JP2016142301A - Controller of continuously variable transmission - Google Patents

Abstract

PROBLEM TO BE SOLVED: To provide a controller capable of making the difference between driving force in advance of a vehicle and driving force in backing small, in a continuously variable transmission in which a maximum advance transmission ratio is different from a maximum backing transmission ratio.SOLUTION: During clutch control for shift from an N range to an R range, namely during transient from the N range to the R range, a belt transmission ratio γis made small. Consequently, a unit transmission ratio γat a time point when shift to the R range has been completed is made smaller than a maximum value (maximum backing transmission ratio) of the unit transmission ratio γin the R range.SELECTED DRAWING: Figure 6

Description

  The present invention relates to a control device for a continuously variable transmission.

  As a transmission mounted on a vehicle such as an automobile, a continuously variable transmission mechanism that continuously changes engine power, a gear mechanism that transmits engine power without going through a continuously variable transmission mechanism, and a continuously variable transmission mechanism There has been proposed a planetary gear mechanism for synthesizing the power from the gear and the power from the gear mechanism. In this transmission, the power from the engine can be divided into a continuously variable transmission mechanism and a gear mechanism, and the divided powers can be combined by the planetary gear mechanism and transmitted to the wheels.

JP 2004-176890 A

  The applicant has also proposed a transmission capable of dividing and transmitting the power of the drive source into two systems as a power split type continuously variable transmission.

  The power split type continuously variable transmission according to the proposal includes a continuously variable transmission mechanism that continuously changes power by changing a transmission ratio, a constant transmission mechanism that changes power at a constant transmission ratio, and a continuously variable transmission mechanism. And a planetary gear mechanism for synthesizing the power passing through the power and the power passing through the constant speed change mechanism. The continuously variable transmission mechanism has the same configuration as a known belt-type continuously variable transmission (CVT). A secondary shaft of a continuously variable transmission mechanism is connected to the sun gear of the planetary gear mechanism. An output shaft is connected to the ring gear of the planetary gear mechanism. The rotation of the output shaft is transmitted to the differential gear, and is transmitted from the differential gear to the left and right drive wheels.

  This power split type continuously variable transmission has a belt mode in which power is transmitted to a planetary gear mechanism only through a continuously variable transmission mechanism as a power transmission mode in the D range (forward range), and the power is a continuously variable transmission mechanism. And a split mode transmitted to the planetary gear mechanism via a constant speed change mechanism.

  In the belt mode, the sun gear and the ring gear of the planetary gear mechanism are directly connected, and the carrier of the planetary gear mechanism is brought into a free state. Therefore, the sun gear and the ring gear rotate integrally with the power output from the continuously variable transmission mechanism, and the output shaft rotates integrally with the ring gear. Therefore, in the belt mode, the gear ratio of the power split continuously variable transmission matches the gear ratio (belt gear ratio) of the continuously variable transmission mechanism.

  In the split mode, the direct connection between the sun gear and the ring gear is released, and the power passing through the constant speed change mechanism is input to the carrier. Since the speed ratio (split speed ratio) of the constant speed change mechanism is constant and unchanged, if the rotational speed input to the power split continuously variable transmission is constant, the rotational speed of the carrier is kept constant. Therefore, when the belt transmission ratio is increased and the rotation speed of the sun gear is decreased, the rotation speed of the output shaft connected to the ring gear and the ring gear is increased, and the transmission ratio of the power split continuously variable transmission is decreased. Therefore, in the split mode, the gear ratio (unit gear ratio) of the power split continuously variable transmission is equal to or less than the split gear ratio.

  Further, in the R range (reverse range), the direct connection between the sun gear and the ring gear is released, and the carrier is fixed. Therefore, when the sun gear is rotated by the power output from the continuously variable transmission mechanism, the output shaft rotates integrally with the ring gear in the direction opposite to the rotation direction of the sun gear. At this time, due to the configuration of the planetary gear mechanism, the rotational speed of the ring gear is necessarily lower than the rotational speed of the sun gear.

  Therefore, the maximum value of the unit speed ratio (maximum reverse speed ratio) in the R range is larger than the maximum value of the unit speed ratio (maximum forward speed ratio) in the D range. As a result, if the power input to the power split type continuously variable transmission is the same when the vehicle starts, the driving force transmitted from the differential gear to the left and right driving wheels in the R range compared to the D range. Becomes larger. This driving force difference may cause problems such as jumping out due to a creep phenomenon of the vehicle, insufficient braking force, and the generation of a glone sound (brake noise).

  An object of the present invention is to provide a control capable of reducing a difference between a driving force at the time of forward start of a vehicle and a driving force at the time of reverse start in a continuously variable transmission having different maximum forward speed ratio and maximum reverse speed ratio. Is to provide a device.

  In order to achieve the above object, a control device according to the present invention includes a continuously variable transmission mechanism that continuously changes power input to an input shaft by changing a gear ratio, and power transmitted from at least the continuously variable transmission mechanism. A continuously variable transmission having a maximum forward transmission ratio that is the maximum value of the transmission ratio in the forward range and a maximum reverse transmission ratio that is the maximum value of the transmission ratio in the reverse range. A control device for controlling, a shift instruction detecting means for detecting that a shift to a forward range or a reverse range having a larger one of the maximum forward transmission ratio and the maximum reverse transmission ratio from the neutral range is instructed, and a shift instruction When an instruction is detected by the detection means, the shift control means for controlling the shift from the neutral range to the forward range or the reverse range according to the instruction, and the shift control means During control that, as the speed ratio of the forward range or reverse range after the shift control is smaller, and a ratio changing means for changing the gear ratio of the continuously variable transmission mechanism.

  According to this configuration, for example, when the maximum reverse transmission gear ratio is larger than the maximum forward transmission gear ratio, reverse shift after shift control is being performed during shift control from the neutral range to the reverse range, that is, during transition from the neutral range to the reverse range. The speed ratio of the continuously variable transmission mechanism is changed so that the speed ratio of the range becomes smaller. Thereby, the driving force at the time of backward start of the vehicle can be reduced, and the difference between the driving force and the driving force at the time of forward start of the vehicle can be reduced. Therefore, it is possible to suppress popping out due to the creep phenomenon of the vehicle, insufficient braking force, occurrence of a glone sound, and the like.

