the continuously variable transmission 18, generating a belt clamping pressure, controlling the engagement release of the lockup clutch 26, or supplying lubricating oil to each part. Is coupled to a mechanical oil pump 28 that is generated by being driven to rotate by the engine 12.

  The forward / reverse switching device 16 is mainly composed of a double pinion type planetary gear device, the turbine shaft 34 of the torque converter 14 is integrally connected to the sun gear 16s, and the input shaft 36 of the continuously variable transmission 18 is a carrier. The carrier 16c and the sun gear 16s are selectively connected via the forward clutch C1, and the ring gear 16r is selectively fixed to the housing via the reverse brake B1. It has become. The forward clutch C1 and the reverse brake B1 correspond to an intermittent device, both of which are hydraulic friction engagement devices that are frictionally engaged by a hydraulic cylinder.

  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 34 is directly connected to the input shaft 36, and the forward power 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 36 is connected to the turbine shaft 34. 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. Further, when both the forward clutch C1 and the reverse brake B1 are released, the forward / reverse switching device 16 becomes neutral (interrupted state) for interrupting power transmission.

  The continuously variable transmission 18 has an input-side variable pulley (primary pulley) 42 having a variable effective diameter that is an input-side member provided on the input shaft 36, and an effective diameter that is an output-side member provided on the output shaft 44. A variable output side variable pulley (secondary pulley) 46 and a transmission belt 48 wound around the variable pulleys 42 and 46 are provided, and a frictional force between the variable pulleys 42 and 46 and the transmission belt 48 is provided. Power is transmitted via the.

The variable pulleys 42 and 46 are fixed rotation bodies 42 a and 46 a fixed to the input shaft 36 and the output shaft 44, respectively. Provided movable rotors 42b and 46b, and an input side hydraulic cylinder (primary pulley side hydraulic cylinder) 42c and an output side hydraulic cylinder (secondary pulley side hydraulic cylinder) 46c that apply thrust to change the V groove width therebetween. The hydraulic fluid supply / discharge flow rate to the input side hydraulic cylinder 42c is controlled by the hydraulic control circuit 100, so that the V-groove widths of both variable pulleys 42 and 46 are changed to change the transmission belt. changed consuming diameter of 48 (effective diameter), the speed ratio gamma (= input shaft speed _{N iN} / output shaft speed _{N OUT)} continuous It is changed to. Further, the hydraulic pressure of the output side hydraulic cylinder 46c (belt clamping pressure Pd) is regulated by the hydraulic control circuit 100, whereby the belt clamping pressure is controlled so that the transmission belt 48 does not slip. As a result of such control, the hydraulic pressure (shift control pressure Pin) of the input side hydraulic cylinder 42c is generated.

  FIG. 2 is a block diagram for explaining a main part of a control system provided in the vehicle for controlling the vehicle drive device 10 of FIG. 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. By performing signal processing, 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 are executed. This is divided into control and hydraulic control for the continuously variable transmission 18 and the lockup clutch 26.

The electronic control unit 50, representing the crankshaft rotation speed corresponding to the engine rotational speed crankshaft detected by the sensor 52 rotation angle (position) A _{CR (°)} and the rotational speed of the engine 12 (engine rotational speed) N _E Signal, a signal representing the rotational speed (turbine rotational speed) _NT of the turbine shaft 34 detected by the turbine rotational speed sensor 54, and an input that is the input rotational speed of the continuously variable transmission 18 detected by the input shaft rotational speed sensor 56. rotational speed (input shaft rotational speed) signal representing the N _iN of the shaft 36, a vehicle speed sensor (output shaft rotation speed sensor) 58 by the rotational speed (output of the output shaft 44 is the output rotational speed of the continuously variable transmission 18 detected axis rotation speed) N _OUT ie vehicle speed signal representing a vehicle speed V corresponding to the output shaft speed N _OUT, en detected by the throttle sensor 60 Intake pipe 32 throttle valve opening signal representing the throttle valve opening theta _TH of the electronic throttle valve 30 provided in (see FIG. 1) of the emission 12, the cooling water temperature T _W of the engine 12 detected by a coolant temperature sensor 62 signals representative signal representative of the oil temperature T _CVT of a hydraulic circuit of the continuously variable such as transmission 18 detected by the CVT oil temperature sensor 64, the accelerator opening is an operation amount of the accelerator pedal 68 detected by the accelerator opening sensor 66 an accelerator opening signal representative of the acc, brake operation signal indicating whether B _ON operation of the foot brake is a service brake, which is detected by a foot brake switch 70, a lever position (operation of the shift lever 74 detected by a lever position sensor 72 Position) An operation position signal representing _PSH is supplied.

Further, the electronic control device 50 receives an engine output control command signal S _E for controlling the output of the engine 12, for example, a throttle signal for driving a throttle actuator 76 for controlling the opening / closing of the electronic throttle valve 30, and a fuel injection device 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, a command 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, the input-side hydraulic cylinder 42c for changing the speed ratio γ of the continuously variable transmission 18 signal, a 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 line for which control the line pressure P _L such command signal for driving the hydraulic control command signal S _PL example, a linear solenoid valve pressure line pressure P _L tone SLT is output to the hydraulic control circuit 100.

  The shift lever 74 is arranged, for example, in the vicinity of the driver's seat and is sequentially positioned in five lever positions “P”, “R”, “N”, “D”, and “L” (see FIG. 3). Any one of them is manually operated.

  The “P” position (range) releases the power transmission path of the vehicle drive device 10, that is, a neutral state (neutral state) where the power transmission of the vehicle drive device 10 is interrupted, and is mechanically output by the mechanical parking mechanism. The parking position (position) for preventing (locking) the rotation of 44, the “R” position is the reverse traveling position (position) for reversely rotating the output shaft 44, and the “N” position. Is a neutral position (position) for setting the neutral state in which the power transmission of the vehicle drive device 10 is interrupted, and the “D” position establishes an automatic transmission mode within a transmission range that allows the transmission of the continuously variable transmission 18. This is a forward travel position (position) that allows automatic shift control to be executed, and the “L” position is operated by a strong engine brake. An engine brake position (position). Thus, the “P” position and the “N” position are non-traveling positions that are selected when the vehicle is not traveling, and the “R” position, the “D” position, and the “L” position are when the vehicle is traveling. This is the selected driving position.

