JP7330628B2 - vehicle controller - Google Patents

Description

The present invention relates to a vehicle control device.

2. Description of the Related Art A conventional vehicle that runs by the power of an engine is equipped with a transmission. The power of the engine is input to the input shaft of the transmission, changed within the transmission, and transmitted from the output shaft of the transmission to the left and right drive wheels via a differential gear or the like.

A belt-type continuously variable transmission (CVT), which is a type of transmission, has a configuration in which an endless belt is wound around a primary pulley and a secondary pulley. By changing the winding diameter of the belt on the primary pulley and the secondary pulley, the pulley ratio continuously and steplessly changes, and the gear ratio changes as the pulley ratio changes.

Further, some belt-type continuously variable transmissions have a planetary gear mechanism for switching between forward and reverse travel of the vehicle. A planetary gear mechanism is provided, for example, between a secondary shaft that supports a secondary pulley and an output shaft. The secondary shaft is non-rotatably connected to the sun gear of the planetary gear mechanism, and the output shaft is non-rotatably connected to the ring gear of the planetary gear mechanism. When the vehicle moves forward, the carrier of the planetary gear mechanism is set in a free-rotating state, and the sun gear and the ring gear are directly connected. As a result, the power transmitted to the secondary shaft causes the sun gear and the ring gear to rotate together, and the output shaft to rotate together with the ring gear. On the other hand, when the vehicle is moving in reverse (reversing), the direct connection between the sun gear and the ring gear is released, and the carrier is non-rotatably fixed. As a result, when the sun gear rotates due to the power of the secondary shaft, the ring gear rotates in the opposite direction to the sun gear, and the output shaft rotates in the opposite direction to the forward direction.

When the vehicle is moving forward, the rotation speed of the sun gear and the ring gear are the same due to the direct connection between the sun gear and the ring gear. Always lower than speed. Therefore, if the maximum pulley ratio is configured in the reverse range (R range), the gear ratio becomes larger than the maximum pulley ratio, and the power output from the output shaft increases when the vehicle starts to move backward. Concerns such as jumping out due to creep and lack of braking force arise. In addition, when the reverse range is switched to the parking range (P range) and the parking pole is engaged with the parking gear supported by the output shaft so as not to rotate relative to each other, if a large torque is applied to the output shaft, The large torque is applied to the parking gear, and the parking gear comes to a stop with a strong contact with the parking pole. Therefore, when the parking range is switched to another gear range and the parking pole is disengaged from the parking gear (the parking lock is released), a loud noise is generated.

Therefore, when the reverse range is configured by engaging a clutch (brake) that fixes the carrier of the planetary gear mechanism so that the carrier of the planetary gear mechanism is not rotatable, the pulley ratio is changed from the maximum pulley ratio to the maximum pulley ratio. A reduction to a smaller target pulley ratio has been proposed. By reducing the pulley ratio, the power output from the output shaft is reduced and the driving force is reduced, so it is possible to prevent the vehicle from creeping, lack of braking force, and loud noises when the parking lock is released. can be done.

Japanese Patent No. 6529773

After full engagement of the clutch, the primary and secondary pulleys stop and the pulley ratio cannot be changed. Therefore, there are cases where the control for engaging the clutch is completed before the control for reducing the pulley ratio is completed, the pulley ratio does not decrease to the target pulley ratio, and the driving force cannot be sufficiently suppressed.

Also, when the pulley ratio is shifted from the maximum pulley ratio to the high side when the reverse range is configured, if the vehicle is immediately started after switching from the reverse range to the forward range, the start will be sluggish. do.

SUMMARY OF THE INVENTION An object of the present invention is to provide a configuration in which the maximum gear ratio in the second range is larger than the maximum gear ratio in the first range, so that the driving force at the time of starting the vehicle in the second range can be suppressed, and To provide a vehicle control device capable of suppressing the occurrence of sluggishness when starting a vehicle.

In order to achieve the above object, a vehicle control device according to one aspect of the present invention has an endless belt connected to a primary pulley and a secondary pulley on a power transmission path between an input shaft and an output shaft. A belt-type continuously variable transmission mechanism in a wound configuration, a first friction engagement element engaged to form a first range in which power in a first direction is output from an output shaft, and an output and a second friction engagement element engaged to constitute a second range in which power is output from the shaft in a second direction opposite to the first direction, and the continuously variable transmission mechanism is operated in the second range. A control device used in a vehicle equipped with a continuously variable transmission in which the maximum gear ratio when the pulley ratio is maximum is the first range and is greater than the maximum gear ratio when the pulley ratio of the continuously variable transmission mechanism is maximum. and when the second range is configured from a state in which the second range is not configured, the indicated torque, which is the indicated value of the transmission torque of the second frictional engagement element, is set to the maximum transmission torque of the second frictional engagement element. command torque setting means for setting a value smaller than the command torque, transmission torque estimation means for estimating the transmission torque of the second frictional engagement element, and command torque set by the command torque setting means and the transmission torque estimated by the transmission torque estimation means and engagement control means for feedback-controlling the engagement of the second frictional engagement element based on the deviation from the transmission torque.

