JP7033236B2 - Continuously variable transmission hydraulic pressure control device and continuously variable transmission hydraulic pressure control method - Google Patents

Description

The present invention relates to a speed change hydraulic control device for a continuously variable transmission mounted on a vehicle and a speed change hydraulic control method for the continuously variable transmission.

Conventionally, there is known a continuously variable transmission shift control device capable of suppressing a decrease in followability and convergence when the actual gear ratio of the continuously variable transmission is brought closer to the target gear ratio (for example, Patent Document 1). reference). This conventional device is a deviation determination means for determining the degree of change in the deviation between the target gear ratio and the actual gear ratio in the gear shift control device of the stepless transmission that executes feedback control to bring the actual gear ratio closer to the target gear ratio. It has a control switching means for selectively switching between a feedback control including a proportional operation and a feedback control including a proportional operation and an integrating operation based on the determination result of the deviation determining means.

In the above-mentioned conventional device, a hydraulic pressure control system for reducing actual hydraulic pressure vibration is provided by using one feedback compensator that refers to actual hydraulic pressure. However, since the characteristics of the feedback compensator are fixed, they do not match the attenuation characteristics of the feedback compensator when the controlled object characteristics change. If the two characteristics do not match, there is a problem that the hydraulic amplitude may increase or the hydraulic vibration characteristics may become unstable.

Japanese Unexamined Patent Publication No. 2005-98329

The present invention has been made by paying attention to the above problem, and suppresses the power train resonance by reducing the amplitude of hydraulic vibration caused by the torsional fluctuation of the drive system and stabilizing the characteristics regardless of the change in the controlled object characteristics. With the goal.

In order to achieve the above object, the continuously variable transmission of the present invention
Set the target primary pressure and target secondary pressure so that the target gear ratio is set based on the operating condition,
Based on the target primary pressure and the target secondary pressure, the primary pulley command pressure signal and the secondary pulley command pressure signal are generated by control using a feedback compensator that refers to the actual oil pressure.
Determine the characteristic fluctuation factor that the damping characteristic of the feedback compensator does not match due to the change of the controlled object characteristic including the variator.
When the characteristic fluctuation factor is determined, the attenuation characteristic of the feedback compensator is switched according to the control target characteristic.

In this way, when the characteristic fluctuation factor that the feedback compensator's damping characteristic does not match due to the change in the controlled object characteristic is determined, the damping characteristic of the feedback compensator is switched according to the controlled object characteristic. Therefore, regardless of the change in the controlled object characteristic, the power train resonance can be suppressed by reducing the amplitude of the hydraulic vibration caused by the torsional fluctuation of the drive system and stabilizing the characteristic.

FIG. 3 is an overall system diagram showing a drive system and a control system of an engine vehicle to which the shift hydraulic control of the continuously variable transmission of the first embodiment is applied. It is a shift schedule diagram which shows an example of the D range stepless shift schedule used when the stepless shift control in an automatic shift mode is executed by a variator. It is a main part block diagram which shows the variator shift control system by a hydraulic control system which performs switching control of a feedback compensator, and an electronic control system. It is a control block diagram which shows the hydraulic pressure compensator composition of the primary pulley command pressure signal generation part, the characteristic fluctuation factor determination part, and the switching execution part. It is a control block diagram which shows the hydraulic pressure compensator composition of the secondary pulley command pressure signal generation part, the characteristic fluctuation factor determination part, and the switching execution part. It is a flowchart which shows the flow of the switching control processing of the feedback compensator executed in the pulley command pressure signal generation part, the characteristic fluctuation factor determination part, and the switching execution part of a speed change controller. It is a control block diagram which shows the hydraulic pressure compensator composition in the background art. It is a mechanism explanatory drawing which shows the generation mechanism of a power train resonance vibration. It is a hydraulic pressure control correspondence diagram which shows the table which summarized the area judgment and the compensator selection for a purpose and a target frequency. It is a hydraulic vibration characteristic comparison diagram which shows the comparison of the hydraulic vibration characteristic of the shifting hydraulic pressure in the background technique, and the hydraulic vibration characteristic of shifting hydraulic pressure in the switching control of the feedback compensator of Example 1. FIG.

Hereinafter, a mode for carrying out variable speed hydraulic control of the continuously variable transmission of the present invention will be described with reference to Example 1 shown in the drawings.

The shift hydraulic control in the first embodiment is applied to an engine vehicle equipped with a belt-type continuously variable transmission including a torque converter, a forward / backward switching mechanism, a variator, and a final deceleration mechanism. Hereinafter, the configuration of the first embodiment will be described separately by dividing it into an “overall system configuration”, a “variator shift control system configuration”, a “hydraulic pressure compensator configuration”, and a “feedback compensator switching control processing configuration”.

[Overall system configuration]
FIG. 1 shows a drive system and a control system of an engine vehicle to which the shift hydraulic control of the belt-type continuously variable transmission of the first embodiment is applied. Hereinafter, the overall system configuration will be described with reference to FIG.

As shown in FIG. 1, the drive system of the engine vehicle includes an engine 1, a torque converter 2, a forward / backward switching mechanism 3, a variator 4, a final deceleration mechanism 5, and drive wheels 6 and 6. There is.
Here, the belt-type continuously variable transmission CVT is configured by incorporating a torque converter 2, a forward / backward switching mechanism 3, a variator 4, and a final deceleration mechanism 5 in a transmission case (not shown).