  FIG. 3 shows the main part of the hydraulic control circuit 100 relating to the belt clamping pressure control, the transmission gear ratio control of the continuously variable transmission 18, and the engagement hydraulic pressure control of the forward clutch C1 or the reverse brake B1 accompanying the operation of the shift lever 74. FIG. In FIG. 3, the hydraulic control circuit 100 continuously changes the clamping pressure control valve 110 that regulates the belt clamping pressure Pd that is the hydraulic pressure of the output hydraulic cylinder 46 c so that the transmission belt 48 does not slip, and the speed ratio γ continuously changes. The transmission ratio control valve UP114 and the transmission ratio control valve DN116 for controlling the flow rate of hydraulic fluid to the input side hydraulic cylinder 42c so that the ratio between the transmission control pressure Pin and the belt clamping pressure Pd is set in a predetermined relationship. A manual valve 120 or the like that mechanically switches the oil path according to the operation of the shift lever 74 is provided so that the thrust ratio control valve 118, the forward clutch C1 and the reverse brake B1 are engaged or released.

Line pressure P _L is the linear solenoid as a source pressure of working oil pressure output (generated) from the mechanical oil pump 28 which is rotated by the engine 12, for example, by the primary regulator valve (line pressure regulating valve) 122 of the relief type Based on the control hydraulic pressure _PSLT which is the output hydraulic pressure of the valve SLT, the pressure is adjusted to a value corresponding to the engine load, the shift control pressure Pin, the belt clamping pressure Pd, and the like.

More specifically, the primary regulator valve 122 is provided so as to be movable in the axial direction, thereby opening and closing the input port 122i and discharging the hydraulic pressure generated from the oil pump 28 to the intake oil passage 124 via the output port 122t. The spool valve element 122a, the spring 122b as an urging means for urging the spool valve element 122a in the valve closing direction, and the spring 122b for accommodating the spool valve element 122a and applying a thrust force in the valve closing direction to the spool valve element 122a. the control oil pressure and oil chamber 122c for receiving the P _SLT, and an oil chamber 122d that receives the hydraulic pressure generated from the oil pump 28 to apply a thrust force in the valve opening direction to the spool valve element 122a to.

In the primary regulator valve 122 configured as described above, and the biasing force _{F S} of the spring 122b, the pressure receiving area of the control oil pressure _{P SLT} in the oil chamber 122c a, the pressure receiving area difference of the line pressure _{P L} in the oil chamber 122d b Then, it will be in an equilibrium state by following Formula (1).
P _L × b = P _SLT × a + F _S (1)
Therefore, the line pressure _{P L} is represented by the following formula (2), is proportional to the control pressure _{P SLT.}
P _L = P _SLT × (a / b) + F _S / b (2)

Thus, the primary regulator valve 122 and the linear solenoid valve SLT, a line oil pressure control command signal S _PL pressure regulating pressure regulating hydraulic oil to the line pressure P _L to be discharged from the oil pump 28 on the basis of as the hydraulic pressure command value Functions as a device.

Modulator pressure P _M, as well is used as the basic pressure of the control oil pressure P _SLS is the output hydraulic pressure of the control pressure P _SLT and the linear solenoid valve SLS, by the electronic control unit 50 by the output hydraulic pressure of the solenoid valve DS1 that is duty-controlled a used as the basic pressure of a certain control oil pressure P _DS1 and the control pressure P _DS2 is the output hydraulic pressure of the solenoid valve DS2, the modulator valve 126 to line pressure P _L as source pressure adapted to be pressure regulated to a constant pressure ing.

Output hydraulic pressure _{P LM2} is adapted to line pressure _{P L} to be pressure regulated on the basis of the control hydraulic pressure _{P SLT} by the line pressure modulator NO.2 valve 128 as an original pressure.

In the manual valve 120, the output oil pressure _PLM2 is supplied to the input port 120a. When the shift lever 74 is operated to the “D” position or the “L” position, the output hydraulic pressure _PLM2 is supplied to the forward clutch C1 via the forward output port 120f as the forward travel output pressure and the reverse brake. The oil passage of the manual valve 120 is switched so that the hydraulic oil in B1 is drained (discharged) to the atmospheric pressure, for example, from the reverse output port 120r through the discharge port EX, and the forward clutch C1 is engaged and reversely moved. The brake B1 is released.

Further, when the shift lever 74 is operated to the "R" position, output pressure P _LM2 is the hydraulic fluid in the fed and the forward clutch C1 to the reverse brake B1 via the reverse output port 120r as reverse running output pressure Is switched from the forward output port 120f through the discharge port EX to the atmospheric pressure, for example, so that the oil passage of the manual valve 120 is switched, the reverse brake B1 is engaged, and the forward clutch C1 is released. Be made.

  When the shift lever 74 is operated to the “P” position and the “N” position, the oil path from the input port 120a to the forward output port 120f and the oil path from the input port 120a to the reverse output port 120r are both And the oil passage of the manual valve 120 is switched so that the hydraulic oil in the forward clutch C1 and the reverse brake B1 is drained from the manual valve 120, and both the forward clutch C1 and the reverse brake B1 are connected. Be released.

The speed ratio control valve UP114 is closed the suppliable and output port 114k from the input port 114i to the line pressure P _L by being movable in the axial direction to the input side variable pulley 42 via the input and output ports 114j A spool valve element 114a positioned at an upshift position and an original position where the input-side variable pulley 42 communicates with the input / output port 114k via the input / output port 114j, and the spool valve element 114a toward the original position side. A spring 114b as an urging means for _urging , an oil chamber 114c that accommodates the spring 114b and receives the control hydraulic pressure _PDS2 to _apply a thrust toward the original position to the spool valve element 114a, and the spool valve element 114a control pressure P to apply a thrust force toward the upshift position side in And an oil chamber 114d that accepts _S1.

Further, the transmission ratio control valve DN116 is provided so as to be movable in the axial direction, whereby a downshift position where the input / output port 116j communicates with the discharge port EX and an original position where the input / output port 116j communicates with the input / output port 116k. A spool valve element 116a positioned at the first position, a spring 116b as an urging means for urging the spool valve element 116a toward the original position, and a spring 116b that accommodates the spool valve element 116a in the original position side. An oil chamber 116c that receives the control hydraulic pressure _PDS1 to apply a thrust toward the engine, and an oil chamber 116d that receives the control hydraulic pressure _PDS2 to apply a thrust toward the downshift position to the spool valve element 116a. .

In the transmission ratio control valve UP114 and the transmission ratio control valve DN116 thus configured, in the closed state in which the spool valve element 114a is held in the original position in accordance with the urging force of the spring 114b as shown in the left half of the center line, The input / output port 114j and the input / output port 114k are communicated with each other, and the hydraulic oil in the input side variable pulley 42 (input side hydraulic cylinder 42c) is allowed to flow to the input / output port 116j. In the closed state in which the spool valve element 116a is held in the original position according to the urging force of the spring 116b as shown in the right half of the center line, the input / output port 116j and the input / output port 116k are communicated with each other, and the thrust ratio it is allowed that the thrust ratio control oil pressure P _tau from control valve 118 to flow to the input-output port 114k.