According to this configuration, when the second range is configured, the indicated torque, which is the indicated value of the transmission torque of the second frictional engagement element, is set to a value lower than the maximum transmission torque of the second frictional engagement element. be. Then, the transmission torque of the second frictional engagement element is estimated, and the engagement of the second frictional engagement element is feedback-controlled based on the deviation between the estimated transmission torque and the command torque.

As a result, even if the pulley ratio of the continuously variable transmission mechanism is the maximum pulley ratio when the vehicle starts moving in the second range, the power output from the output shaft can be suppressed, and the driving force of the drive wheels of the vehicle can be suppressed. can do. As a result, it is possible to prevent the vehicle from creeping out due to creep, insufficient braking force, and loud noise when the parking lock is released.

Further, when the second range is configured, it is not necessary to reduce the pulley ratio of the continuously variable transmission mechanism from the maximum pulley ratio in order to suppress the driving force. Even when the vehicle is immediately started, it is possible to suppress the occurrence of sluggishness in the start.

A vehicle control device according to another aspect of the present invention is a belt-type control device in which an endless belt is wound around a primary pulley and a secondary pulley on a power transmission path between an input shaft and an output shaft. a continuously variable transmission mechanism, a first frictional engagement element engaged to form a first range in which power in the first direction is output from the output shaft, and a first direction opposite to the first direction from the output shaft and a second friction engagement element that is engaged to form a second range in which power in the second direction is output, and is used when the pulley ratio of the continuously variable transmission mechanism is maximum in the second range. A control device used in a vehicle equipped with a continuously variable transmission that is larger than the maximum gear ratio when the maximum gear ratio is the first range and the pulley ratio of the continuously variable transmission mechanism is the maximum, and the second range is configured. An instruction to set the indicated torque, which is the indicated value of the transmission torque of the second frictional engagement element, to be lower than the maximum transmission torque of the second frictional engagement element when the second range is configured from the state in which the torque setting means; and engagement control means for controlling engagement of the second frictional engagement element based on the indicated torque set by the indicated torque setting means, wherein the indicated torque setting means is in the second range. After starting the vehicle, increase the indicated torque.

According to this configuration, when the second range is configured, the indicated torque, which is the indicated value of the transmission torque of the second frictional engagement element, is set to a value lower than the maximum transmission torque of the second frictional engagement element. be. Engagement of the second friction engagement element is controlled based on the indicated torque.

As a result, even if the pulley ratio of the continuously variable transmission mechanism is the maximum pulley ratio when the vehicle starts moving in the second range, the power output from the output shaft can be suppressed, and the driving force of the drive wheels of the vehicle can be suppressed. can do. As a result, it is possible to prevent the vehicle from creeping out due to creep, insufficient braking force, and loud noise when the parking lock is released. After the vehicle starts moving, the instructed torque is increased, so that it is possible to prevent the driving force from becoming insufficient and to ensure drivability.

Further, when the second range is configured, it is not necessary to reduce the pulley ratio of the continuously variable transmission mechanism from the maximum pulley ratio in order to suppress the driving force. Even when the vehicle is immediately started, it is possible to suppress the occurrence of sluggishness in the start.

The vehicle is equipped with a braking device that generates a braking force corresponding to the operating force of the braking operation member, and the instruction torque setting means is based on the gradient of the road surface on which the vehicle is located, the vehicle speed of the vehicle, and the braking force of the braking operation member. The command torque may be set by setting a coefficient based on at least one of the values that vary depending on the input shaft and multiplying the input torque input to the input shaft by the set coefficient.

As a result, after the vehicle starts in the second range, the command torque can be increased according to the state of the vehicle. As a result, it is possible to ensure drivability when starting the vehicle in the second range.

According to the present invention, in a configuration in which the maximum gear ratio in the second range is larger than the maximum gear ratio in the first range, it is possible to suppress the driving force at the time of starting the vehicle in the second range, and in the first range. It is possible to suppress the occurrence of sluggishness in starting the vehicle.

1 is a skeleton diagram showing the configuration of a drive system of a vehicle in which a vehicle control device according to an embodiment of the invention is mounted; FIG. FIG. 5 is a diagram showing the states of the clutch and brake when the vehicle is moving forward and backward; FIG. 4 is a collinear diagram showing the relationship between the number of revolutions (rotational speed) of the sun gear, carrier and ring gear of the planetary gear mechanism; FIG. 3 is a diagram showing the relationship between the belt gear ratio of the belt transmission mechanism and the total gear ratio of the transmission as a whole. 2 is a block diagram showing the configuration of a vehicle control system; FIG. 4 is a block diagram showing functions of a command torque setting unit; FIG. 4 is a block diagram showing functions of a transmission torque estimator; FIG. 4 is a block diagram showing functions of a feedback calculation unit; FIG.

BEST MODE FOR CARRYING OUT THE INVENTION Below, embodiments of the present invention will be described in detail with reference to the accompanying drawings.

<Vehicle drive system>
FIG. 1 is a skeleton diagram showing the configuration of the driving system of the vehicle 1. As shown in FIG.

A vehicle 1 is an automobile having an engine 2 as a drive source.