The engine 1 can control the output torque by an engine control signal from the outside in addition to the control of the output torque by the accelerator operation by the driver. Torque control is performed on the engine 1 by opening and closing the throttle valve, cutting fuel, and the like. For example, fuel cut control is executed when the vehicle is running on the coast by releasing the accelerator foot.

A motor generator 10 having a starter motor function and a regenerative power generation function for charging when the charging capacity of the vehicle-mounted battery is low is connected to the crankshaft of the engine 1.
The motor generator 10 may be provided with an engine assist function in the starting range.

The torque converter 2 is a starting element with a fluid coupling having a torque amplification function and a torque fluctuation absorption function. It has a lockup clutch 20 capable of directly connecting the engine output shaft 11 (= torque converter input shaft) and the torque converter output shaft 21 when the torque amplification function or the torque fluctuation absorption function is not required. The torque converter 2 includes a pump impeller 23, a turbine runner 24, and a stator 26 as components. The pump impeller 23 is connected to the engine output shaft 11 via the converter housing 22. The turbine runner 24 is connected to the torque converter output shaft 21. The stator 26 is provided in the transmission case via the one-way clutch 25.

The forward / backward switching mechanism 3 is a mechanism that switches the input rotation direction to the variator 4 between a forward rotation direction during forward travel and a reverse rotation direction during reverse travel. The forward / backward switching mechanism 3 has a double pinion type planetary gear 30, a forward clutch 31 with a plurality of clutch plates, and a reverse brake 32 with a plurality of brake plates. The forward clutch 31 is hydraulically engaged by the forward clutch pressure Pfc when the forward traveling range such as the D range is selected. The reverse brake 32 is hydraulically fastened by the reverse brake pressure Prb when the reverse travel range such as the R range is selected. The forward clutch 31 and the reverse brake 32 are both released by draining the forward clutch pressure Pfc and the reverse brake pressure Prb when the N range (neutral range) is selected.

The variator 4 has a primary pulley 42, a secondary pulley 43, and a pulley belt 44, and the gear ratio (ratio of variator input rotation and variator output rotation) is steplessly changed by changing the belt contact diameter. Equipped with a shifting function. The primary pulley 42 is composed of a fixed pulley 42a and a slide pulley 42b arranged coaxially with the variator input shaft 40, and the slide pulley 42b slides by the primary pulley pressure Ppri guided to the primary pressure chamber 45. The secondary pulley 43 is composed of a fixed pulley 43a and a slide pulley 43b arranged coaxially with the variator output shaft 41, and the slide pulley 43b slides by the secondary pulley pressure Psec guided to the secondary pressure chamber 46. The pulley belt 44 is hung between the V-shaped sheave surface of the primary pulley 42 and the V-shaped sheave surface of the secondary pulley 43. The pulley belt 44 is formed of two sets of laminated rings in which a large number of annular rings are stacked from the inside to the outside and a punched plate material, and a large number of ring-shaped laminated rings are attached by sandwiching the two sets of laminated rings. It is composed of elements. The pulley belt 44 may be a chain type belt in which a large number of chain elements arranged in the pulley traveling direction are connected by pins penetrating in the pulley axial direction.

The final deceleration mechanism 5 is a mechanism that decelerates the rotation of the variator output from the variator output shaft 41, gives a differential function, and transmits the differential function to the left and right drive wheels 6 and 6. As a reduction gear mechanism, the final deceleration mechanism 5 includes an output gear 52 provided on the variator output shaft 41, an idler gear 53 and a reduction gear 54 provided on the idler shaft 50, and a final provided at the outer peripheral position of the differential case. It has a gear 55 and. Further, as a differential gear mechanism, it has a differential gear 56 interposed between the left and right drive shafts 51 and 51.

As shown in FIG. 1, the control system of the engine vehicle includes a hydraulic control unit 7, a CVT control unit 8 (abbreviated as "CVTCU"), and an engine control unit 9 (abbreviated as "ECU"). The CVT control unit 8 and the engine control unit 9, which are electronic control systems, are connected by a CAN communication line 13 capable of exchanging information with each other.

The hydraulic control unit 7 has a primary pulley pressure Ppri guided to the primary pressure chamber 45, a secondary pulley pressure Psec guided to the secondary pressure chamber 46, a forward clutch pressure Pfc to the forward clutch 31, and a reverse brake pressure Prb to the reverse brake 32. It is a unit that regulates pressure. The hydraulic control unit 7 includes an oil pump source 70 and a hydraulic control circuit 71 that regulates the control pressure of various hydraulic devices based on the amount of oil discharged from the oil pump source 70. Here, the hydraulic device refers to a hydraulically operated device including a variator 4, a lockup clutch 20, a forward clutch 31, a reverse brake 32, and the like.

Here, the oil pump source 70 is a mechanical oil pump that is rotationally driven by the engine 1 as described later. However, an electric oil pump that is rotationally driven by an electric motor different from the engine 1 may be provided and used in combination with the mechanical oil pump.