Further, when the control oil pressure _PDS1 is supplied to the oil chamber 114d, the spool valve element 114a is increased against the urging force of the spring 114b by a thrust according to the control oil pressure _PDS1 as shown in the right half of the center line. is moved to the shift position side, the line pressure _{P L} is supplied to the control oil pressure _{P DS1} to the corresponding flow rate at the input port output from 114i port 114j menstrual the input side hydraulic cylinder 42c, input and output ports 114k is blocked Accordingly, the flow of the hydraulic oil to the speed ratio control valve DN116 side is prevented. As a result, the shift control pressure Pin is increased, the V groove width of the input side variable pulley 42 is narrowed, and the speed ratio γ is decreased, that is, the continuously variable transmission 18 is upshifted.

When the control oil pressure _PDS2 is supplied to the oil chamber 116d, the spool valve element 116a is lowered against the urging force of the spring 116b by the thrust according to the control oil pressure _PDS2 , as shown in the left half of the center line. The hydraulic fluid in the input side hydraulic cylinder 42c is discharged from the input / output port 114j through the input / output port 114k and the input / output port 116j through the input / output port 116j at a flow rate corresponding to the control hydraulic pressure _PDS2 . As a result, the shift control pressure Pin is lowered, the V-groove width of the input side variable pulley 42 is increased, and the speed ratio γ is increased, that is, the continuously variable transmission 18 is downshifted.

Thus, the line pressure P _L is a used as the basic pressure of the shift control pressure Pin, the control pressure P _DS1 is to be output speed ratio control line pressure P _L input to the valve UP114 input side hydraulic cylinder When the control hydraulic pressure _PS2 is output, the hydraulic oil in the input side hydraulic cylinder 42c is discharged from the discharge port EX and the shift control pressure Pin is reduced. Lowered and continuously downshifted.

For example, as shown in FIG. 4, it is actually determined from a previously stored relationship (shift map) between the vehicle speed V and the target input shaft rotational speed N _IN ^* that is the target input rotational speed of the continuously variable transmission 18 using the accelerator opening Acc as a parameter. The target input shaft rotational speed N _IN ^* set based on the vehicle state indicated by the vehicle speed V and the accelerator opening Acc matches the actual input shaft rotational speed (hereinafter referred to as the actual input shaft rotational speed) N _IN. As described above, the shift of the continuously variable transmission 18 is executed by feedback control in accordance with the rotational speed difference (deviation) ΔN _IN (= N _IN ^* −N _IN ), that is, hydraulic oil for the input side hydraulic cylinder 42c. By supplying and discharging, the V-groove widths of both variable pulleys 42 and 46 are changed, and the speed ratio γ is continuously changed by feedback control.

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} ^* is the target speed ratio ^{_{γ * (= N IN * /}} N OUT) Is determined within the range of the minimum speed ratio γmin and the maximum speed ratio γmax of the continuously variable transmission 18.

Further, the control hydraulic pressure _PDS1 is supplied to the oil chamber 116c of the transmission ratio control valve DN116, and regardless of the control hydraulic pressure _PDS2 , the transmission ratio control valve DN116 is closed to limit the downshift, while the control hydraulic pressure _{PDS2 changes} the speed. The oil ratio is supplied to the oil chamber 114c of the ratio control valve UP114, and regardless of the control oil pressure _PDS1 , the transmission ratio control valve UP114 is closed to prohibit the upshift. That is, the control when the hydraulic _{P DS1} and the control pressure _{P DS2} are not supplied together but of course, also, the speed change ratio control valve UP114 and speed ratio control valve when the control oil pressure _{P DS1} and the control pressure _{P DS2} is supplied together Each of the DNs 116 is in a closed state held in its original position. As a result, one of the solenoid valves DS1 and DS2 does not function due to a failure in the electrical system, and a sudden upshift occurs even when the control hydraulic pressure _PDS1 or the control hydraulic pressure _PDS2 continues to be output at the maximum pressure. It is possible to prevent a downshift or a belt slip due to the sudden shift.

The clamping force control valve 110, via an output port 110t to line pressure P _L by opening and closing an input port 110i from the input port 110i output side variable pulley 46 and the thrust ratio control by being movable in the axial direction valve 118 The spool valve element 110a that enables the belt clamping pressure Pd to be supplied, the spring 110b as an urging means that urges the spool valve element 110a in the valve opening direction, and the spring 110b is accommodated in the spool valve element 110a. An oil chamber 110c that receives the control hydraulic pressure _PSLS to give thrust in the valve opening direction, and feedback that receives belt clamping pressure Pd output from the output port 110t to give thrust in the valve closing direction to the spool valve element 110a. The thrust in the valve closing direction is applied to the oil chamber 110d and the spool valve element 110a. And an oil chamber 110e that accepts modulator pressure P _M in order to impart.

In the clamping pressure control valve 110 thus configured, by the transmission belt 48 is line pressure P _L is continuously regulated pressure control control oil pressure P _SLS so as not slip as a pilot pressure, an output port 110t From this, the belt clamping pressure Pd is output. Thus, the line pressure P _L is used as the basic pressure of the belt clamping pressure Pd. A hydraulic pressure sensor 130 is provided in the oil passage between the output port 110t and the output side hydraulic cylinder 46c, and the belt clamping pressure Pd is detected by the hydraulic pressure sensor 130.

For example, as shown in FIG. 5, a relationship (belt) that is experimentally obtained in advance and stored so that belt slip does not occur between the transmission gear ratio γ and the belt clamping pressure Pd ^* using the accelerator opening Acc corresponding to the transmission torque as a parameter. The belt clamping pressure Pd of the output side hydraulic cylinder 46c is obtained so that the belt clamping pressure Pd ^* determined (calculated) based on the vehicle state indicated by the actual gear ratio γ and the accelerator opening Acc is obtained from the clamping pressure map). The pressure is regulated, and the belt clamping pressure Pd ^*, that is, the frictional force between the variable pulleys 42 and 46 and the transmission belt 48 is increased or decreased according to the belt clamping pressure Pd.

The thrust ratio control valve 118, a thrust ratio control oil pressure P the line pressure _{P L} by opening and closing an input port 118i by being movable in the axial direction from the input port 118i via an output port 118t to the speed ratio control valve DN116 a spool valve element 118a that can supply _τ , a spring 118b as an urging means that urges the spool valve element 118a in the valve opening direction, and the spring 118b is accommodated in the valve opening direction in the valve opening direction. an oil chamber 118c that receives the belt clamping pressure Pd to apply a thrust force, a feedback oil chamber for receiving the thrust ratio control oil pressure P _tau output from the output port 118t to apply a thrust force in the valve closing direction to the spool valve element 118a 118d.