The engine 2 is provided with an electronic throttle valve for adjusting the amount of intake air into the combustion chamber of the engine 2, an injector (fuel injection device) for injecting fuel into the intake air, and a spark plug for generating electrical discharge in the combustion chamber. It is The engine 2 is also provided with a starter for starting the engine. The power of the engine 2 is transmitted to the differential gear 5 via the torque converter 3 and the transmission 4, and from the differential gear 5 to the left and right driving wheels 7L and 7R via the left and right drive shafts 6L and 6R, respectively. .

The engine 2 has an E/G output shaft 11 . The E/G output shaft 11 is rotated by power generated by the engine 2 .

The torque converter 3 has a front cover 21 , a pump impeller 22 , a turbine runner 23 and a lockup mechanism 24 . The E/G output shaft 11 is connected to the front cover 21 , and the front cover 21 rotates together with the E/G output shaft 11 . The pump impeller 22 is arranged on the opposite side of the front cover 21 from the engine 2 side. The pump impeller 22 is provided so as to be rotatable together with the front cover 21 . The turbine runner 23 is arranged between the front cover 21 and the pump impeller 22 and is rotatable about a rotation axis shared with the front cover 21 .

The lockup mechanism 24 has a lockup piston 25 . Lockup piston 25 is provided between front cover 21 and turbine runner 23 . The lockup mechanism 24 is operated by the differential pressure between the oil pressure in the release oil chamber 26 between the lockup piston 25 and the front cover 21 and the oil pressure in the engagement oil chamber 27 between the lockup piston 25 and the pump impeller 22. Lockup is turned on (engaged)/off (released). That is, when the hydraulic pressure in the disengagement oil chamber 26 is higher than the hydraulic pressure in the engagement oil chamber 27, the lockup piston 25 is moved away from the front cover 21 due to the pressure difference, and the lockup is turned off. When the oil pressure in the engagement oil chamber 27 is higher than the oil pressure in the release oil chamber 26, the lockup piston 25 is pressed against the front cover 21 due to the difference in pressure, and the lockup is turned on.

In the lockup off state, the pump impeller 22 rotates when the E/G output shaft 11 rotates. Rotation of the pump impeller 22 causes a flow of oil from the pump impeller 22 towards the turbine runner 23 . This oil flow is received by the turbine runner 23 and the turbine runner 23 rotates. At this time, an amplifying action of the torque converter 3 occurs, and torque larger than the torque of the E/G output shaft 11 is generated in the turbine runner 23 .

In the lockup ON state, when the E/G output shaft 11 is rotated, the E/G output shaft 11, the pump impeller 22 and the turbine runner 23 rotate together.

The transmission 4 is provided with an input shaft 31 and an output shaft 32, and is so-called power split type (torque split type) transmission. The transmission 4 includes a belt transmission mechanism 33, a front reduction gear mechanism 34, a planetary gear mechanism 35 and a split transmission mechanism 36 to form two power transmission paths.

The input shaft 31 is connected to the turbine runner 23 of the torque converter 3 and is provided so as to be integrally rotatable about the same rotational axis as the turbine runner 23 .

The output shaft 32 is provided parallel to the input shaft 31 . An output gear 37 is supported on the output shaft 32 so as not to rotate relative to it. The output gear 37 meshes with the differential gear 5 (ring gear of the differential gear 5).

The belt transmission mechanism 33 has the same configuration as a known belt-type continuously variable transmission (CVT: Continuously Variable Transmission). Specifically, the belt transmission mechanism 33 includes a primary shaft 41 , a secondary shaft 42 provided parallel to the primary shaft 41 , a primary pulley 43 supported by the primary shaft 41 so as not to rotate relative to the primary shaft 41 , and a secondary shaft 42 . A secondary pulley 44 supported so as not to rotate relatively, and a belt 45 wound around the primary pulley 43 and the secondary pulley 44 are provided.

The primary pulley 43 is arranged to face a fixed sheave 51 fixed to the primary shaft 41 with a belt 45 interposed therebetween. 52. A cylinder 53 fixed to the primary shaft 41 is provided on the opposite side of the movable sheave 52 from the fixed sheave 51 , and a hydraulic chamber 54 is formed between the movable sheave 52 and the cylinder 53 .

The secondary pulley 44 is opposed to a fixed sheave 55 fixed to the secondary shaft 42 with the belt 45 interposed therebetween. 56. A cylinder 57 fixed to the secondary shaft 42 is provided on the opposite side of the movable sheave 56 from the fixed sheave 55 , and a hydraulic chamber 58 is formed between the movable sheave 56 and the cylinder 57 . In the rotation axis direction, the positional relationship between the fixed sheave 55 and the movable sheave 56 is reversed from the positional relationship between the fixed sheave 51 and the movable sheave 52 of the primary pulley 43 .

In the belt transmission mechanism 33, the hydraulic pressure supplied to the hydraulic chamber 54 of the primary pulley 43 and the hydraulic chamber 58 of the secondary pulley 44 are respectively controlled to change the groove widths of the primary pulley 43 and the secondary pulley 44. The belt gear ratio (pulley ratio) is continuously and steplessly changed.