The hydraulic control circuit 71 includes a line pressure solenoid valve 72, a primary pressure solenoid valve 73, a secondary pressure solenoid valve 74, a select solenoid valve 75, and a lockup pressure solenoid valve 76. Each solenoid valve 72, 73, 74, 75, 76 performs a pressure adjustment operation according to a control command value (instructed current) output from the CVT control unit 8.

The line pressure solenoid valve 72 adjusts the discharge pressure from the oil pump source 70 to the commanded line pressure PL according to the line pressure command value output from the CVT control unit 8. This line pressure PL is the original pressure when adjusting various control pressures, and is a hydraulic pressure that suppresses belt slip and clutch slip with respect to the torque transmitted to the drive system.

The primary pressure solenoid valve 73 adjusts the pressure reduction to the commanded primary pulley pressure Ppri with the line pressure PL as the original pressure in response to the primary pulley command pressure signal output from the CVT control unit 8. The secondary pressure solenoid valve 74 adjusts the pressure reduction to the secondary pulley pressure Psec commanded by using the line pressure PL as the original pressure in response to the secondary pulley command pressure signal output from the CVT control unit 8.

The select solenoid valve 75 adjusts the pressure reduction to the forward clutch pressure Pfc or the backward brake pressure Prb commanded with the line pressure PL as the original pressure according to the forward clutch pressure command value or the reverse brake pressure command value output from the CVT control unit 8. do.

The lockup pressure solenoid valve 76 adjusts the pressure to the LU instruction pressure Pl that engages / slips / releases the lockup clutch 20 according to the indicated current Alu output from the CVT control unit 8.

The CVT control unit 8 performs line pressure control, shift control, forward / backward switching control, lockup control, and the like. In the line pressure control, a command value for obtaining a target line pressure according to the accelerator opening or the like is output to the line pressure solenoid valve 72. In shift control, when the target gear ratio (target primary rotation speed Npri *) is determined, the pulley command pressure signal to obtain the determined target gear ratio (target primary rotation speed Npri *) is sent to the primary pressure solenoid valve 73 and the secondary pressure solenoid valve 74. Output to. In the forward / backward switching control, a command value for controlling engagement / release of the forward clutch 31 and the reverse brake 32 is output to the select solenoid valve 75 according to the selected range position. In the lockup control, the instruction current Alu that controls the LU instruction pressure Pl that engages / engages / releases the lockup clutch 20 is output to the lockup pressure solenoid valve 76.

Sensor information and switch information from the primary rotation sensor 90, the vehicle speed sensor 91, the secondary pressure sensor 92, the oil temperature sensor 93, the inhibitor switch 94, the brake switch 95, and the turbine rotation sensor 96 are input to the CVT control unit 8. Further, sensor information from the secondary rotation sensor 97, the primary pressure sensor 98, the line pressure sensor 99, and the like is input. In addition to these sensors, it has a primary pulley stroke sensor, a secondary pulley stroke sensor, and the like.

Sensor information from the engine rotation sensor 12, the accelerator opening sensor 14, and the like is input to the engine control unit 9. When the CVT control unit 8 requests the engine rotation information and the accelerator opening information to the engine control unit 9, the CVT control unit 8 receives information on the engine rotation speed Ne and the accelerator opening APO via the CAN communication line 13. Further, when the engine torque information is requested to the engine control unit 9, the information of the actual engine torque Te estimated and calculated in the engine control unit 9 is received via the CAN communication line 13.

FIG. 2 shows an example of a D-range continuously variable transmission schedule used when the continuously variable transmission control in the automatic transmission mode is executed by the variator 4.

The shift control when the D range is selected is the operating point (VSP, APO) on the D range continuously variable transmission schedule of FIG. 2 specified by the vehicle speed VSP (vehicle speed sensor 91) and the accelerator opening APO (accelerator opening sensor 14). ) Determines the target primary rotation speed Npri *. Then, the actual primary rotation speed Npri from the primary rotation sensor 90 is matched with the target primary rotation speed Npri * by feedforward compensation + feedback compensation of the pulley hydraulic pressure.

The gear ratio is represented by the slope of the gear ratio line drawn from the zero operating point, as is clear from the lowest gear ratio line and the highest gear ratio line of the D range continuously variable transmission schedule. Therefore, determining the target primary rotation speed Npri * by the operating point (VSP, APO) determines the target gear ratio of the variator 4.

That is, as shown in FIG. 2, the D-range continuously variable transmission schedule changes the gear ratio steplessly within the range of the gear ratio range according to the lowest gear ratio and the highest gear ratio according to the operating point (VSP, APO). It is set to change. For example, when the vehicle speed VSP is constant, when the accelerator is depressed, the target primary rotation speed Npri * rises and shifts in the downshift direction, and when the accelerator return operation is performed, the target primary rotation speed Npri * decreases and rises. Shift in the shift direction. When the accelerator opening APO is constant, the gear shifts in the upshift direction when the vehicle speed VSP increases, and shifts in the downshift direction when the vehicle speed VSP decreases.

[Variator shift control system configuration]
FIG. 3 shows a variator shift control system using a hydraulic control system and an electronic control system that perform switching control of the feedback compensator. Hereinafter, the variator shift control system configuration will be described with reference to FIG.