In the thrust ratio control valve 118 configured as described above, the pressure receiving area of the belt clamping pressure Pd in the oil chamber 118c a, the pressure receiving area of the thrust ratio control oil pressure P _tau in the feedback oil chamber 118d b, the biasing force of the spring 118b When F _S, an equilibrium state in the following equation (3).
_Pτ × b = Pd × a + F _S (3)
Accordingly, the thrust ratio control hydraulic pressure _Pτ is expressed by the following equation (4) and is proportional to the belt clamping pressure Pd.
_Pτ = Pd × (a / b) + F _S / b (4)

Then, the control oil pressure _P or _DS1 and the control pressure _{P DS2} is not supplied together, or the predetermined pressure or more control pressure _{P DS1} and the predetermined pressure or more control pressure _{P DS2} is both supplied, the speed ratio control valve UP114 and the speed ratio control when the valve DN116 has both been a closed state is held in the original position, since the thrust ratio control oil pressure P _tau is supplied to the input side hydraulic cylinder 42c, and a shift control pressure Pin is the thrust ratio control oil pressure P _tau Matched. That is, the thrust ratio control oil pressure P _tau i.e. the shift control pressure Pin maintain a predetermined relationship between the ratio between the shift control pressure Pin and the belt clamping pressure Pd by the thrust ratio control valve 118 is output.

For example, because of the accuracy of the input shaft rotation speed sensor 56 and the vehicle speed sensor 58, the detection accuracy of the input shaft rotation speed _NIN and the vehicle speed V is inferior in a low vehicle speed state below a predetermined vehicle speed V ′. Instead of feedback control of the gear ratio γ for eliminating the rotational speed difference (deviation) ΔN _IN at the time of starting, for example, neither the control hydraulic pressure P _DS1 nor the control hydraulic pressure P _DS2 is supplied, and the gear ratio control valve UP114 and the gear ratio control valve A so-called closing control is performed in which all the DNs 116 are closed. As a result, the shift control pressure Pin proportional to the belt clamping pressure Pd is applied to the input side hydraulic cylinder 42c so that the ratio between the transmission control pressure Pin and the belt clamping pressure Pd is a predetermined relationship during low vehicle speed traveling or starting. The belt slippage of the transmission belt 48 from the time when the vehicle is stopped to the time of extremely low vehicle speed is prevented, and at this time, for example, the thrust ratio τ corresponding to the maximum gear ratio γmax (= output side hydraulic cylinder thrust W _OUT / input side) The hydraulic cylinder thrust W _IN ; W _OUT is the belt clamping pressure Pd × the pressure receiving area S ₄₆ of the output side hydraulic cylinder ₄₆ c, and W _IN is the shift control pressure Pin × the pressure receiving area S ₄₂ of the input side hydraulic cylinder 42 c). If (a / b) or F _S / b in the first term on the right side of the above equation (4) is set as possible, a good start can be made at the maximum gear ratio γmax or a gear ratio γmax 'in the vicinity thereof. Done The predetermined vehicle speed V ′ is a lower limit vehicle speed at which a predetermined feedback control can be executed as a vehicle speed V at which the rotational speed of the predetermined rotating member, for example, the input shaft rotational speed _NIN cannot be detected. For example, it is set to about 2 km / h.

FIG. 6 shows a relationship that is obtained and stored in advance between the speed ratio γ and the thrust ratio τ with the vehicle speed V as a parameter, and the relationship (a) of the first term on the right side of the above equation (4) is as shown in the figure. It is a figure which shows an example when / b) is set. The parameter of the vehicle speed V indicated by the one-dot chain line in FIG. 6 is a thrust ratio τ calculated in consideration of the centrifugal hydraulic pressure in the input side hydraulic cylinder 42c and the output side hydraulic cylinder 46c, and the intersections with the solid lines (V ₀ , V ₂₀ , V ₅₀ ) obtains a gear ratio γ as a predetermined gear ratio that can be held during the closing control. For example, as shown in FIG. 6, in the continuously variable transmission 18 according to this embodiment, the maximum speed ratio γmax can be maintained as a predetermined speed ratio when the vehicle speed V is 0 km / h, that is, the closing control is performed while the vehicle is stopped. .

FIG. 7 is a functional block diagram for explaining the main part of the control function by the electronic control unit 50. In FIG. 7, the target input rotation setting means 150 is configured to change the input shaft rotation speed N _IN based on the vehicle state indicated by the actual vehicle speed V and the accelerator opening Acc from a shift map stored in advance as shown in FIG. Set the target input shaft speed N _IN ^* .

The speed change control means 152 determines the rotational speed difference ΔN _IN (= N _IN ^* −) so that the actual input shaft rotational speed N _IN matches the target input shaft rotational speed N _IN ^* set by the target input rotation setting means 150. N _IN ), the shift of the continuously variable transmission 18 is feedback-executed. That is, the input side hydraulic shift control command signal for changing the V groove widths of both variable pulleys 42 and 46 by controlling the flow of hydraulic fluid to the cylinder 42c (hydraulic pressure command) is output to S _T to the hydraulic control circuit 100 shift The ratio γ is continuously changed.

The belt clamping pressure setting means 154, for example, from the belt clamping pressure map obtained and stored experimentally in advance as shown in FIG. 5, for example, the actual accelerator opening Acc and the actual input shaft rotational speed N by the electronic control unit 50. _The belt clamping pressure Pd ^* is set based on the vehicle state indicated by the actual speed ratio γ (= N _IN / N _OUT ) calculated based on _IN and the output shaft rotational speed N _OUT . That is, the belt clamping pressure setting means 154 sets the belt clamping pressure Pd of the output side hydraulic cylinder 46c for obtaining the belt clamping pressure Pd ^* .

Belt clamping pressure control means 156, the belt clamping pressure setting means 154 clamping pressure control command signal S _B for pressurizing regulating the belt clamping pressure Pd of the output side hydraulic cylinder 46c for the set belt clamping pressure Pd ^* is obtained by Output to the hydraulic control circuit 100 to increase or decrease the belt clamping pressure Pd ^* .

The hydraulic control circuit 100, the supply of hydraulic fluid by operating the solenoid valve DS1 and the solenoid valve DS2 so shifting of the continuously variable transmission 18 is executed to the input side hydraulic cylinder 42c in accordance with the shift control command signal S _T · to control the emissions, by operating the linear solenoid valve SLS so that the belt clamping pressure Pd ^* is increased or decreased pressure of the belt clamping pressure Pd adjusted in accordance with the above clamping force control command signal S _B.