Specifically, when the belt gear ratio is decreased, the hydraulic pressure supplied to the hydraulic chamber 54 of the primary pulley 43 is increased. As a result, the movable sheave 52 of the primary pulley 43 moves toward the fixed sheave 51, and the gap (groove width) between the fixed sheave 51 and the movable sheave 52 is reduced. Accordingly, the winding diameter of the belt 45 around the primary pulley 43 increases, and the interval (groove width) between the fixed sheave 55 and the movable sheave 56 of the secondary pulley 44 increases. As a result, the pulley ratio between the primary pulley 43 and the secondary pulley 44 becomes smaller.

When the belt gear ratio is increased, the hydraulic pressure supplied to the hydraulic chamber 54 of the primary pulley 43 is decreased. As a result, the thrust ratio, which is the ratio of the thrust (primary thrust) of the primary pulley 43 to the thrust (secondary thrust) of the secondary pulley 44, decreases, and the distance between the fixed sheave 55 and the movable sheave 56 of the secondary pulley 44 decreases. , the distance between the fixed sheave 51 and the movable sheave 52 increases. As a result, the pulley ratio between the primary pulley 43 and the secondary pulley 44 increases.

On the other hand, the thrust of the primary pulley 43 and the secondary pulley 44 must be large enough to prevent slippage (belt slippage) between the primary pulley 43 and secondary pulley 44 and the belt 45 . Therefore, the hydraulic pressure supplied to the hydraulic chamber 54 of the primary pulley 43 and the hydraulic chamber 58 of the secondary pulley 44 is controlled so as to obtain necessary and sufficient clamping pressure that does not cause belt slippage.

The front reduction gear mechanism 34 reverses and decelerates the power input to the input shaft 31 and transmits it to the primary shaft 41 . Specifically, the front reduction gear mechanism 34 includes an input shaft gear 61 supported by the input shaft 31 so as not to relatively rotate, and a gear having a larger diameter and a larger number of teeth than the input shaft gear 61 and spline-fitted to the primary shaft 41 . includes a primary shaft gear 62 which is non-rotatably supported by and meshes with the input shaft gear 61 .

The planetary gear mechanism 35 has a sun gear 71 , a carrier 72 and a ring gear 73 . The sun gear 71 is spline-fitted to the secondary shaft 42 so as to be non-rotatable relative to the secondary shaft 42 . The carrier 72 is fitted onto the output shaft 32 so as to be relatively rotatable. Carrier 72 rotatably supports a plurality of pinion gears 74 . A plurality of pinion gears 74 are arranged on the circumference and mesh with the sun gear 71 . The ring gear 73 has an annular shape that collectively surrounds the plurality of pinion gears 74 , and meshes with each pinion gear 74 from the outside in the rotation radial direction of the secondary shaft 42 . The output shaft 32 is connected to the ring gear 73 , and the ring gear 73 is integrally rotatable about the same rotational axis as the output shaft 32 .

The split transmission mechanism 36 is a parallel shaft gear mechanism including a split drive gear 81 and a split driven gear 82 meshing with the split drive gear 81 .

The split drive gear 81 is fitted on the input shaft 31 so as to be relatively rotatable.

The split driven gear 82 is integrally rotatable around the same rotation axis as the carrier 72 of the planetary gear mechanism 35 . The split driven gear 82 is formed with a smaller diameter than the split drive gear 81 and has fewer teeth than the split drive gear 81 .

A parking gear 83 is supported on the output shaft 32 so as not to be relatively rotatable. A parking pole (not shown) is provided around the parking gear 83 . When the parking pole engages with the tooth space of the parking gear 83, the rotation of the parking gear 83 is restricted (parking lock), and when the parking pole disengages from the tooth space of the parking gear 83, the rotation of the parking gear 83 is prevented. Allowed (parking lock released).

The transmission 4 also includes clutches C1, C2 and a brake B1.

The clutch C1 is switched by hydraulic pressure between an engaged state in which the input shaft 31 and the split drive gear 81 are directly connected (coupled so as to be able to rotate integrally) and a released state in which the direct connection is released.

The clutch C2 is switched by hydraulic pressure between an engaged state in which the sun gear 71 and the ring gear 73 of the planetary gear mechanism 35 are directly connected (coupled so as to rotate integrally) and a released state in which the direct connection is released.

The brake B1 is hydraulically switched between an engaged state for braking the carrier 72 of the planetary gear mechanism 35 and a released state for allowing the carrier 72 to rotate.

<Power transmission mode>
FIG. 2 shows the states of the clutches C1, C2 and the brake B1 when the vehicle 1 is traveling forward and backward. FIG. 3 is a collinear diagram showing the relationship between the number of rotations (rotational speed) of the sun gear 71, the carrier 72 and the ring gear 73 of the planetary gear mechanism 35. As shown in FIG. FIG. 4 is a diagram showing the relationship between the belt transmission ratio of the belt transmission mechanism 33 and the total transmission ratio of the transmission 4 as a whole.

In FIG. 2, "o" indicates that the clutches C1, C2 and the brake B1 are engaged. "X" indicates that the clutches C1, C2 and the brake B1 are released.

A shift lever (select lever) is arranged in the vehicle interior of the vehicle 1 at a position that can be operated by the driver. In the movable range of the shift lever, for example, a P (parking) position, an R (reverse) position, an N (neutral) position, and a D (drive) position are arranged in this order.