  The continuously variable transmission may include a forward engagement element that is engaged when shifting from the neutral range to the forward range, and a reverse engagement element that is engaged when shifting from the neutral range to the reverse range. In that case, the shift control means controls the engagement of the forward engagement element or the reverse engagement element in accordance with the instruction detected by the shift instruction detection means.

  Further, the shift control means may perform slip control for slipping the forward engagement element or the reverse engagement element to be engaged during the change of the speed ratio by the speed ratio changing means.

  By slip control, the engagement of the forward engagement element or the reverse engagement element can be delayed, and the speed ratio of the continuously variable transmission mechanism can be changed to the target speed ratio before the engagement is completed.

  The control device further includes a calorific value calculating unit that calculates a total amount of heat absorbed by the forward engagement element or the reverse engagement element that is the object of slip control during the slip control, and the shift control unit is calculated by the calorie calculating unit. When the amount of heat to be reached reaches a predetermined value, the slip control may be stopped and the forward engagement element or the reverse engagement element that was the object of the slip control may be completely engaged.

  With this configuration, it is possible to suppress the forward engagement element or the reverse engagement element that is the object of slip control from becoming an abnormally high temperature state, and it is possible to suppress the deterioration of the forward engagement element or the reverse engagement element.

Further, the control device according to the present invention is a power split type continuously variable transmission in which the maximum forward transmission ratio that is the maximum value of the transmission ratio in the forward range and the maximum reverse transmission ratio that is the maximum value of the transmission ratio in the reverse range are different. It can use suitably also as a control apparatus to control.
In other words, the power split type continuously variable transmission has a continuously variable transmission mechanism that continuously changes the power input to the input shaft by changing the speed ratio, and the power input to the input shaft is changed at a constant speed ratio. A constant transmission mechanism and an output gear mechanism that outputs the power transmitted from the continuously variable transmission mechanism or the power transmitted from both the continuously variable transmission mechanism and the constant transmission mechanism to the output shaft. When the power from the continuously variable transmission mechanism and the power from the constant transmission mechanism are transmitted to the output gear mechanism, the output gear mechanism needs to combine these powers and output them to the output shaft. For this reason, the planetary gear mechanism may be employed as the output gear mechanism, and in this case, a difference due to the gear ratio of the planetary gear mechanism is likely to occur between the maximum forward transmission gear ratio and the maximum reverse transmission gear ratio. Therefore, the control device according to the present invention can be suitably used as a control device for controlling the power split continuously variable transmission.

  According to the present invention, it is possible to reduce the driving force when the vehicle starts moving backward, and to reduce the difference between the driving force and the driving force when the vehicle moves forward. Therefore, it is possible to suppress popping out due to the creep phenomenon of the vehicle, insufficient braking force, occurrence of a glone sound, and the like.

It is a figure which shows the structure of the principal part of the vehicle by which the control apparatus which concerns on one Embodiment of this invention is mounted. It is a skeleton figure which shows the structure of the drive system of a vehicle. It is a figure which shows the state of the low clutch, reverse brake, and high brake at the time of advance of a vehicle, and reverse drive. It is a figure which shows the relationship between the gear ratio of a continuously variable transmission mechanism, and the gear ratio of a power division type continuously variable transmission. It is an alignment chart which shows the relationship of the rotation speed of the carrier of an output gear mechanism, a sun gear, and a ring gear. In the control which concerns on 1st Embodiment, it is a figure which shows an example of the time change of the turbine speed, belt speed ratio, and reverse control pressure when a shift range is switched from N range to R range. In the control which concerns on 1st Embodiment, it is a figure which shows the other example of the time change of the turbine speed, belt speed ratio, and reverse control pressure when a shift range is switched from N range to R range. In the control which concerns on 2nd Embodiment, it is a figure which shows an example of the time change of the turbine speed, belt gear ratio, and reverse control pressure when a shift range is switched from N range to R range. It is a flowchart which shows the flow of the process performed in the period from the start of slip control to completion | finish. It is a flowchart which shows the flow of an estimated rotation speed calculation process.

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

<Vehicle configuration>

  FIG. 1 is a diagram illustrating a configuration of a main part of a vehicle 1 on which a control device according to an embodiment of the present invention is mounted.

  The vehicle 1 is an automobile that uses the engine 2 as a drive source.

  The output of the engine 2 is transmitted to driving wheels (for example, left and right front wheels) of the vehicle 1 via the torque converter 3 and the power split type continuously variable transmission 4. The engine 2 is provided with a throttle valve for adjusting the amount of intake air into the combustion chamber of the engine 2 and an ignition plug for generating electric discharge in the combustion chamber. The engine 2 is also provided with a starter for starting the engine 2.

  The vehicle 1 is provided with a plurality of ECUs (electronic control units) including a CPU, a ROM, a RAM, and the like. The ECU includes an engine ECU 11 and a transmission ECU 12. The plurality of ECUs are connected so as to be capable of bidirectional communication using a CAN (Controller Area Network) communication protocol.

  An accelerator sensor 13 and an engine speed sensor 14 are connected to the engine ECU 11.

  The accelerator sensor 13 inputs a signal corresponding to an operation amount of an accelerator pedal (not shown) to the engine ECU 11. Based on the signal input from the accelerator sensor 13, the engine ECU 11 sets the ratio of the operation amount to the maximum operation amount of the accelerator pedal, that is, 0% when the accelerator pedal is not depressed, and the accelerator pedal is depressed to the maximum. The accelerator opening, which is a percentage with the time as 100%, is calculated.

  The engine speed sensor 14 inputs a pulse signal synchronized with the rotation of the engine 2 (rotation of the crankshaft) to the engine ECU 11. The engine ECU 11 converts the frequency of the pulse signal input from the engine speed sensor 14 into the speed of the engine 2 (engine speed).