The engine output control means 158 outputs an engine output control command signal S _E , for example, a throttle signal, an injection signal, an ignition timing signal, etc., to the throttle actuator 76, the fuel injection device 78, and the ignition device 80, respectively, for output control of the engine 12. To do. For example, the engine output control means 158 controls the engine torque T _E and outputs a throttle signal for opening and closing the electronic throttle valve 30 such that the throttle opening theta _TH corresponding to the accelerator opening Acc to the throttle actuator 76.

  The engine output control means 158 functions as a fuel cut control means for shutting off fuel to the engine 12 when a predetermined condition is satisfied in order to improve fuel consumption.

Specifically, the engine output control means 158 is configured to perform the engine deceleration when the vehicle is decelerating with the throttle off such that the throttle valve opening θ _TH is substantially fully closed or is slightly opened below about 2-3%. rotational speed N _E, for example, such as the fuel cut start rotational speed predetermined to approximately 1400 rpm N _EK traversal and engine rotational speed N _E has been elapsed traveling state becomes a predetermined delay time from T _FC that towards the reduction when a predetermined condition is satisfied, outputs a stop instruction S _FC of the fuel supply to the engine 12 as fuel cut operation is started in the fuel injection device 78.

Then, the engine output control means 158, the fuel supply stop command S to the engine 12 as fuel cut operation is terminated when the engine rotational speed N _E with deceleration is equal to or less than a predetermined fuel cut speed N _EF Stop _FC output. When the fuel cut operation is released, the fuel supply is resumed and the engine 12 is started quickly. The fuel cut return rotational speed N _EF is an engine rotational speed determination value that is experimentally obtained and stored in advance so that the engine 12 is quickly started when the fuel supply is resumed, and is, for example, about 550 rpm. idle rotation speed _{N IDL} or less 300~500rpm about the engine rotational speed _{N E} is set.

The rotational speed difference between the predetermined delay time T _FC and the fuel cut start rotational speed N _EK and the fuel cut return rotational speed N _EF is a judgment condition for providing hysteresis for stabilizing the start / end judgment of the fuel cut control. is there. In other words, a determination condition is provided to prevent the switching busy fuel cut operation due to busy switching and engine speed N _E changes in the fuel cut operation by the accelerator on-off.

By the way, in order to generate a shift control pressure Pin (hereinafter referred to as Pin pressure) and a belt clamping pressure Pd ^* which are necessary when the speed ratio γ is feedback controlled in order to improve the speed change performance, for example, the speed change response. the belt clamping pressure Pd (hereinafter, Pd referred pressure) to protect itself, it is necessary to set the line pressure P _L as the original pressure thereof Pin pressure and the Pd pressure.

In other words, the line pressure P _L is relatively high and even less likely to occur often belt slip shift response than Pin pressure and the Pd pressure, causes the fuel efficiency is deteriorated unnecessarily high. Also, or a factor of the line pressure P _L is reduced as low as shifting response than Pin pressure, the line pressure P _L is easily made factors occur less belt slippage than the Pd pressure.

Accordingly, the line hydraulic pressure setting means 160 is controlled by the linear solenoid valve SLT independently of the flow rate control of the hydraulic oil to the input side hydraulic cylinder 42c by the solenoid valves DS1 and DS2 and the control of the Pd pressure by the linear solenoid valve SLS. that the line pressure P _L, is set based on either the higher oil pressure of the Pin pressure and the Pd pressure.

For example, the line oil pressure setting means 160, based on whichever is higher hydraulic pressure of the hydraulic pressure by adding a predetermined margin value in consideration of the control accuracy and the vehicle state of the line pressure P _L in each of the Pin pressure and the Pd pressure line to set the hydraulic pressure _{P L.}

Figure 8 is a diagram for explaining an example of a concept of setting the line hydraulic pressure P _L. As shown in FIG. 8, the reference margin value EX that is experimentally determined in advance as the predetermined margin value is the Pin pressure reference margin value EXin at the Pin pressure, and the Pd pressure reference margin value EXd at the Pd pressure. It is. Thus, different values are set for the reference margin value EX for the Pin pressure and the Pd pressure. Then, the line oil pressure P _L that requires the higher oil pressure of the oil pressure obtained by adding the Pin pressure reference margin value EXin to the Pin pressure and the oil pressure obtained by adding the Pd pressure reference margin value EXd to the Pd pressure, that is, the target line oil pressure ( Hereinafter, it is set as the target line oil pressure) P _L ^* .

The setting of the target line oil pressure P _L ^* will be described in detail below.

Pd pressure is directly regulated pressure control by operating the linear solenoid valve SLS accordance clamping pressure control command signal S _B in accordance with the belt clamping pressure control means 156 such that the Pd pressure for the belt clamping pressure Pd ^* is obtained It is hydraulic. Therefore, as the Pd pressure used for setting the target line oil pressure P _L ^* , the Pd pressure for obtaining the belt clamping pressure Pd ^* set by the belt clamping pressure setting means 154 is used as it is.

On the other hand, the Pin pressure is a hydraulic pressure generated as a result of the flow control of the hydraulic oil and the belt clamping pressure control when feedback control is performed so that the target input shaft rotational speed N _IN ^* (or the target speed ratio γ ^* ) is obtained. Therefore, pressure regulation is not directly controlled. Therefore, the Pin pressure used for setting the target line oil pressure P _L ^* is calculated as an estimated value.

As described above, when the thrust of the output side hydraulic cylinder 46c is W _OUT and the thrust of the input side hydraulic cylinder 42c is W _IN , the thrust ratio τ is expressed by the following equation (5):
τ = W _OUT / W _IN (5)
Assuming that the pressure receiving area of the output side hydraulic cylinder 46c is S ₄₆ and the pressure receiving area of the input side hydraulic cylinder 42c is S ₄₂ , the output side hydraulic cylinder thrust W _OUT and the input side hydraulic cylinder thrust W _IN are respectively expressed by the following equations (6), (7)
W _OUT = Pd × S ₄₆ (6)
W _IN = Pin × S ₄₂ (7)
From the equations (5) to (7), the Pin pressure is expressed by the following equation (8).
Pin = (Pd × S ₄₆ ) / (τ × S ₄₂ ) (8)

The input side estimated hydraulic pressure calculating means 162 calculates the thrust ratio τ, the Pd pressure set by the belt clamping pressure setting means 154, and the pressure receiving area based on S ₄₆ and S ₄₂ according to the above equation (8). An estimated value (hereinafter referred to as estimated Pin pressure) is calculated.

  The thrust ratio calculation means 164 uses the vehicle speed V as shown in FIG. 6 as a parameter, and from the relationship (thrust ratio characteristic, map) obtained and stored in advance between the speed ratio γ and the thrust ratio τ, the actual vehicle speed V and The thrust ratio τ is calculated based on the actual speed ratio γ.