When the shift lever is in the P position, the clutches C1, C2 and brake B1 are all disengaged, and the parking gear 83 is fixed. ) is constructed. When the shift lever is in the N position, all of the clutches C1, C2 and brake B1 are released, and the parking lock gear is not locked. neutral range) is configured. When both the clutch C1 and the brake B1 are released, the power of the engine 2 is transmitted to the secondary shaft 42, and the secondary shaft 42 rotates. 2 is not transmitted to the driving wheels 7L, 7R.

When the shift lever is positioned at the D position, the D range (forward range), which is one of the shift ranges of the transmission 4, is configured. The power transmission modes in this D range include belt mode and split mode. The belt mode and the split mode are switched by switching between a state in which the clutch C1 is engaged and a state in which the clutch C2 is engaged (replacement of the clutches C1 and C2).

In belt mode, clutch C1 and brake B1 are disengaged and clutch C2 is engaged, as shown in FIG. As a result, the split drive gear 81 is disconnected from the input shaft 31, the carrier 72 of the planetary gear mechanism 35 becomes free (free rotation state), and the sun gear 71 and ring gear 73 of the planetary gear mechanism 35 are directly connected.

The power input to the input shaft 31 is reversed and decelerated by the front reduction gear mechanism 34 and transmitted to the primary shaft 41 of the belt transmission mechanism 33 to rotate the primary shaft 41 and the primary pulley 43 . Rotation of the primary pulley 43 is transmitted to the secondary pulley 44 via the belt 45 to rotate the secondary pulley 44 and the secondary shaft 42 . Since the sun gear 71 and the ring gear 73 of the planetary gear mechanism 35 are directly connected, the sun gear 71 , the ring gear 73 and the output shaft 32 rotate integrally with the secondary shaft 42 . Therefore, in the belt mode, as shown in FIGS. 3 and 4, the total transmission gear ratio of the entire transmission 4 is equal to the belt transmission gear ratio of the belt transmission mechanism 33 and the front reduction gear ratio (rotational speed of input shaft 31/primary shaft 41). number of revolutions).

In split mode, clutch C1 is engaged and clutch C2 and brake B1 are disengaged, as shown in FIG. As a result, the input shaft 31 and the split drive gear 81 are coupled, and the rotation of the input shaft 31 can be transmitted to the carrier 72 of the planetary gear mechanism 35 via the split drive gear 81 and the split driven gear 82. The sun gear 71 of 35 and the ring gear 73 are separated.

Power input to the input shaft 31 is accelerated and transmitted from the split drive gear 81 to the carrier 72 of the planetary gear mechanism 35 via the split driven gear 82 . Power transmitted to the carrier 72 is divided and transmitted from the carrier 72 to the sun gear 71 and the ring gear 73 . The power of sun gear 71 is transmitted to primary shaft gear 62 via secondary shaft 42 , secondary pulley 44 , belt 45 , primary pulley 43 and primary shaft 41 , and from primary shaft gear 62 to input shaft gear 61 . Therefore, in the belt mode, the input shaft gear 61 is the driving gear and the primary shaft gear 62 is the driven gear, whereas in the split mode the primary shaft gear 62 is the driving gear and the input shaft gear 61 is the driven gear. .

Since the gear ratio between the split drive gear 81 and the split driven gear 82 is constant and unchanged (fixed), in the split mode, if the power input to the input shaft 31 is constant, the rotation of the carrier 72 of the planetary gear mechanism 35 is held at a constant speed. Therefore, when the belt transmission ratio is increased, the number of rotations of the sun gear 71 of the planetary gear mechanism 35 decreases. rises. As a result, in the split mode, as shown in FIG. 4, the larger the belt transmission ratio of the belt transmission mechanism 33, the smaller the total transmission ratio of the transmission 4, and the sensitivity of the total transmission ratio to the belt transmission ratio (belt transmission ratio of the amount of change in the total gear ratio to the amount of change in the ratio) is lower than in the belt mode.

Rotation of the output shaft 32 in belt mode and split mode is transmitted to the differential gear 5 via the output gear 37 . As a result, the drive shafts 6L, 6R and the driving wheels 7L, 7R of the vehicle 1 rotate forward.

When the shift lever is positioned at the R position, the R range (reverse range), which is one of the shift ranges of the transmission 4, is configured. In the R range, as shown in FIG. 2, the clutches C1 and C2 are released and the brake B1 is engaged. As a result, the split drive gear 81 is separated from the input shaft 31, the sun gear 71 and the ring gear 73 of the planetary gear mechanism 35 are separated, and the carrier 72 of the planetary gear mechanism 35 is braked.