The drive system that controls the switching of the feedback compensator includes an engine 1 (driving drive source), a torque converter 2, a forward / backward switching mechanism 3, a variator 4 (continuously variable transmission mechanism), and a final deceleration mechanism 5. It is equipped with a drive wheel 6. The engine 1 has a motor generator 10 connected to a crankshaft. The torque converter 2 has a lockup clutch 20. The forward / backward switching mechanism 3 has a forward clutch 31 and a reverse brake 32. The variator 4 has a primary pulley 42, a secondary pulley 43, and a pulley belt 44.

The hydraulic control system that controls the switching of the feedback compensator includes an oil pump source 70 by a mechanical oil pump driven by the engine 1, a hydraulic control circuit 71, a primary pressure solenoid valve 73, and a secondary pressure solenoid valve 74. To prepare for.

The electronic control system that controls the switching of the feedback compensator includes a CVT control unit 8. The CVT control unit 8 has a speed change controller 80 that outputs a pulley command pressure signal to the variator 4 to the primary pressure solenoid valve 73 and the secondary pressure solenoid valve 74 of the hydraulic control circuit 71.

As shown in FIG. 3, the speed change controller 80 includes a target hydraulic pressure setting unit 801, a primary pulley command pressure signal generation unit 802, a secondary pulley command pressure signal generation unit 803, a characteristic fluctuation factor determination unit 804, and a switching execution unit. 805 and.

The target hydraulic pressure setting unit 801 inputs information on the vehicle speed VSP from the vehicle speed sensor 91 and the accelerator opening APO from the accelerator opening sensor 14. Then, the target primary pressure Ppri * and the target secondary pressure Psec * are set so as to be the target primary rotation speed Npri * (target gear ratio) set based on the operating point (VSP, APO) in the operating state.

The primary pulley command pressure signal generation unit 802 (pulley command pressure signal generation unit) inputs information on the target primary pressure Ppri * from the target hydraulic pressure setting unit 801. Then, based on the target primary pressure Ppri *, F / F compensation + F / B compensation using a feedback compensator (hereinafter referred to as “F / B compensator”) that refers to the actual hydraulic pressure supplied to the primary pulley 42. Generates the primary pulley command pressure signal Ppri (C). The generated primary pulley command pressure signal Ppri (C) is output to the primary pressure solenoid valve 73.

The secondary pulley command pressure signal generation unit 803 (pulley command pressure signal generation unit) inputs information on the target secondary pressure Psec * from the target hydraulic pressure setting unit 801. Then, based on the target secondary pressure Psec *, the secondary is F / F compensation + F / B compensation using the first F / B compensator 803g and the second F / B compensator 803h that refer to the actual hydraulic pressure supplied to the secondary pulley 43. Generates the pulley command pressure signal Psec (C). The generated secondary pulley command pressure signal Psec (C) is output to the secondary pressure solenoid valve 74.

The characteristic fluctuation factor determination unit 804 inputs shift factor information such as the primary actual pressure sensor signal Ppri (S) from the primary pressure sensor 98, the secondary actual pressure sensor signal Psec (S) from the secondary pressure sensor 92, and the oil temperature. .. Then, based on the shift factor information, it is determined that the characteristic fluctuation factor that the damping characteristic of the F / B compensator does not match due to the change in the control target characteristic including the variator 4. The determination result of the characteristic fluctuation factor is output to the switching execution unit 805.

The switching execution unit 805 inputs the determination result of the characteristic variation factor from the characteristic variation factor determination unit 804. Then, when the characteristic fluctuation factor is determined, the attenuation characteristic of the F / B compensator is switched according to the control target characteristic. In the case of the first embodiment, two types of F / B compensators prepared in advance in the primary pulley command pressure signal generation unit 802 are selected. Further, two types of I compensators and two types of PD compensators prepared in advance in the secondary pulley command pressure signal generation unit 803 are selected. Both the I compensator and the PD compensator are examples of F / B compensators.

[Hydraulic pressure compensator configuration]
FIG. 4 shows the hydraulic pressure compensator configuration of the primary pulley command pressure signal generation unit 802, the characteristic fluctuation factor determination unit 804, and the switching implementation unit 805. FIG. 5 shows the hydraulic pressure compensator configuration of the secondary pulley command pressure signal generation unit 803, the characteristic fluctuation factor determination unit 804, and the switching implementation unit 805. Hereinafter, the hydraulic pressure compensator configuration will be described with reference to FIGS. 4 and 5.

As shown in FIG. 4, the primary pulley pressure compensator includes an F / F compensator 802a, a first F / B compensator 802b, and a second F / B compensator 802c in the primary pulley command pressure signal generation unit 802. It has a subtractor 802d and.

The F / F compensator 802a inputs the information of the target primary pressure Ppri * and generates the primary pulley F / F compensation command Ppri (F / F) by the F / F compensation operation.

The first F / B compensator 802b is an F / B compensator having damping characteristics matched to the low frequency hydraulic vibration (several Hz) of the primary actual pressure caused by the torsional fluctuation of the drive system. That is, the first F / B compensator 802b has a higher damping characteristic than the second F / B compensator 802c, which delays the actual pressure response and attenuates the primary actual pressure. Therefore, when the first F / B compensator 802b is selected by the switching implementation unit 805, the primary actual pressure sensor signal Ppri (S) is input from the primary pressure sensor 98 to reduce the hydraulic vibration amount of several Hz. The primary pulley F / B compensation command Ppri (F / B) is generated by compensation.