As described above, the Pd pressure is a hydraulic pressure that is directly pressure-controlled so as to be a Pd pressure for obtaining the belt clamping pressure Pd ^* . Thus, the line oil pressure setting means 160, the only consideration of variations of the actual line oil pressure P _L to the line oil pressure control command signal S _PL, i.e. the actual line oil pressure P _L to the line pressure control command signal S _PL its actual line pressure P _L even variations in to exceed the Pd pressure, set the Pd pressure reference margin value EXd stored predetermined experimentally.

On the other hand, as described above, the Pin pressure is not directly regulated and is a hydraulic pressure calculated as an estimated value. In addition, a certain margin is required to perform feedback control so as to maintain a constant gear ratio γ. Thus, the line oil pressure setting means 160, the line oil pressure control command signal S _PL actual line pressure P _L of only to be larger than the Pd pressure reference margin value EXd that has been set in consideration variation and, in advance experimentally The Pin pressure reference margin value EXin determined and stored in is set. Thus, different reference margin values EX are set for the Pin pressure and the Pd pressure, respectively.

  Further, the line hydraulic pressure setting means 160 may change the Pd pressure reference margin value EXd and the Pin pressure reference margin value EXin, respectively, based on the vehicle state (running state).

For example, the line oil pressure setting means 160 is configured to change the actual line oil pressure P _L with respect to the line oil pressure control command signal S _PL when the hydraulic oil temperature is low and when the hydraulic oil temperature is low compared to the normal oil temperature when the hydraulic oil temperature is normal. Since the control accuracy is lowered, the Pd pressure reference margin value EXd and the Pin pressure reference margin value EXin which are larger than those at the normal oil temperature are set. The normal oil temperature is assumed to be, for example, the oil temperature after completion of warm-up.

  Further, the line hydraulic pressure setting means 160 changes the speed ratio γ during steady running so that a constant speed ratio γ is maintained, and the change in the speed ratio γ is extremely small compared to the speed change that changes the speed ratio γ. Therefore, the Pd pressure reference margin value EXd and the Pin pressure reference margin value EXin, which are smaller than those at the time of shifting, are set.

Further, the line hydraulic pressure setting means 160 greatly changes the gear ratio γ during traveling that requires a sudden gear change, that is, during a sudden gear change in which the gear ratio γ exceeds a predetermined value and the gear ratio γ is changed. Therefore, a large Pd pressure reference margin value EXd and a Pin pressure reference margin value EXin are set as compared with those during normal gear shifting and steady running. This normal shift is a shift in which the speed ratio γ is changed within a range in which the speed of change of the speed ratio γ does not exceed a predetermined value. During a sudden shift, the accelerator pedal 68 suddenly returns as indicated by an arrow A in FIG. When the vehicle is operated to perform an off-up shift, which is a rapid acceleration, or when the speed ratio γ is changed along the lower limit value of the target input shaft rotational speed N _IN ^* as shown by the arrow B, the rapid acceleration It is assumed that the vehicle will be running.

Then, the line oil pressure setting means 160, the oil pressure value obtained by adding the Pd pressure reference margin value EXd to the Pd pressure set by the belt clamping pressure setting means 154 calculates the Pd pressure-purpose line oil pressure P _{L d,} the input A hydraulic pressure value obtained by adding the Pin pressure reference margin value EXin to the estimated Pin pressure calculated by the side hydraulic pressure estimated value calculating means 162 is calculated as a Pin pressure line hydraulic pressure P _L in, and the Pd pressure line hydraulic pressure P _L d and the Pin pressure the line hydraulic pressure P _L in alternatively select either the higher line hydraulic pressure of the, to set the selected higher line hydraulic pressure as the target line pressure P _L ^*. As shown in FIG. 8, the target line oil pressure P _L ^* is set with the line oil pressure MAX guard that has been experimentally determined in advance so that the line oil pressure P _L does not exceed the allowable load of the transmission belt 48 as an upper limit. The This line oil pressure MAX guard may be determined in consideration of fuel consumption instead of or in addition to the allowable load of the transmission belt 48.

Line pressure control means 166, for example, pre-stored relationship shown in FIG. 9 and the line oil pressure control command signal S _PL to the line pressure P _L which is pressure regulated on the basis of the line pressure control command signal S _PL (line Line pressure control command signal S _PL (control current) for obtaining the target line oil pressure P _L ^* based on the target line oil pressure P _L ^* set by the line oil pressure setting means 160 from the hydraulic pressure characteristic). output to 100 to pressure the line hydraulic pressure _{P L} tone is.

The hydraulic control circuit 100 actuates the linear solenoid valve SLT pressure to line pressure P _L regulated in accordance with the above line oil pressure control command signal S _PL.

Line pressure characteristic of FIG. 9, the line hydraulic pressure control command signal S _PL and the line oil pressure control command signal S relationship between the control hydraulic pressure P _SLT is the output oil pressure of the linear solenoid valve SLT normally open type which is driven based on the _PL Table in relation to the relationship between the line oil pressure _{P L} to be pressure regulated by the primary regulator valve 122, and line hydraulic pressure control command signal _{S PL} and the line pressure _{P L} based, and control pressure _{P SLT} and its control oil pressure _{P SLT} The hydraulic characteristics are experimentally obtained in advance.

  Here, the thrust ratio τ differs between when the vehicle is in a driven state and when it is in a driven state. That is, the actual Pin pressure (hereinafter referred to as the actual Pin pressure) is lower when in the driven state than when in the driven state, and the thrust ratio characteristic as shown in FIG. Is different. Therefore, it is necessary to calculate the estimated Pin pressure by obtaining the thrust ratio τ from different thrust ratio characteristics in consideration of whether the vehicle is in a driving state or a driven state.

  The storage unit 168 calculates a thrust ratio characteristic during driving (hereinafter referred to as a driving map) for determining the thrust ratio τ when the vehicle is in a driving state, and a thrust ratio τ when the vehicle is in a driven state. The driven thrust ratio characteristic for driving (hereinafter referred to as a driven map) is stored.

  The thrust ratio calculation means 164 calculates the thrust ratio τ based on the actual vehicle speed V and the actual speed ratio γ from the driving time map when the vehicle is in a driving state, while being driven when the vehicle is in a driven state. A thrust ratio τ is calculated based on the actual vehicle speed V and the actual speed ratio γ from the time map.