Power input to the input shaft 31 is reversed and decelerated by the front reduction gear mechanism 34 , transmitted to the primary shaft 41 of the belt transmission mechanism 33 , and transmitted from the primary shaft 41 through the primary pulley 43 , belt 45 and secondary pulley 44 . is transmitted to the secondary shaft 42 and rotates the sun gear 71 of the planetary gear mechanism 35 integrally with the secondary shaft 42 . Since the carrier 72 of the planetary gear mechanism 35 is braked, the ring gear 73 of the planetary gear mechanism 35 rotates in the opposite direction to the sun gear 71 when the sun gear 71 rotates. The direction of rotation of the ring gear 73 is opposite to the direction of rotation of the ring gear 73 during forward movement (belt mode and split mode). The output shaft 32 rotates integrally with the ring gear 73 . Rotation of the output shaft 32 is transmitted to the differential gear 5 via the output gear 37 . As a result, the drive shafts 6L, 6R and the drive wheels 7L, 7R of the vehicle 1 rotate in the reverse direction.

When the vehicle 1 moves forward, the sun gear 71 and the ring gear 73 of the planetary gear mechanism 35 are directly connected so that the rotation speed of the sun gear 71 and the ring gear 73 match. , the rotation speed of the ring gear 73 is always lower than the rotation speed of the sun gear 71 . Therefore, in the R range, the transmission gear ratio is larger than the maximum pulley ratio, and when the maximum pulley ratios are configured in the D range and the R range, the output from the output shaft 32 is greater when the vehicle 1 is moving backward than when moving forward. power increases.

<Vehicle control system>
FIG. 5 is a block diagram showing the configuration of the control system of the vehicle 1. As shown in FIG.

The vehicle 1 is equipped with an ECU (Electronic Control Unit) including a microcomputer (microcontroller unit). A microcomputer includes, for example, a CPU, a nonvolatile memory such as a flash memory, and a volatile memory such as a DRAM (Dynamic Random Access Memory). Although only one ECU 91 is shown in FIG. 5, the vehicle 1 is equipped with a plurality of ECUs having the same configuration as the ECU 91 in order to control each part. A plurality of ECUs including the ECU 91 are connected so as to be capable of two-way communication by a CAN (Controller Area Network) communication protocol.

A unit including the torque converter 3 and the transmission 4 is provided with a hydraulic circuit 92 for supplying hydraulic pressure to each part. The ECU 91 controls various valves and the like included in the hydraulic circuit 92 for gear shift control of the transmission 4 and the like.

Various sensors required for the control are connected to the ECU 91 . As an example of a sensor, the ECU 91 is connected to an engine rotation sensor 93 that outputs a pulse signal synchronized with rotation of the engine 2 (rotation of the crankshaft) as a detection signal. The ECU 91 obtains the engine speed, which is the speed of the engine 2 , from the detection signal of the engine speed sensor 93 .

For control purposes, the ECU 91 receives information from other ECUs such as the gradient of the road surface on which the vehicle 1 is located, the master cylinder pressure that is the oil pressure generated by the master cylinder, and the vehicle speed of the vehicle 1 .

In the vehicle 1, for example, when a brake pedal provided in the passenger compartment is stepped on, the force applied to the brake pedal is transmitted to the brake booster. The pedaling force transmitted to the brake booster is amplified (boosted) by the negative pressure of the brake booster and input from the brake booster to the master cylinder. In the master cylinder, hydraulic pressure is generated according to the force input from the brake booster. The hydraulic pressure generated by the master cylinder is transmitted to the brake actuator. By the function of the brake actuator, hydraulic pressure is distributed to the wheel cylinders of the brakes provided on each wheel, and braking force is applied from each brake to the wheels including the drive wheels 7L and 7R.

Information input to the ECU 91 from another ECU may be obtained from a detection signal of the sensor by connecting a sensor to the ECU 91 . For example, the ECU 91 may be connected to a vehicle speed sensor that outputs a pulse signal having a frequency corresponding to the vehicle speed of the vehicle 1 as a detection signal, and the ECU 91 may obtain the vehicle speed from the detection signal of the vehicle speed sensor.

<Reverse decompression control>
When the shift lever is shifted from a position other than the R position to the R position, a signal corresponding to the shift operation is input from the shift position sensor to the ECU 91. Based on the input of this signal, the ECU 91 selects the R range by the shift operation. An instruction to switch (shift) from a shift range other than the R range to the R range is detected. In response to this detection, the ECU 91 executes reverse pressure reduction control for controlling the engagement of the brake B1. In this reverse pressure reduction control, the engagement pressure of the brake B1 is reduced to cause the brake B1 to slip, thereby suppressing the power output from the output shaft 32 .

The ECU 91 includes a command torque setting section 94 , a transmission torque estimation section 95 , a deviation calculation section 96 and a feedback calculation section 97 as functional processing sections necessary for reverse pressure reduction control. These functional processing units are realized in software by program processing, or in hardware such as logic circuits.

FIG. 6 is a block diagram showing functions of the command torque setting section 94. As shown in FIG.

The command torque setting section 94 has a multiplier 101 . A multiplier 101 receives the turbine torque, the master cylinder pressure/vehicle speed coefficient, and the slope coefficient.

The turbine torque is the torque output to the turbine runner 23 of the torque converter 3 when the speed range is the D range and the speed of the engine 2 is the idling speed, and the capacity coefficient of the torque converter 3 is used. can be calculated by a known method.