  The engine ECU 11 is a throttle provided in the engine 2 for starting, stopping, and adjusting the output of the engine 2 based on numerical values obtained from signals inputted from various sensors and various information inputted from other ECUs. Control valves and spark plugs.

  A shift position sensor 15, a primary rotation speed sensor 16, a secondary rotation speed sensor 17, an output rotation speed sensor 18, and the like are connected to the transmission ECU 12.

  The shift position sensor 15 inputs a signal corresponding to the position of the shift lever (select lever) to the transmission ECU 12. As positions of the shift lever, for example, a P position, an R position, an N position, and a D position are provided. The P position, the R position, the N position, and the D position correspond to the P range (parking range), R range (reverse range), N range (neutral range), and D range (forward range), respectively. The shift lever can be shifted between the P position, the R position, the N position, and the D position, and the shift range can be instructed by the shift operation.

  The primary rotational speed sensor 16 inputs, for example, a pulse signal synchronized with the rotation of the primary pulley 53 (see FIG. 2) of the power split type continuously variable transmission 4 to the transmission ECU 12. The transmission ECU 12 converts the frequency of the pulse signal input from the primary rotational speed sensor 16 into a primary rotational speed that is the rotational speed of the primary pulley 53 (primary shaft 51).

  The secondary rotational speed sensor 17 inputs, for example, a pulse signal synchronized with the rotation of the secondary pulley 54 (see FIG. 2) of the power split type continuously variable transmission 4 to the transmission ECU 12. The transmission ECU 12 converts the frequency of the pulse signal input from the secondary rotational speed sensor 17 into a secondary rotational speed that is the rotational speed of the secondary pulley 54 (secondary shaft 52).

  The output rotation speed sensor 18 inputs, for example, a pulse signal synchronized with the rotation of the output gear 85 (see FIG. 2) to the transmission ECU 12. The transmission ECU 12 converts the frequency of the pulse signal input from the output rotation speed sensor 18 into an output rotation speed that is the rotation speed of the output shaft 42 (see FIG. 2).

The transmission ECU 12 is based on numerical values obtained from signals input from various sensors, various information input from other ECUs, and the like (hereinafter referred to as “belt transmission ratio”) of the continuously variable transmission mechanism 43. for the control of gamma _b, it controls the valve (not shown) included in the hydraulic circuit 19 for supplying hydraulic pressure to each part of the CVT 43. Further, the transmission ECU 12 controls the hydraulic pressure supplied to the low clutch C1, the reverse brake B1 and the high brake B2 (see FIG. 2) by controlling a valve included in the hydraulic circuit 19, and switches between the forward mode and the reverse mode. Switching between belt mode and split mode.

  The valves of the hydraulic circuit 19 include a hydraulic control valve that adjusts the hydraulic pressure supplied to the primary pulley 53, the secondary pulley 54, the low clutch C1, the reverse brake B1, and the high brake B2, respectively. As the hydraulic control valve, a valve capable of controlling the output hydraulic pressure based on a current value, for example, a linear solenoid valve is used.

<Configuration of drive system>

  FIG. 2 is a skeleton diagram showing the configuration of the drive system of the vehicle 1.

  The engine 2 includes an E / G output shaft 21. The E / G output shaft 21 is rotated by the power generated by the engine 2.

  The torque converter 3 includes a pump impeller 31, a turbine runner 32, and a lockup clutch 33. An E / G output shaft 21 is connected to the pump impeller 31, and the pump impeller 31 is provided so as to be integrally rotatable about the same rotation axis as the E / G output shaft 21. The turbine runner 32 is provided to be rotatable about the same rotation axis as the pump impeller 31. The lockup clutch 33 is provided to directly connect / separate the pump impeller 31 and the turbine runner 32. When the lockup clutch 33 is engaged, the pump impeller 31 and the turbine runner 32 are directly connected, and when the lockup clutch 33 is released, the pump impeller 31 and the turbine runner 32 are separated.

  When the E / G output shaft 21 is rotated in a state where the lockup clutch 33 is released, the pump impeller 31 rotates. When the pump impeller 31 rotates, an oil flow from the pump impeller 31 toward the turbine runner 32 is generated. This oil flow is received by the turbine runner 32 and the turbine runner 32 rotates. At this time, the amplifying action of the torque converter 3 occurs, and the turbine runner 32 generates a power larger than the power (torque) of the E / G output shaft 21.

  In a state where the lockup clutch 33 is engaged, when the E / G output shaft 21 is rotated, the E / G output shaft 21, the pump impeller 31, and the turbine runner 32 are rotated together.

  An oil pump 5 is provided between the torque converter 3 and the power split type continuously variable transmission 4. The pump shaft of the oil pump 5 is provided so as to be rotatable integrally with the pump impeller 31. Accordingly, when the pump impeller 31 is rotated by the power of the engine 2, the pump shaft of the oil pump 5 is rotated and oil is discharged from the oil pump 5.

  The power split type continuously variable transmission 4 transmits the power input from the torque converter 3 to the differential gear 6. The power split type continuously variable transmission 4 includes an input shaft 41, an output shaft 42, a continuously variable transmission mechanism 43, a constant transmission mechanism 44, and an output gear mechanism 45.

  The input shaft 41 is connected to the turbine runner 32 of the torque converter 3 and is provided so as to be integrally rotatable about the same rotation axis as the turbine runner 32.

  The output shaft 42 is provided in parallel with the input shaft 41.

  The continuously variable transmission mechanism 43 has the same configuration as a known belt-type continuously variable transmission (CVT). Specifically, the continuously variable transmission mechanism 43 includes a primary shaft 51 coupled to the input shaft 41, a secondary shaft 52 provided in parallel with the primary shaft 51, and a primary shaft supported by the primary shaft 51 so as not to be relatively rotatable. The pulley 53, the secondary pulley 54 supported by the secondary shaft 52 so that relative rotation is impossible, and the belt 55 wound around the primary pulley 53 and the secondary pulley 54 are provided.