The second F / B compensator 802c is an F / B compensator having damping characteristics matched to high-frequency hydraulic vibration (several to several tens of Hz) of the primary actual pressure caused by torsional fluctuation of the drive system. That is, the second F / B compensator 802c has a lower damping characteristic that delays the actual pressure response and attenuates the primary actual pressure than the first F / B compensator 802b. Therefore, when the second F / B compensator 802c is selected by the switching implementation unit 805, the primary actual pressure sensor signal Ppri (S) is input from the primary pressure sensor 98 to reduce the hydraulic vibration amount of several to several tens of Hz. The primary pulley F / B compensation command Ppri (F / B) is generated by the damping characteristics.

The subtractor 802d subtracts the primary pulley F / B compensation command Ppri (F / B) from the primary pulley F / F compensation command Ppri (F / F) and outputs it to the primary pressure solenoid valve 73 of the controlled object P (plant). Generates the primary pulley command pressure signal Ppri (C).

Regarding the transition from the 1st F / B compensator 802b to the 2nd F / B compensator 802c, or the transition from the 2nd F / B compensator 802c to the 1st F / B compensator 802b, the transition rate (transition speed). ) Can be set to the same or differently.

As shown in FIG. 5, the secondary pulley pressure compensator includes an F / F compensator 803a, a normative response compensator 803b, a subtractor 803c, and a first I compensator 803d in the secondary pulley command pressure signal generation unit 803. , A second I compensator 803e and an adder 803f. Further, it has a first F / B compensator 803g, a second F / B compensator 803h, and a subtractor 803i.

The F / F compensator 803a inputs the information of the target secondary pressure Psec * and generates the secondary pulley F / F compensation command Psec (F / F) by the F / F compensation calculation.

The normative response compensator 803b inputs the information of the target secondary pressure Psec * and generates the secondary pulley normative response compensation command Psec (N / R) by the normative response compensation operation.

The subtractor 803c generates the secondary pulley norm response difference Psec (ΔN) by subtracting the secondary actual pressure sensor signal Psec (S) from the secondary pulley norm response compensation command Psec (N / R).

The 1st I compensator 803d is an integral compensator having a damping characteristic matched to the low frequency hydraulic vibration (several Hz) of the secondary actual pressure caused by the torsional fluctuation of the drive system. That is, the first I compensator 803d has a higher damping characteristic than the second I compensator 803e, which delays the actual pressure response and attenuates the secondary actual pressure. Therefore, when the first I compensator 803d is selected by the switching implementation unit 805, the secondary pulley norm response difference Psec (ΔN) is input, and the secondary pulley integral compensation command Psec ( I) is generated.

The second I compensator 803e is an integral compensator having a damping characteristic matched to high frequency hydraulic vibration (several to several tens of Hz) of the secondary actual pressure caused by the torsional fluctuation of the drive system. That is, the second I compensator 803e has a damping characteristic that delays the actual pressure response and attenuates the secondary actual pressure, which is lower than that of the first I compensator 803d. Therefore, when the second I compensator 803e is selected by the switching implementation unit 805, the secondary pulley norm response difference Psec (ΔN) is input, and the secondary pulley integral compensation is performed by the integral compensation that reduces the hydraulic vibration amount of several to several tens of Hz. Generate command Psec (I).

The adder 803f adds the secondary pulley F / F compensation command Psec (F / F) and the secondary pulley integral compensation command Psec (I) to generate a pulley integral compensation command pressure signal Psec (C').

The first F / B compensator 803g is an F / B compensator having damping characteristics matched to low-frequency hydraulic vibration (several Hz) of secondary actual pressure caused by torsional fluctuation of the drive system. That is, the first F / B compensator 803g has a higher damping characteristic than the second F / B compensator 803h, which delays the actual pressure response and attenuates the secondary actual pressure. Therefore, when the first F / B compensator 803g is selected by the switching implementation unit 805, the secondary actual pressure sensor signal Psec (S) is input from the secondary pressure sensor 92 to reduce the hydraulic vibration amount of several Hz. The secondary pulley F / B compensation command Psec (F / B) is generated by compensation.

The second F / B compensator 803h is an F / B compensator having damping characteristics matched to high-frequency hydraulic vibration (several to several tens of Hz) of secondary actual pressure caused by torsional fluctuation of the drive system. That is, the second F / B compensator 803h has a damping characteristic that delays the actual pressure response and attenuates the secondary actual pressure, which is lower than that of the first F / B compensator 803g. Therefore, when the second F / B compensator 803h is selected by the switching implementation unit 805, the secondary actual pressure sensor signal Psec (S) is input from the secondary pressure sensor 92 to reduce the hydraulic vibration amount of several to several tens. Generates secondary pulley F / B compensation command Psec (F / B) by / B compensation.

The subtractor 803i subtracts the secondary pulley F / B compensation command Psec (F / B) from the pulley integral compensation command pressure signal Psec (C') and outputs it to the secondary pressure solenoid valve 74 of the controlled object P (plant). Generates the pulley command pressure signal Psec (C).