The master cylinder pressure/vehicle speed coefficient is a coefficient that is set according to the combination of master cylinder pressure and vehicle speed. The non-volatile memory of the ECU 91 stores a map that defines the relationship between the combination of master cylinder pressure and vehicle speed and the master cylinder pressure/vehicle speed coefficient. For example, this map is such that the master cylinder pressure/vehicle speed coefficient increases stepwise or continuously as the master cylinder pressure decreases, and the master cylinder pressure/vehicle speed coefficient increases stepwise or continuously as the vehicle speed increases. is created in Further, when the master cylinder pressure is equal to or higher than a predetermined value and the vehicle speed is 0 km/h, the master cylinder pressure/vehicle speed coefficient takes a value of 1, for example. The instruction torque setting unit 94 sets the master cylinder pressure/vehicle speed coefficient according to the combination of the master cylinder pressure and the vehicle speed according to the map.

The slope coefficient is a coefficient set according to the slope of the road surface on which the vehicle 1 is located. The non-volatile memory of the ECU 91 stores a map that defines the relationship between gradients and gradient coefficients. This map is created, for example, so that the gradient coefficient increases stepwise or continuously as the downhill gradient of the road surface on which the vehicle 1 is located (the gradient that becomes an uphill slope when moving in reverse) increases. The instruction torque setting unit 94 sets the master cylinder pressure/vehicle speed coefficient according to the gradient of the road surface on which the vehicle 1 is located, according to the map.

The multiplier 101 multiplies the turbine torque (the torque of the turbine runner 23 when the transmission range is the D range and the speed of the engine 2 is the idling speed), the master cylinder pressure/vehicle speed coefficient, and the slope coefficient. be. Then, in the command torque setting unit 94, the multiplied value by the multiplier 101 is set as the command torque, which is the command value (target value) of the transmission torque of the brake B1. The instructed torque set in this manner takes a value smaller than the maximum transmission torque capacity of the brake B1 due to the setting of the master cylinder pressure/vehicle speed coefficient and the gradient coefficient.

FIG. 7 is a block diagram showing functions of the transmission torque estimator 95. As shown in FIG.

The transmission torque estimator 95 has a multiplier 102 . Turbine torque and unit efficiency are input to multiplier 102 . The unit efficiency is the power transmission efficiency of the entire transmission unit including the torque converter 3 and the transmission 4 . A multiplier 102 multiplies the turbine torque and the unit efficiency. Then, in the transmission torque estimator 95, the multiplied value by the multiplier 102 is estimated as the actual value of the transmission torque of the brake B1.

As shown in FIG. 5 , the deviation calculator 96 subtracts the transmission torque estimated by the transmission torque estimator 95 from the command torque set by the command torque setting unit 94 . Then, the subtracted value is output from the deviation calculator 96 as a feedback deviation.

FIG. 8 is a block diagram showing the functions of the feedback calculator 97. As shown in FIG.

Feedback calculator 97 includes multipliers 103 and 104 , integrator 105 and adder 106 .

The multiplier 103 receives the feedback deviation output from the deviation calculator 96 and the P gain, which is the feedback gain of the proportional control. The multiplier 103 multiplies the feedback deviation and the P gain, and outputs the multiplied value as the P control value.

The multiplier 104 receives the feedback deviation output from the deviation calculator 96 and the I gain, which is the feedback gain of the integral control. The multiplier 104 multiplies the feedback deviation and the I gain, and outputs the multiplied value.

Integrator 105 time-integrates the multiplied value of the feedback deviation output from multiplier 104 and the I gain, and outputs the integrated value as an I control value.

Adder 106 receives the P control value output from multiplier 103 and the I control value output from integrator 105 . The adder 106 adds the P control value and the I control value, and the added value is output from the adder 106 as a feedback amount.

In the ECU 91, a command value for the current supplied to the solenoid valve included in the hydraulic circuit 92 is set based on the command torque set by the command torque setting section 94 and the feedback amount calculated by the feedback calculation section 97. be. Then, the current of the command value is supplied to the solenoid valve, and hydraulic pressure corresponding to the opening of the solenoid valve is supplied to the brake B1, so that the transmission torque of the brake B1 matches the command torque. is controlled.

<Effect>
As described above, when the R range is configured, the master cylinder pressure is added to the turbine torque, which is the torque of the turbine runner 23 of the torque converter 3 when the shift range is the D range and the rotational speed of the engine 2 is the idling rotational speed. /The vehicle speed coefficient and the gradient coefficient are multiplied, and the multiplied value is set as the indicated torque for the brake B1. By setting the master cylinder pressure/vehicle speed coefficient and the gradient coefficient, the indicated torque is set to a value lower than the maximum transmission torque of the brake B1. Then, the transmission torque of the brake B1 is estimated, and the engagement of the brake B1 is feedback-controlled based on the deviation between the estimated transmission torque and the command torque.

As a result, when the vehicle 1 starts in the R range, even if the pulley ratio of the belt transmission mechanism 33 is the maximum pulley ratio, the power output from the output shaft 32 can be suppressed. driving force can be suppressed. As a result, it is possible to prevent the vehicle 1 from creeping out due to creep, insufficient braking force, and the generation of a loud noise when the parking lock is released.