  The primary pulley 53 is disposed so as to face the fixed sheave 61 fixed to the primary shaft 51 with the belt 55 sandwiched between the fixed sheave 61 and is supported by the primary shaft 51 so as to be movable in the axial direction but not to be relatively rotatable. 62. A cylinder 63 fixed to the primary shaft 51 is provided on the opposite side of the movable sheave 62 from the fixed sheave 61, and an oil chamber 64 is formed between the movable sheave 62 and the cylinder 63.

  The secondary pulley 54 is arranged so as to be opposed to the fixed sheave 65 fixed to the secondary shaft 52 with the belt 55 sandwiched between the fixed sheave 65 and supported on the secondary shaft 52 so as to be movable in the axial direction but not to be relatively rotatable. 66. A cylinder 67 fixed to the secondary shaft 52 is provided on the opposite side of the movable sheave 66 from the fixed sheave 61, and an oil chamber 68 is formed between the movable sheave 66 and the cylinder 67.

In the continuously variable transmission mechanism 43, the hydraulic pressure supplied to the oil chamber 64 of the primary pulley 53 and the oil chamber 68 of the secondary pulley 54 is controlled, and the groove widths of the primary pulley 53 and the secondary pulley 54 are changed. The belt speed ratio γ _b is continuously changed in a stepless manner.

More specifically, when the belt speed ratio gamma _b is lowered, the hydraulic pressure is raised to be supplied to the oil chamber 64 of the primary pulley 53. As a result, the movable sheave 62 of the primary pulley 53 moves to the fixed sheave 61 side, and the interval (groove width) between the fixed sheave 61 and the movable sheave 62 is reduced. Accordingly, the winding diameter of the belt 55 around the primary pulley 53 is increased, and the interval (groove width) between the fixed sheave 65 and the movable sheave 66 of the secondary pulley 54 is increased. As a result, the pulley ratio between the primary pulley 53 and the secondary pulley 54 decreases, and the belt speed ratio γ _b decreases.

When the belt speed ratio gamma _b is raised, the hydraulic pressure is reduced is supplied to an oil chamber 64 of the primary pulley 53. Thereby, the thrust of the secondary pulley 54 with respect to the belt 55 becomes larger than the thrust of the primary pulley 53 with respect to the belt 55, the interval between the fixed sheave 65 and the movable sheave 66 of the secondary pulley 54 is reduced, and the fixed sheave 61 and the movable sheave The distance from 62 increases. As a result, the pulley ratio between the primary pulley 53 and the secondary pulley 54 is increased, and the belt speed ratio γ _b is increased.

  On the other hand, the thrust of the primary pulley 53 and the secondary pulley 54 needs to be large enough to prevent slippage between the primary pulley 53 and the secondary pulley 54 and the belt 55. Therefore, the hydraulic pressure supplied to the oil chamber 64 of the primary pulley 53 and the oil chamber 68 of the secondary pulley 54 is controlled so that a thrust according to the magnitude of the torque input to the input shaft 41 is obtained.

  The constant speed change mechanism 44 includes a planetary gear mechanism 71, a split drive gear 72, a split driven gear 73, and an idle gear 74.

  The planetary gear mechanism 71 includes a sun gear 75, a carrier 76 and a ring gear 77. The sun gear 75 is fitted on the input shaft 41 so as to be relatively rotatable. The carrier 76 is supported by the input shaft 41 so as not to be relatively rotatable. The carrier 76 supports a plurality of pinion gears 78 in a rotatable manner. The plurality of pinion gears 78 are arranged on the circumference and mesh with the sun gear 75. The ring gear 77 has an annular shape surrounding the carrier 76 and meshes with each pinion gear 78 from the outside in the rotational radial direction of the input shaft 41.

  The split drive gear 72 is provided so as to be able to rotate integrally with the sun gear 75.

  The split driven gear 73 is provided on the outer periphery of the carrier 82 of the output gear mechanism 45 described below so as to be able to rotate integrally with the carrier 82. That is, the constant speed change mechanism 44 is connected to the carrier 82 of the output gear mechanism 45.

  The idle gear 74 meshes with the split drive gear 72 and the split driven gear 73.

  The output gear mechanism 45 has a configuration of a planetary gear mechanism. That is, the output gear mechanism 45 includes a sun gear 81, a carrier 82, and a ring gear 83. The sun gear 81 is externally fitted to the secondary shaft 52 so as not to be relatively rotatable. The secondary shaft 52 of the continuously variable transmission mechanism 43 is inserted in the center of the carrier 82 so as to be relatively rotatable. The carrier 82 rotatably supports a plurality of pinion gears 84. The plurality of pinion gears 84 are arranged on the circumference and mesh with the sun gear 81. The ring gear 83 has an annular shape surrounding the periphery of the carrier 82, and meshes with each pinion gear 84 from the outer side in the rotational radial direction of the secondary shaft 52. One end of the output shaft 42 is connected to the ring gear 83, and the ring gear 83 is provided so as to be integrally rotatable about the same rotation axis as the output shaft 42. An output gear 85 is supported at the other end of the output shaft 42 so as not to be relatively rotatable.

  The rotation of the output gear 85 is transmitted to the differential gear 6 via the first idle gear 86, the idle shaft 87 and the second idle gear 88. The first idle gear 86 is supported by the idle shaft 87 so as not to be relatively rotatable, and meshes with the output gear 85. The idle shaft 87 is provided in parallel with the output shaft 42. The second idle gear 88 is supported by the idle shaft 87 so as not to be relatively rotatable, and meshes with a ring gear 91 provided in the differential gear 6.

  The power split type continuously variable transmission 4 includes a low clutch C1, a reverse brake B1, and a high brake B2.

  The low clutch C1 is switched between an engaged state (on) in which the output shaft 42 and the secondary shaft 52 are directly connected, and a released state (off) in which the direct connection is released.