However, when the driving state is switched from the driving state to the driven state, when the driving determination flag is turned on, the driving time map is switched to the driving time map and the thrust ratio τ is calculated from the driving time map. Since there is a response delay in the reduction from the actual Pin pressure in the driving state (hereinafter referred to as the actual Pin pressure during driving) to the actual Pin pressure in the driven state (hereinafter referred to as the actual Pin pressure during driving), the actual Pin pressure The estimated Pin pressure is calculated based on the thrust ratio τ obtained from the driven map before the actual driven Pin pressure drops to the driven Pin pressure (hereinafter referred to as the driven Pin estimated pin pressure). In the process from the actual driving Pin pressure to the driven actual Pin pressure, the estimated Pin pressure becomes lower than the actual Pin pressure, and the controllability of the continuously variable transmission 18 may be reduced. For example, the estimated Pin pressure is lower than the actual control pressure Pin, which may be insufficient line pressure P _L required to relatively small target line pressure P _L ^* is set Pin pressure. The driven determination flag may be, for example, one of the predetermined conditions for the fuel cut operation, that is, the throttle valve opening θ _TH is substantially fully closed, or is slightly opened with about 2 to 3% or less. Is turned on by the electronic control unit 50 when the condition that the vehicle is traveling at a reduced throttle speed while the throttle is off is satisfied.

Therefore, when the fuel cut operation is started on the condition that a predetermined delay time T _{FC has} elapsed since the vehicle is decelerated while the throttle is off, the driving map is switched to the driven map, and the thrust ratio is calculated from the driven map. τ is calculated. As a result, when the actual pin pressure at the time of driving is reliably reduced, the estimated Pin pressure at the time of driving is calculated as the estimated Pin pressure.

More specifically, a fuel cut state determining means 170 determines whether or not the engine 12 is in a fuel cut state based on the stop command S _FC output from the engine output control means 158. Further, a fuel cut state determining means 170, when it is determined that the engine 12 is in a fuel cut state by being output stop command S _FC, the fuel cut flag ON.

  The thrust ratio calculating means 164 does not calculate the thrust ratio τ depending on whether the vehicle is in the driven state or in the driven state when the vehicle is switched from the driving state to the driven state. When it is determined by the cut state determination means 170 that the engine 12 is not in the fuel cut state, the thrust ratio τ is calculated based on the actual vehicle speed V and the actual speed ratio γ from the driving time map, while the fuel cut state determination means 170 Thus, when it is determined that the engine 12 is in the fuel cut state, the thrust ratio τ is calculated based on the actual vehicle speed V and the actual speed ratio γ from the driven map.

  That is, the thrust ratio calculation means 164 switches from the driving time map to the driven time map when the driven determination flag is turned on when the vehicle is switched from the driving state to the driven state, and from the driven time map. Instead of calculating the thrust ratio τ, when the fuel cut flag is turned on, the driving time map is switched to the driven time map, and the thrust ratio τ is calculated from the driven time map.

  FIG. 10 calculates the estimated Pin pressure by obtaining the thrust ratio τ from different thrust ratio characteristics in consideration of the main part of the control operation of the electronic control unit 50, that is, whether the vehicle is in the driven state or the driven state. For example, the control operation is repeatedly executed with a very short cycle time of about several milliseconds to several tens of milliseconds. FIG. 11 is a time chart for explaining the control operation shown in the flowchart of FIG.

In FIG 10, first, the step corresponding to the fuel cut state determining means 170 (hereinafter, omitting step) in S1, based on the stop command S _FC of the fuel supply to the engine 12 is output to the fuel injection device 78 It is determined whether or not the engine 12 is fuel cut. When the stop command S _FC is output, fuel cut flag is turned on from off.

Time point t ₁ in FIG. 11, the vehicle in the driving state is shown that is a driven state with the driven determination flag is turned on. As shown after time t _{1 in} FIG. 11, the change from the actual driving pressure to the actual driving pressure has a response delay with respect to switching from the driving state to the driven state of the vehicle.

Next, in S2 corresponding to the thrust ratio calculating means 164, the thrust ratio τ is calculated based on the actual vehicle speed V and the actual speed ratio γ from the thrust ratio characteristics as shown in FIG. More specifically, in S21 if the result is negative judgment of step S1 does not output the stop command S _FC, the thrust ratio τ is calculated based on the actual vehicle speed V and the actual gear ratio γ from a drive time map Is done. On the other hand, in S22, if the result is affirmative judgment of said output stop command S _FC S1, the thrust ratio τ is calculated based on the actual vehicle speed V and the actual gear ratio γ from the drive-time map.

Then, in S3 corresponding to the input estimated hydraulic pressure calculation means 162, Pin = according _{(Pd × S 46) / (} τ × S 42), the thrust ratio calculated by the S2 tau, the belt clamping pressure setting The estimated Pin pressure is calculated based on the Pd pressure set by the means 154 and the pressure receiving area based on S ₄₆ and S ₄₂ .

T ₂ time points 11 show that the fuel cut flag has been turned on from off. The one-dot chain line in FIG. 11 is a case where the driven-time estimated Pin pressure is calculated by switching to the driven-time map when the fuel cut flag is turned on, and the broken line is when the driven determination flag is turned on. In this case, the drive-time estimated Pin pressure is calculated by switching to the drive-time map. In the broken line, since the estimated Pin pressure is set to the estimated Pin pressure at the time of driving before being lowered to the actual Pin pressure at the time of driving, the estimated Pin pressure is higher than the actual Pin pressure until it is reduced to the actual Pin pressure at the time of driving. It is low. On the other hand, in the alternate long and short dash line in this embodiment that is switched to the driven map by the fuel cut flag, the estimated Pin pressure is calculated based on the thrust ratio τ obtained from the driving map until the fuel cut flag is turned on. Since the Pin pressure (hereinafter referred to as the estimated Pin pressure during driving) is used, it is avoided that the estimated Pin pressure is lower than the actual Pin pressure.

  As described above, according to the present embodiment, when the thrust ratio calculating means 164 determines that the engine 12 is not in the fuel cut state by the fuel cut state determining means 170, the actual vehicle speed V and the actual speed change are determined from the driving time map. While the thrust ratio τ is calculated based on the ratio γ, when the fuel cut state determination means 170 determines that the engine 12 is in the fuel cut state, the actual vehicle speed V and the actual speed ratio γ are determined from the driven map. Since the thrust ratio τ is calculated based on the map, a map for determining the thrust ratio τ based on the fuel cut of the engine 12 executed at a timing delayed from the timing of switching from the driving state of the vehicle to the driven state ( (Thrust ratio characteristics) will be switched, and the timing of switching from the driven state to the driven state The map is switched at the timing when the actual Pin pressure with a delay in response actually decreases. Therefore, the estimated Pin pressure calculated based on the thrust ratio τ is avoided from being lower than the actual Pin pressure, and the controllability when the continuously variable transmission 18 is controlled based on the estimated Pin pressure is improved. .