Further, when the R range is configured, it is not necessary to reduce the pulley ratio of the belt transmission mechanism 33 from the maximum pulley ratio in order to suppress the driving force. Even when the vehicle is started immediately, it is possible to suppress the occurrence of sluggishness in the start. Then, when the vehicle speed of the vehicle 1 increases after the vehicle 1 starts, the master cylinder pressure/vehicle speed coefficient is increased and the command torque is increased. be able to.

It should be noted that when the brake B1 is brand new until the brake B1 becomes used, the instructed torque of the brake B1 is reduced, and when the hydraulic pressure supplied to the brake B1 is reduced, the brake B1 does not slip smoothly and abnormal vibration, so-called judder, occurs. may occur. Therefore, the reverse pressure reduction control is not executed (prohibited) until the brake B1 becomes used. good too.

<Modification>
Although the embodiments of the present invention have been described above, the present invention can also be implemented in other forms.

For example, during the above-described reverse pressure reduction control, the total amount of heat absorbed by the brake B1 (the amount of heat generated by the slip) is calculated, and when the total amount of heat reaches a predetermined value, the reverse pressure reduction control is stopped. , the brake B1 may be fully engaged. As a result, it is possible to prevent the brake B1 from becoming abnormally high temperature due to the reverse pressure reduction control, thereby suppressing the deterioration of the brake B1.

Further, in the above-described embodiment, when the maximum pulley ratio is configured in the D range and the R range, the power output from the output shaft 32 is larger when the vehicle 1 is moving backward than when moving forward. However, in a configuration in which the power output from the output shaft 32 is greater when the vehicle 1 moves forward than when the vehicle 1 moves backward, the engagement pressure of the clutch C2 is reduced when the D range is configured, and the clutch C2 slips. , the power output from the output shaft 32 may be suppressed.

Further, the present invention provides a belt-type continuously variable transmission mechanism in which an endless belt is wound around a primary pulley and a secondary pulley on a power transmission path between an input shaft and an output shaft. , a vehicle having a forward clutch/reverse clutch provided in series with the continuously variable transmission mechanism.

In addition, various design changes can be made to the above configuration within the scope of the matters described in the claims.

1: Vehicle 4: Transmission (continuously variable transmission)
31: Input shaft 32: Output shaft 33: Belt transmission mechanism (continuously variable transmission mechanism)
43: Primary pulley 44: Secondary pulley 45: Belt 91: ECU (vehicle control device, engagement control means)
94: Instruction torque setting unit (instruction torque setting means)
95: Transmission torque estimation unit (transmission torque estimation means)
96: Deviation calculator (engagement control means)
97: Feedback calculation unit (engagement control means)
B1: Brake (second frictional engagement element, first frictional engagement element)
C2: Clutch (first friction engagement element, second friction engagement element)

Claims (3)

a belt-type continuously variable transmission mechanism in which an endless belt is wound around a primary pulley and a secondary pulley on a power transmission path between an input shaft and an output shaft; A first friction engagement element engaged to form a first range in which directional power is output, and a second friction engagement element in which power in a second direction opposite to the first direction is output from the output shaft. A second frictional engagement element is provided to be engaged to form a range , and in comparison when the pulley ratio of the continuously variable transmission mechanism is the maximum pulley ratio , the gear ratio is in the second range. A control device used in a vehicle equipped with a continuously variable transmission that is larger than the first range ,
When the second range is configured from a state in which the second range is not configured, the indicated torque, which is the indicated value of the transmission torque of the second frictional engagement element, is set to the maximum transmission of the second frictional engagement element. an indicated torque setting means for setting the torque to be lower than the torque;
transmission torque estimating means for estimating the transmission torque of the second friction engagement element;
engagement control means for feedback-controlling the engagement of the second friction engagement element based on the deviation between the command torque set by the command torque setting means and the transmission torque estimated by the transmission torque estimation means; A control device for a vehicle, comprising: a belt-type continuously variable transmission mechanism in which an endless belt is wound around a primary pulley and a secondary pulley on a power transmission path between an input shaft and an output shaft; A first friction engagement element engaged to form a first range in which directional power is output, and a second friction engagement element in which power in a second direction opposite to the first direction is output from the output shaft. A second frictional engagement element is provided to be engaged to form a range , and in comparison when the pulley ratio of the continuously variable transmission mechanism is the maximum pulley ratio , the gear ratio is in the second range. A control device used in a vehicle equipped with a continuously variable transmission that is larger than the first range ,
When the second range is configured from a state in which the second range is not configured, the indicated torque, which is the indicated value of the transmission torque of the second frictional engagement element, is set to the maximum transmission of the second frictional engagement element. an indicated torque setting means for setting the torque to be lower than the torque;
engagement control means for controlling engagement of the second frictional engagement element based on the indicated torque set by the indicated torque setting means;
The vehicle control device, wherein the command torque setting means increases the command torque after the vehicle starts moving in the second range. The vehicle is equipped with a braking device that generates a braking force corresponding to the operating force of the braking operation member,
The command torque setting means sets a coefficient based on at least one of values that vary according to the slope of the road surface on which the vehicle is located, the vehicle speed of the vehicle, and the braking force of the braking operation member, and 3. The vehicle control device according to claim 1, wherein the command torque is set by multiplying the input torque input to the input torque by the set coefficient.

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