  The reverse brake B1 is switched between an engaged state (on) for braking the split drive gear 72 (sun gear 75) and a released state (off) for allowing the split drive gear 72 to rotate.

  The high brake B <b> 2 is switched between an engaged state (on) in which the ring gear 77 is braked and a released state (off) in which the rotation of the ring gear 77 is allowed.

<Power transmission mode>

FIG. 3 is a diagram illustrating states of the low clutch C1, the reverse brake B1, and the high brake B2 when the vehicle 1 moves forward and backward. Figure 4 is a diagram showing a relationship between the gear ratio of the belt speed ratio gamma _b power split type continuously variable transmission 4 (hereinafter referred to as "unit speed ratio".) And gamma _u.

  In FIG. 3, “◯” indicates that the low clutch C1, the reverse brake B1, and the high brake B2 are engaged. In the P range and the N range, the low clutch C1, the reverse brake B1, and the high brake B2 are released.

  The power split type continuously variable transmission 4 has a belt mode and a split mode as power transmission modes in the D range.

  In the belt mode, the high brake B2 and the reverse brake B1 are released, and the low clutch C1 is engaged. As a result, the split drive gear 72, the split driven gear 73 and the idle gear 74 of the constant speed change mechanism 44, and the carrier 82 of the output gear mechanism 45 become free (free rotation state), and the output shaft 42 and the secondary shaft 52 are directly connected. .

The power input to the input shaft 41 is transmitted to the primary shaft 51 of the continuously variable transmission mechanism 43 to rotate the primary shaft 51 and the primary pulley 53. The rotation of the primary pulley 53 is transmitted to the secondary pulley 54 via the belt 55 to rotate the secondary pulley 54 and the secondary shaft 52. Since the low clutch C <b> 1 is engaged, the output shaft 42 rotates integrally with the secondary shaft 52. Therefore, in the belt mode, as shown in FIG. 4, the unit speed ratio γ _u coincides with the belt speed ratio γ _b .

  The rotation of the output shaft 42 is transmitted to the ring gear 91 of the differential gear 6 through the output gear 85, the first idle gear 86, the idle shaft 87 and the second idle gear 88. As a result, the drive shafts 92 and 93 of the vehicle 1 rotate in the forward direction.

  FIG. 5 is a collinear diagram showing the relationship among the rotation speeds of the carrier 82, sun gear 81, and ring gear 83 of the output gear mechanism 45.

  In the split mode, as shown in FIG. 3, the high brake B2 is engaged, and the reverse brake B1 and the low clutch C1 are released. As the high brake B2 is engaged, the ring gear 77 of the constant speed change mechanism 44 is braked. Further, the direct coupling between the output shaft 42 and the secondary shaft 52 is released by releasing the low clutch C1.

  The power input to the input shaft 41 is transmitted to the primary shaft 51 of the continuously variable transmission mechanism 43 to rotate the primary shaft 51 and the primary pulley 53. The rotation of the primary pulley 53 is transmitted to the secondary pulley 54 via the belt 55 to rotate the secondary pulley 54 and the secondary shaft 52. As the secondary shaft 52 rotates, the sun gear 81 of the output gear mechanism 45 rotates.

  Since the ring gear 77 of the constant speed change mechanism 44 is braked, the power input to the input shaft 41 revolves the carrier 76 of the constant speed change mechanism 44 and rotates the pinion gear 78 held by the carrier 76. Let As the pinion gear 78 rotates, power is input from the pinion gear 78 to the sun gear 75. Thereby, the pinion gear 78 and the split drive gear 72 rotate. The rotation of the split drive gear 72 is transmitted to the split driven gear 73 via the idle gear 74 and rotates the split driven gear 73 and the carrier 82 of the output gear mechanism 45.

Since the speed ratio γ _g of the constant speed change mechanism 44 is constant and unchanged (fixed), in the split mode, if the power input to the input shaft 41 is constant, the rotation of the carrier 82 of the output gear mechanism 45 is constant speed. Retained. Therefore, when the belt speed ratio γ _b is increased, the rotational speed of the sun gear 81 of the output gear mechanism 45 is decreased. Therefore, as shown by a two-dot chain line in FIG. ) Increases the rotation speed. As a result, in the split mode, as shown in FIG. 4, the larger the belt speed ratio γ _b , the smaller the unit speed ratio γ _u .

  The rotation of the output shaft 42 is transmitted to the ring gear 91 of the differential gear 6 through the output gear 85, the first idle gear 86, the idle shaft 87 and the second idle gear 88. As a result, the drive shafts 92 and 93 of the vehicle 1 rotate in the forward direction.

  In the R range, as shown in FIG. 3, the high brake B2 and the low clutch C1 are released. Then, the reverse brake B1 is engaged. Thereby, the split drive gear 72 (sun gear 75) is braked. Due to braking of the split drive gear 72, the idle gear 74 of the constant speed change mechanism 44 becomes non-rotatable, and the split driven gear 73 and the carrier 82 become non-rotatable.

  The power input to the input shaft 41 is transmitted to the primary shaft 51 of the continuously variable transmission mechanism 43 to rotate the primary shaft 51 and the primary pulley 53. The rotation of the primary pulley 53 is transmitted to the secondary pulley 54 via the belt 55 to rotate the secondary pulley 54 and the secondary shaft 52. As the secondary shaft 52 rotates, the sun gear 81 of the output gear mechanism 45 rotates. Since the carrier 82 cannot rotate, as shown by a one-dot chain line in FIG. 5, when the sun gear 81 rotates, the ring gear 83 rotates in the opposite direction to the sun gear 81. The rotation direction of the ring gear 83 is opposite to the rotation direction of the ring gear 83 in the D range (belt mode and split mode). Then, the output shaft 42 rotates integrally with the ring gear 83. The rotation of the output shaft 42 is transmitted to the ring gear 91 of the differential gear 6 through the output gear 85, the first idle gear 86, the idle shaft 87 and the second idle gear 88. As a result, the drive shafts 92 and 93 of the vehicle 1 rotate in the reverse direction.