Further, according to this embodiment, since the line pressure P _L is controlled based on the estimated Pin pressure by the line oil pressure control unit 166, it is avoided insufficient line pressure P _L required Pin pressure.

  Next, another embodiment of the present invention will be described. In the following description, parts common to the embodiments are denoted by the same reference numerals and description thereof is omitted.

  In the above-described embodiment, when the fuel cut flag is turned on when the vehicle is switched from the driving state to the driven state, the driving ratio map is switched to the driving time map to calculate the input side hydraulic pressure estimated value. Although the calculating means 162 calculates the estimated Pin pressure based on the thrust ratio τ, in this embodiment, when the vehicle is switched from the driving state to the driven state, the input side hydraulic pressure estimated value calculating means 162 is driven. From the time when the determination flag is turned on, the estimated Pin pressure is gradually decreased from the estimated Pin pressure during driving toward the estimated Pin pressure during driving with a predetermined gradient. Even in this case, it is avoided that the estimated Pin pressure becomes lower than the actual Pin pressure as in the above-described embodiment.

  FIG. 12 is a time chart showing an embodiment in which the estimated Pin pressure is gradually decreased from the driving estimated Pin pressure toward the driven estimated Pin pressure.

Time point t ₁ in FIG. 12, the vehicle in the driving state is shown that is a driven state with the driven determination flag is turned on. T ₂ point of FIG. 12 shows that the fuel cut flag has been turned on from off. One-dot chain line in FIG. 12 is an estimated Pin pressure from driving time estimated Pin pressure toward the driven time estimated Pin pressure was allowed to gradually reduced from time point t ₁ at a predetermined inclination. Thus, it is avoided that the estimated Pin pressure becomes lower than the actual Pin pressure in the one-dot chain line of the present embodiment.

Further, t ₁ estimated Pin pressure from the time the time and up to the time that is a driven time estimated Pin pressure, the time point t ₁ t time from time point t ₁ to suppress the time difference ΔT predetermined to ₂ times The estimation accuracy of the estimated Pin pressure may be further improved by learning the inclination.

  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 Pd pressure set by the belt clamping pressure setting unit 154 is used as the Pd pressure in the setting of the target line oil pressure P _L ^* and the calculation of the estimated Pin pressure. The belt clamping pressure Pd to be used may be used. When the belt clamping pressure Pd detected by the hydraulic sensor 130 is not used as the Pd pressure, the hydraulic sensor 130 is not necessarily provided.

Further, the input shaft rotational speed N _IN and the related target input shaft rotational speed N _IN ^* in the above-described embodiment are replaced with the input rotational speed N _{IN and the} like, and the engine rotational speed _NE and the related target. The engine rotational speed N _E ^* or the like, or the turbine rotational speed _NT or a target turbine rotational speed _NT ^* related thereto may be used.

  In the above-described embodiment, the torque converter 14 provided with the lock-up clutch 26 is used as the fluid transmission device. However, the lock-up clutch 26 is not necessarily provided. In addition, other fluid type power transmission devices such as a fluid coupling (fluid coupling) having no torque amplification function may be used.

  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.

1 is a skeleton diagram illustrating a vehicle drive device to which the present invention is applied. It is a block diagram explaining the principal part of the control system provided in the vehicle in order to control the vehicle drive device etc. of FIG. FIG. 3 is a hydraulic circuit diagram showing a main part relating to belt clamping pressure control of a continuously variable transmission, gear ratio control, and engagement hydraulic control of a forward clutch or reverse brake accompanying operation of a shift lever in the hydraulic control circuit. It is a figure which shows an example of the shift map used when calculating | requiring a target input rotational speed in the shift control of a continuously variable transmission. It is a figure which shows an example of the belt clamping pressure map which calculates | requires belt clamping pressure according to gear ratio etc. in clamping pressure control of a continuously variable transmission. This is a relationship obtained and stored in advance between the gear ratio and the thrust ratio with the vehicle speed as a parameter. It is a functional block diagram explaining the principal part of the control function of the electronic control apparatus of FIG. It is a figure for demonstrating an example of the way of thinking which sets line oil pressure. It is a figure which shows an example of the relationship (line oil pressure characteristic) memorize | stored beforehand with a line oil pressure control command signal and the line oil pressure adjusted based on the line oil pressure control command signal. Control for calculating the estimated Pin pressure by obtaining the thrust ratio from different thrust ratio characteristics in consideration of the main part of the control operation of the electronic control unit of FIG. It is a flowchart explaining an action | operation. It is a time chart explaining the control action shown in the flowchart of FIG. FIG. 12 is a time chart showing an embodiment when the estimated Pin pressure is gradually decreased from the estimated Pin pressure during driving toward the estimated Pin pressure during driving, and is another embodiment of FIG. 11.

Explanation of symbols

12: engine 18: continuously variable transmission 24: drive wheel 42: input side variable pulley 42c: input side hydraulic cylinder 46: output side variable pulley 46c: output side hydraulic cylinder 48: transmission belt (belt)
50: Electronic control device (control device)
162: Input side hydraulic pressure estimated value calculating means 164: Thrust ratio calculating means 166: Line hydraulic pressure control means 168: Storage means 170: Fuel cut state determining means

Claims (2)

In a vehicle in which a continuously variable transmission having an input-side variable pulley, an output-side variable pulley, and a belt wound around both pulleys is disposed in a power transmission path between an engine and a drive wheel, the input-side variable A thrust ratio, which is a ratio of a thrust by the input side hydraulic cylinder for changing the groove width of the pulley and a thrust by the output side hydraulic cylinder for changing the groove width of the output side variable pulley, and the output side hydraulic cylinder Input side hydraulic pressure estimated value calculation means for calculating an estimated value of the hydraulic pressure in the input side hydraulic cylinder based on the hydraulic pressure of the continuously variable transmission, based on the estimated value of the hydraulic pressure in the input side hydraulic cylinder A control device for a continuously variable transmission for a vehicle for controlling
Stores a previously determined driving relationship for determining a thrust ratio when the vehicle is in a driving state and a previously determined driving relationship for determining a thrust ratio when the vehicle is in a driven state Storage means for
Fuel cut state determination means for determining whether or not the engine is in a fuel cut state;
When the fuel cut state determining means determines that the fuel cut state is not established, the thrust ratio is obtained from the driving relationship, while when the fuel cut state determining means determines that the fuel cut state is established, And a thrust ratio calculating means for determining the thrust ratio from a relationship during driving.   The control device for a continuously variable transmission for a vehicle according to claim 1, further comprising line hydraulic pressure control means for controlling the line hydraulic pressure of the continuously variable transmission based on an estimated value of the hydraulic pressure in the input hydraulic cylinder.

Images