The transition rate (transition speed) should be set to the same for the transition from the 1st I compensator 803d to the 2nd I compensator 803e or the transition from the 2nd I compensator 803e to the 1st I compensator 803d. You can do it, or you can set it separately. Also, regarding the transition from the 1st F / B compensator 803g to the 2nd F / B compensator 803h, or from the 2nd F / B compensator 803h to the 1st F / B compensator 803g, the transition rate (transition speed). ) Can be set to the same or differently.

[Feedback compensator switching control processing configuration]
FIG. 6 shows the flow of switching control processing of the feedback compensator executed by the pulley command pressure signal generation units 802 and 803 of the speed change controller 80, the characteristic fluctuation factor determination unit 804, and the switching execution unit 805. Hereinafter, each step in FIG. 6 will be described. It should be noted that this process is repeatedly performed according to a predetermined control cycle.

In step S1, it is determined whether or not the switching operating condition for switching the attenuation characteristic of the feedback compensator is satisfied. If YES (switching operation condition is satisfied), the process proceeds to step S2, and if NO (switching operation condition is not satisfied), the process proceeds to step S9.

Here, it is determined that "A. Switching operation condition" is satisfied when all the conditions of the oil amount balance condition, the hydraulic pressure eigenvalue condition, and the oil temperature condition are satisfied, and even one condition is not satisfied. Is determined to be unsuccessful.
-The oil amount balance condition is satisfied when the engine speed Ne or more by the oil temperature axis that does not cause a shortage in the oil amount balance.
-The hydraulic pressure eigenvalue condition is satisfied if the hydraulic eigenvalue is not lower than the specified value. For the hydraulic pressure eigenvalues, the map output from the hydraulic pressure eigenvalues map based on the pulley command pressure (X-axis), target stroke speed (Y-axis), and hydraulic pressure eigenvalues (Z-axis) is corrected by oil temperature (multiply the temperature correction coefficient, or the temperature). Calculated by adding the correction amount).
-As for the oil temperature condition, if the oil temperature is not lower than the judged oil temperature value, the oil temperature condition is satisfied. The determined oil temperature value is calculated from the determined oil temperature value map based on the command pressure difference (X axis) between the line pressure and the pulley pressure, the pulley command pressure (Y axis), and the determined oil temperature value (Z axis).

In step S2, following the determination that the switching operation condition is satisfied in S1, it is determined whether or not the power train resonance countermeasure region is established. If YES (determined to be in the P / T resonance countermeasure region), the process proceeds to step S4, and if NO (determined to be not in the P / T resonance countermeasure region), the process proceeds to step S3.

Here, the shift phase lead operation flag is used to determine the establishment of the "B.P / T resonance countermeasure region condition". The "shift phase lead operation flag" includes the target shift ratio condition, shift speed condition, shift ratio difference condition, shift ratio non-divergence condition, lockup operation condition, travel range selection condition, and non-fail state condition. , It is established that all the conditions with the dither non-operating condition are satisfied.
-As for the target gear ratio condition, the condition is satisfied when the target gear ratio is within a predetermined range.
-As for the shift speed condition, the condition is satisfied when the shift speed is equal to or less than the predetermined speed.
-The gear ratio difference condition is satisfied when the gear ratio difference between the target gear ratio and the actual gear ratio is equal to or less than a predetermined value.
-The gear ratio non-divergence condition is satisfied if the gear ratio does not diverge.
-The lockup operation condition is satisfied when the lockup clutch 20 is in the engaged state.
-The travel range selection condition is other than Ds upshift, and the condition is satisfied when the range position is D, Ds, S, L, M mode, or rattle determination.
-The non-fail state condition is satisfied unless it is in the fail state.
-The dither non-operating condition is satisfied if the dither is not operating.

In step S3, following the determination that the region is not the P / T resonance countermeasure region in S2, it is determined whether or not the hydraulic pressure fixed value is higher than the threshold value. If YES (fixed hydraulic pressure value> threshold value), the process proceeds to step S5, and if NO (fixed hydraulic pressure value ≤ threshold value), the process proceeds to step S4.

Here, the "threshold" refers to a hydraulic pressure specific value that separates low-frequency hydraulic vibration (several Hz) and high-frequency hydraulic vibration (several to several tens of Hz). As for the hydraulic pressure eigenvalue, the map output from the hydraulic pressure eigenvalue map based on the pulley command pressure (X-axis), target stroke speed (Y-axis), and hydraulic eigenvalue (Z-axis) is corrected for oil temperature (multiply the temperature correction coefficient) in the same manner as above. , Or add the temperature correction amount).

In step S4, following the determination in S2 that the area is the P / T resonance countermeasure region or the determination in S3 that the hydraulic pressure fixed value ≤ the threshold value, a low eigenvalue compensator is selected, and the process proceeds to step S6.

Here, the low eigenvalue compensator refers to the first F / B compensator 802b, the first I compensator 803d, and the first F / B compensator 803g selected by the switching execution unit 805.

In step S5, following the determination that the hydraulic pressure fixed value> the threshold value in S3, a high eigenvalue compensator is selected, and the process proceeds to step S6.

Here, the high eigenvalue compensator refers to the second F / B compensator 802c, the second I compensator 803e, and the second F / B compensator 803h selected by the switching execution unit 805.