As can be understood with reference to FIG. 5, in the R range, when the belt speed ratio γ _b is the maximum value (lowest), the unit speed ratio γ _u becomes the maximum, and the maximum unit speed ratio γ _u is , greater than the maximum value of the belt speed ratio gamma _b.

<N-R switching control (first embodiment)>

  FIG. 6 is a diagram illustrating an example of temporal changes in the turbine speed, the belt speed ratio, and the reverse control pressure when the shift range is switched from the N range to the R range.

When the shift lever is shifted from the N position to the R position, a signal corresponding to the shift operation is input from the shift position sensor 15 to the transmission ECU 12, and the transmission ECU 12 performs a shift operation based on the input of this signal. An instruction to switch (shift) from the N range to the R range is detected. Then, in response to the detection, by the transmission ECU 12, transmission ratio control for reducing the clutch control and the belt speed ratio gamma _b for engaging the reverse brake B1 is performed.

  In clutch control, in response to an instruction to switch from the N range to the R range, the reverse control pressure is raised to a backlash hydraulic pressure that is higher than the initial pressure (time T1). The reverse control pressure is a target value of the hydraulic pressure supplied to the reverse brake B1, and corresponds to a current value input to the hydraulic control valve for the reverse brake B1. When the reverse control pressure is held at the backlash hydraulic pressure for a predetermined time (time T1-T2), the reverse control pressure is then lowered from the backlash hydraulic pressure to the initial pressure (time T2). While the reverse control pressure is held at the backlash hydraulic pressure, the piston of the reverse brake B1 is pushed by the hydraulic pressure, and the invalid stroke of the piston is eliminated.

  After the reverse control pressure is maintained at the initial pressure over a predetermined time (time T2-T3), the reverse control pressure is gradually increased from the initial pressure. As a result, the hydraulic pressure supplied to the reverse brake B1 increases, and the turbine rotational speed, which is the rotational speed of the turbine runner 32 of the torque converter 3, begins to deviate from the engine rotational speed. Since the turbine runner 32 rotates integrally with the input shaft 41 of the power split type continuously variable transmission 4, the turbine rotational speed is the same as the rotational speed of the input shaft 41. Further, since the primary shaft 51 is connected to the input shaft 41 and the primary pulley 53 is supported by the primary shaft 51 so as not to be relatively rotatable, the turbine rotational speed is the same as the primary rotational speed.

  Thereafter, when the transmission ECU 12 detects that the ratio between the turbine speed and the engine speed has become a predetermined ratio (out of synchronization), the reverse control pressure is increased at a predetermined time gradient (time change rate). Sweep control is started (time T4). By this sweep control, the hydraulic pressure supplied to the reverse brake B1 further increases, the transmission torque of the reverse brake B1 increases, and the turbine rotational speed decreases.

On the other hand, the gear ratio control, for example, as a belt gear ratio gamma _b is changed from the N range to the target speed ratio α at a predetermined time from the instruction for switching to R-range, the gear ratio control pressure is at a constant time gradient Raised. The transmission ratio control pressure is a target value of the hydraulic pressure supplied to the primary pulley 53 (oil chamber 64) of the continuously variable transmission mechanism 43, and corresponds to the current value input to the hydraulic control valve for the primary pulley 53. As the transmission ratio control pressure increases, the hydraulic pressure supplied to the primary pulley 53 increases, and the movable sheave 62 of the primary pulley 53 moves toward the fixed sheave 61, so that the belt transmission ratio γ _b becomes the target transmission ratio α. Decrease gradually. Target gear ratio α is, for example, unit speed ratio gamma _u in R range maximum value of the unit gear ratio gamma _u in D-range is set so as words to match the maximum value of the belt speed ratio gamma _b.

When the turbine rotational speed decreases due to clutch control and the turbine rotational speed decreases to a predetermined rotational speed (time T5), the transmission ECU 12 determines whether or not the belt speed ratio γ _b has changed to the target speed ratio α. The The predetermined rotational speed is set, for example, to a value obtained by adding a constant value ΔN (rpm) to the turbine rotational speed (for example, 0 rpm) when the reverse brake B1 is completely engaged (synchronization point). Further, the belt speed ratio γ _b is obtained by dividing the primary rotational speed by the secondary rotational speed.

Normally, as shown in FIG. 6, the belt speed ratio γ _b changes to the target speed ratio α before the turbine speed decreases to a predetermined speed. In this case, the sweep control in the clutch control is continued. Then, when the reverse brake B1 is almost completely engaged and the turbine rotational speed decreases to a predetermined end rotational speed (time T6), the sweep control is terminated. Then, the reverse control pressure is increased to the maximum pressure, and the clutch control is terminated (time T7). As a result, the reverse brake B1 is completely engaged, and the decrease in the turbine rotational speed is stopped.

  FIG. 7 is a diagram illustrating another example of the temporal change of the turbine speed, the belt speed ratio, and the reverse control pressure when the shift range is switched from the N range to the R range.

As shown in FIG. 7, due to various factors such as fluctuations in engine speed and oil temperature, when the turbine speed decreases to a predetermined speed (time T5), the belt speed ratio γ _b becomes the target speed ratio α. It may not change until. In this case, the sweep control in the clutch control is terminated, and the hydraulic pressure supplied to the reverse brake B1 is feedback controlled (slip control) so that the turbine speed is maintained at a predetermined speed. During the slip control, the transmission ECU 12 monitors whether or not the belt speed ratio γ _b has changed to the target speed ratio α. When the belt speed ratio γ _b matches the target speed ratio α (time T8), the slip control is finished, the reverse control pressure is increased to the maximum pressure, and the clutch control is finished (time T9). As a result, the reverse brake B1 is completely engaged, and the decrease in the turbine rotational speed is stopped.

By slip control, the complete engagement of the reverse brake B1 can be delayed, and the belt speed ratio γ _b can be changed to the target speed ratio α before the complete engagement.

<N-R switching control (second embodiment)>

  FIG. 8 is a diagram illustrating an example of temporal changes in the turbine speed, the belt speed ratio, and the reverse control pressure when the shift range is switched from the N range to the R range.

  Instead of the above-described clutch control and gear ratio control, clutch control and gear ratio feedback control described below may be executed.

  In clutch control, in response to an instruction to switch from the N range to the R range, the reverse control pressure is raised to a backlash hydraulic pressure that is higher than the initial pressure (time T11). When the reverse control pressure is held at the backlash hydraulic pressure for a predetermined time (time T11-T12), the reverse control pressure is then lowered from the backlash hydraulic pressure to the initial pressure (time T12). While the reverse control pressure is held at the backlash hydraulic pressure, the piston of the reverse brake B1 is pushed by the hydraulic pressure, and the invalid stroke of the piston is eliminated.

  After the reverse control pressure is maintained at the initial pressure for a predetermined time (time T12-T13), the reverse control pressure is gradually increased from the initial pressure. As a result, the hydraulic pressure supplied to the reverse brake B1 increases, and the turbine rotational speed, which is the rotational speed of the turbine runner 32 of the torque converter 3, begins to deviate from the engine rotational speed.

  Thereafter, when the transmission ECU 12 detects that the ratio between the turbine speed and the engine speed has become a predetermined ratio (out of synchronization), the hydraulic pressure supplied to the reverse brake B1 while slipping the reverse brake B1. Is started (time T14). When the hydraulic pressure supplied to the reverse brake B1 increases, the turbine rotational speed decreases and the reverse brake B1 is almost completely engaged, and the turbine rotational speed decreases to a predetermined end rotational speed (time T15). ), Slip control is terminated. Then, the reverse control pressure is increased to the maximum pressure, and the clutch control is terminated (time T16). As a result, the reverse brake B1 is completely engaged, and the decrease in the turbine rotational speed is stopped.

In the transmission ratio feedback control, feedback control is performed so that the belt transmission ratio γ _b decreases to the target transmission ratio α. For example, a deviation between the belt transmission ratio γ _b and the target transmission ratio α is obtained, and the deviation is multiplied by a control gain, whereby a transmission ratio control pressure that is a target value of the hydraulic pressure supplied to the primary pulley 53 is calculated. The Then, a current having a current value corresponding to the gear ratio control pressure is supplied to the hydraulic control valve for the primary pulley 53.

By cooperation with the speed ratio feedback control and slip control in the clutch control, in accordance with the timing at which the state of reverse brake B1 is fully engaged, the belt speed ratio gamma _b matches the target gear ratio alpha.

  FIG. 9 is a flowchart showing a flow of processing executed during a period from the start to the end of the slip control.

  During the period from the start to the end of the slip control, the transmission ECU 12 executes the processing shown in FIG. 9, thereby achieving cooperation between the slip control and the speed ratio feedback control.

First, it is determined whether or not the belt speed ratio γ _b is within a predetermined range, for example, a range where the target speed ratio α is a lower limit and a value obtained by adding a constant value to the target speed ratio α is an upper limit (step). S1).

When the belt speed ratio γ _b is within the predetermined range (YES in step S1), the control gain of the speed ratio feedback control is reduced (step S2).

On the other hand, if the belt speed ratio γ _b is not within the predetermined range, that is, if the belt speed ratio γ _b is larger than the upper limit of the predetermined range (NO in step S1), the control gain of the speed ratio feedback control is increased ( Step S3). Usually, immediately after the start of the slip control, the belt speed ratio γ _b is larger than the upper limit of the predetermined range, so that the control gain of the speed ratio feedback control is increased. As a result, the belt speed ratio γ _b quickly changes to the upper limit of the predetermined range.

  Further, the turbine torque Tt and the engine torque Te are calculated (step S4).

  The turbine torque Tt is the torque of the turbine runner 32 of the torque converter 3, and the capacity coefficient C and the torque ratio R of the torque converter 3 are added to the square value of the estimated engine speed Ne calculated by the estimated speed calculation process described later. It is calculated by multiplying. That is, the turbine torque Tt is obtained according to the following equation.

Tt = Ne ² × C × R

  The engine torque Te is obtained by multiplying the square value of the estimated engine speed Ne by the capacity coefficient C of the torque converter 3. That is, the engine torque Te is obtained according to the following equation.

Te = Ne ² × C

In order to obtain the capacity coefficient C and the torque ratio R of the torque converter 3, the speed ratio of the torque converter 3 is calculated. As for the calculation of the speed ratio, by the current belt speed ratio gamma _b to the output rotational speed is multiplied, the turbine speed is determined. Then, the speed ratio of the torque converter 3 is calculated by dividing the turbine speed by the estimated engine speed Ne. The memory of the transmission ECU 12 stores the relationship between the speed ratio of the torque converter 3 and the capacity coefficient C and the relationship between the speed ratio of the torque converter 3 and the torque ratio R in the form of maps, respectively. A capacity coefficient C and a torque ratio R corresponding to the speed ratio of the torque converter 3 are acquired from the map.

After that, the current belt speed ratio gamma _b and final drive ratio to the turbine torque Tt (final gear ratio) is multiplied, R-range driving force is calculated. Further, the D range assumed driving force is calculated by multiplying the turbine torque Tt by the maximum value of the unit speed ratio γ _u in the D range (the maximum value of the belt speed ratio γ _b ) and the final reduction ratio. Then, the D range assumed driving force is divided by the R range driving force, and it is determined whether or not the division value is 1 or more (step S5).

  If the division value is less than 1 (NO in step S5), the target torque capacity of the reverse brake B1 is calculated so that the R range driving force matches the D range assumed driving force (step S6).

  When the target torque capacity is calculated, the reverse control pressure is calculated so that the transmission torque capacity of the reverse brake B1 matches the target torque capacity (step S7).

  Then, a current having a current value corresponding to the reverse control pressure is input to the hydraulic control valve for the reverse brake B1 (step S8).