In step S6, following the compensator selection in S4 or S5, the hydraulic control of the primary pulley pressure Ppri and the secondary pulley pressure Psec of the variator 4 is operated by F / F compensation and F / B compensation using the selected compensator. Proceed to step S7.

Here, when proceeding in the order of S1 → S2 → S4 → S6, hydraulic pressure damping control using the first F / B compensator 802b, the first I compensator 803d, and the first F / B compensator 803g in S6 and the gear ratio fluctuation are performed. It is also used in combination with the shift phase lead control that reduces the phase lag of.

In step S7, following the hydraulic control operation in S6, it is determined whether or not the selective compensator is different from the previous time. If YES (the selective compensator is different from the previous time), the process proceeds to step S8, and if NO (the selective compensator is the same as the previous time), the process proceeds to the end.

In step S8, following the determination that the selective compensator in S7 is different from the previous time, the compensator transition rate between the previous time and the current time is increased, and the process proceeds to the end.

Here, the increase in the compensator transition rate means that when the compensator selected last time is migrated to the compensator selected this time, the compensator is gradually switched at a preset transition rate (transition speed). ..

In step S9, following the determination that the switching operation condition is not satisfied in S1, the shift hydraulic control is stopped, the variator 4 is fixed to the lowest gear ratio, and the process proceeds to the end.

Here, when the shifting hydraulic control is stopped, the line pressure is supplied to the secondary pulley 43, and the shifting ratio of the variator 4 is fixed at the lowest shifting ratio.

Next, "problems of background technology and problem-solving measures" will be explained. Then, the operation in the first embodiment will be described separately for "feedback compensator switching control processing action" and "hydraulic control action for the target frequency".

[Background technology issues and problem-solving measures]
As a background technique for controlling the primary pressure and the secondary pressure of the variator, for example, as shown in FIG. 7, a configuration having one F / B compensator is used. One F / B compensator has a damping characteristic that attenuates a hydraulic vibration component of a specific frequency, and generates an F / B compensated hydraulic signal with F / B compensation based on an input actual hydraulic signal. A hydraulic pressure signal obtained by subtracting the F / B compensated hydraulic pressure signal from the command hydraulic pressure signal by F / F compensation is output to the variator to be controlled.

However, in the case of having one F / B compensator, if the damping characteristics aiming at the damping of the low frequency hydraulic vibration are set as the hydraulic vibration component of the specific frequency to be damped, the high frequency hydraulic vibration cannot be suppressed and the command pressure is not satisfied. The actual pressure response decreases. On the other hand, if the damping characteristic aims at damping the high-frequency hydraulic vibration, the high-frequency hydraulic vibration can be suppressed while suppressing the decrease in the actual pressure response to the command pressure.

On the other hand, one F / B compensator is currently set to a damping characteristic that suppresses high-frequency hydraulic vibration that ensures actual pressure response to a command pressure. Therefore, when traveling in a low gear ratio range, low-frequency hydraulic vibration due to power train resonance is allowed to occur. Therefore, there is a demand to suppress low-frequency hydraulic vibration that gives a sense of discomfort to the driver and occupants while suppressing high-frequency hydraulic vibration.

Here, the mechanism of generating the power train resonance will be described with reference to FIG. First, when torsional fluctuation occurs in the drive shaft of the drive system due to torque input from the tire, the balance thrust fluctuates in the secondary pulley of the variator due to this torsional fluctuation, and the actual gear ratio vibration of the variator accompanies the fluctuation of the balance thrust. Occurs. When the actual gear ratio vibration of the variator is generated, the actual gear ratio vibration is fed back to the hydraulic system as a flow rate fluctuation, and the hydraulic vibration having a large amplitude and unstable vibration characteristics is generated in the hydraulic system. That is, the hydraulic vibration having a large amplitude and unstable vibration characteristics in this hydraulic system becomes a power train resonance (low frequency resonance vibration) in a low frequency region caused by the torsional fluctuation of the drive system.

In response to the above problems, the present inventors have focused on characteristic fluctuation factors in which the damping characteristics of the feedback compensator do not match due to changes in the controlled object characteristics, and based on the determination of the damping characteristics fluctuation factor, determine the damping characteristics of the F / B compensator. I tried to switch. That is, as a solution to the problem of the background technology, in the speed change hydraulic control device of the belt type continuously variable transmission CVT, the speed change controller 80 includes the target hydraulic pressure setting unit 801 and the pulley command pressure signal generation unit 802, 803, and the characteristic fluctuation factor. A means having a determination unit 804 and a switching execution unit 805 was adopted.

Here, the target hydraulic pressure setting unit 801 sets the target primary pressure Ppri * and the target secondary pressure Psec * so as to have a target gear ratio set based on the operating state. Based on the target primary pressure Ppri * and the target secondary pressure Psec *, the pulley command pressure signal generation unit 802, 803 controls the primary pulley command pressure signal Ppri (C) and the secondary pulley by using a feedback compensator that refers to the actual hydraulic pressure. Generates the command pressure signal Psec (C). The characteristic fluctuation factor determination unit 804 determines a characteristic fluctuation factor in which the attenuation characteristic of the feedback compensator does not match due to a change in the controlled target characteristic including the variator 4. When the characteristic fluctuation factor is determined, the switching implementation unit 805 switches the attenuation characteristic of the feedback compensator according to the control target characteristic.