  After that, it returns to step S1, and the process after step S1 is performed again.

When the slip control progresses and the belt speed ratio γ _b becomes equal to or lower than the upper limit of the predetermined range (YES in step S1), the control gain of the speed ratio feedback control is reduced (step S2). As a result, the belt speed ratio γ _b gradually changes toward the target speed ratio α.

  Then, when the value obtained by dividing the D range driving force by the R range driving force becomes 1 or more (YES in step S5), the process shown in FIG. 9 is terminated together with the slip control.

  FIG. 10 is a flowchart showing the flow of the estimated rotational speed calculation process.

  The estimated rotational speed calculation process is a process for calculating the estimated engine rotational speed Ne, and is repeatedly executed by the transmission ECU 12 at a constant period during the period from the start to the end of the slip control.

  In the estimated rotation speed calculation process, first, the accelerator opening obtained from the output signal of the accelerator sensor 13 by the engine ECU 11 and the output signal of the engine speed sensor 14 by the engine ECU 11 by CAN communication with the engine ECU 11. The engine speed NE is acquired (step S11).

  Next, the engine torque characteristic line is referred to, and an estimated engine torque corresponding to the accelerator opening and the engine speed NE is calculated (step S12). The engine torque characteristic line is a characteristic line that represents the relationship between the accelerator opening, the engine speed, and the engine torque, and is stored in the memory of the transmission ECU 12 in the form of a map.

  When the engine torque characteristic line (map) is stored in the memory of the engine ECU 11, the engine ECU 11 calculates an estimated engine torque corresponding to the accelerator opening and the engine speed NE, and the calculated estimated engine torque. May be transmitted from the engine ECU 11 to the transmission ECU 12.

  Thereafter, a load torque (torque due to an oil pump load or an alternator load) determined by the engine speed NE is obtained, and the estimated input torque is calculated by subtracting the load torque from the estimated engine torque (step S13).

  Next, the difference between the estimated input torque and the engine torque Te calculated in step S4 shown in FIG. 9 is obtained, and the difference of the engine 2 is multiplied by the difference of the engine 2 to calculate the amount of change in the rotational speed of the engine 2. Is done. Then, the calculated fluctuation amount is added to the estimated engine speed Ne calculated in the previous estimated speed calculation process, thereby calculating the temporary engine speed. Then, the temporary engine speed NE and the engine speed NE acquired from the engine ECU 11 are compared, and the larger of the temporary engine speed NE and the engine speed NE is calculated as the estimated engine speed Ne (step). S14). Thus, the estimated rotation speed calculation process ends.

<Effect>

As described above, the belt speed ratio γ _b is reduced during clutch control for shifting from the N range to the R range, that is, during the transition from the N range to the R range. Thus, the unit gear ratio gamma _u at the time when the shift to the R range is completed is less than the maximum value of the unit gear ratio gamma _u in R range (maximum reverse speed ratio). As a result, it is possible to reduce the driving force when the vehicle moves backward (when the vehicle starts moving backward), and to reduce the difference between the driving force and the driving force when the vehicle moves forward (when moving forward). Therefore, it is possible to suppress popping out due to the creep phenomenon of the vehicle, insufficient braking force, occurrence of a glone sound, and the like.

<Modification>

  As mentioned above, although embodiment of this invention was described, this invention can also be implemented with another form.

  For example, when the total amount of heat absorbed by the reverse brake B1 (the amount of heat generated by slip) is calculated during the slip control in each of the above-described embodiments, and the total amount of heat reaches a predetermined value, the slip control is performed. It may be stopped and the reverse brake B1 may be completely engaged. Thereby, it can suppress that reverse brake B1 will be in an abnormally high temperature state by slip control, and can suppress degradation of reverse brake B1.

In the N range, the belt speed ratio γ _b may be held at an intermediate speed ratio between the maximum speed ratio and the target speed ratio α, instead of the maximum speed ratio. In this case, when the shift range is switched from the N range to the R range, the belt speed ratio γ _b is changed from the intermediate speed ratio to the target speed ratio α, and when the shift range is switched from the N range to the D range. The belt speed ratio γv is changed from the intermediate speed ratio to the maximum speed ratio. Thus, it is possible to shorten the time required for changing the belt speed ratio gamma _b during the range switching.

  In a power split continuously variable transmission configured such that the maximum speed ratio in the D range is larger than the maximum speed ratio in the R range, when switching from the N range to the D range, The speed ratio of the continuously variable transmission mechanism may be changed so that the speed ratio becomes small.

  Further, in the above-described embodiment, the example in which the present invention is applied to the power split type continuously variable transmission is taken up. However, the present invention is not limited to the power split type, and is variously provided with a continuously variable transmission mechanism. Can be applied.

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

11 Engine ECU
12 Transmission ECU (shift instruction detecting means, shift control means, speed ratio changing means)
41 Input shaft (input shaft)
42 Output shaft (output shaft)
43 continuously variable transmission mechanism 44 constant transmission mechanism 45 output gear mechanism

Claims (1)

A continuously variable transmission mechanism that continuously changes the power input to the input shaft by changing the transmission gear ratio, and an output gear mechanism that outputs at least the power transmitted from the continuously variable transmission mechanism to the output shaft; A control device that controls a continuously variable transmission in which a maximum forward transmission ratio that is a maximum value of a transmission ratio in the engine and a maximum reverse transmission ratio that is a maximum value of a transmission ratio in a reverse range are different,
A shift instruction detection means for detecting that a shift to the forward range or the reverse range having a larger one of the maximum forward transmission ratio and the maximum reverse transmission ratio from a neutral range is instructed;
Shift control means for controlling a shift from the neutral range to the forward range or the reverse range according to the instruction when the instruction is detected by the shift instruction detection means;
And a gear ratio changing means for changing a gear ratio of the continuously variable transmission mechanism so that a gear ratio of the forward range or the reverse range after the shift control is reduced during the control by the shift control means. apparatus.

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