In this way, when the characteristic fluctuation factor determination unit 804 determines the characteristic fluctuation factor that the attenuation characteristics of the feedback compensator do not match due to the change in the controlled object characteristics including the variator 4, the switching implementation unit 805 determines the feedback compensator. The damping characteristic is switched according to the controlled target characteristic. Therefore, regardless of the change in the controlled object characteristic, the power train resonance can be suppressed by reducing the amplitude of the hydraulic vibration caused by the torsional fluctuation of the drive system and stabilizing the characteristic.

For example, as a feedback compensator, one feedback compensator that can change the damping characteristic to a characteristic aimed at damping low-frequency hydraulic vibration and high-frequency hydraulic vibration is prepared. Alternatively, a feedback compensator having a damping characteristic aimed at damping low-frequency hydraulic vibration and a feedback compensator having a damping characteristic aiming at damping high-frequency hydraulic vibration are prepared. Then, when it is necessary to match the damping characteristic of the feedback compensator to the resonance vibration (low frequency resonance vibration) of several Hz due to the change of the controlled target characteristic, it is low by switching to the damping characteristic aiming at the damping of the low frequency hydraulic vibration. Frequency resonance vibration can be suppressed. In addition, when it is necessary to match the damping characteristics of the feedback compensator to the high-frequency resonance vibration of several to several tens of Hz due to changes in the controlled target characteristics, the high-frequency resonance vibration can be changed by switching to the damping characteristics aimed at damping the high-frequency hydraulic vibration. It can be suppressed. In addition, the problem that the actual pressure response to the command pressure when the feedback compensator is switched to the damping characteristic aiming at the damping of low frequency hydraulic vibration is solved by introducing the shift phase lead control. Is possible.

[Feedback compensator switching control processing action]
When the switching operation condition is not satisfied, in the flowchart shown in FIG. 6, the flow from S1 to S9 to the end is repeated, the shift hydraulic control is stopped, and the variator 4 is fixed to the lowest gear ratio.

When the switching operation condition is satisfied and when it is determined that the P / T resonance countermeasure region condition is satisfied, the flow of S1 → S2 → S4 → S6 → S7 → end is repeated in the flowchart shown in FIG. In S4, the first F / B compensator 802b, the first I compensator 803d, and the first F / B compensator 803g, which are low eigenvalue compensators (first feedback compensators), are selected. At this time, the hydraulic damping control using a low eigenvalue compensator and the shift phase lead control for reducing the phase delay of the shift ratio fluctuation by setting the shift phase lead operation flag are used together.

When the switching operation condition is satisfied and the P / T resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is not satisfied, in the flowchart shown in FIG. 6, S1 → S2 → S3 → S4 → S6 → S7. → The flow to the end is repeated. In S4, the first F / B compensator 802b, the first I compensator 803d, and the first F / B compensator 803g, which are low eigenvalue compensators (first feedback compensators), are selected. At this time, only the hydraulic damping control using the low eigenvalue compensator is executed, and the phase lead control for reducing the phase delay of the gear ratio fluctuation is not executed because the shift phase lead operation flag is not set.

When the switching operation condition is satisfied and the P / T resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is determined to be satisfied, in the flowchart shown in FIG. 6, S1 → S2 → S3 → S5 → S6 → S7. → The flow to the end is repeated. In S5, the second F / B compensator 802c, the second I compensator 803e, and the second F / B compensator 803h, which are high eigenvalue compensators (second feedback compensators), are selected.

When switching from the selection of the low eigenvalue compensator to the selection of the high eigenvalue compensator, in the flowchart shown in FIG. 6, the process proceeds from S6 to S7 → S8 → end, and the low eigenvalue compensator to the high eigenvalue compensator. It can be switched by increasing the transfer rate to the vessel.

Similarly, when switching from the selection of the high eigenvalue compensator to the selection of the low eigenvalue compensator, in the flowchart shown in FIG. 6, the process proceeds from S6 to S7 → S8 → end, and the high eigenvalue compensator is low. It can be switched by increasing the transition rate to the eigenvalue compensator.

As described above, the characteristic fluctuation factor determination unit 804 uses the switching operation condition (S1), the P / T resonance countermeasure region condition (S2), and the hydraulic pressure eigenvalue condition (S3) in which the hydraulic eigenvalue is higher than the threshold value. Then, by grasping the change of the controlled object characteristic including the variator 4 by these condition determinations, the characteristic fluctuation factor that the attenuation characteristic of the feedback compensator does not match is determined.

Then, when the switching operation condition is not satisfied, the switching execution unit 805 stops the shift hydraulic control and fixes the variator 4 to the lowest gear ratio because the hydraulic pressure control by compensation is in the low frequency range (several Hz or less). do.

When the switching operation condition is satisfied, the switching execution unit 805 determines that the P / T resonance countermeasure region condition is satisfied, or the P / T resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is not satisfied. If so, select a low eigenvalue compensator (several Hz). Then, the hydraulic control of the primary pulley pressure Ppri and the secondary pulley pressure Psec of the variator 4 is operated by F / B compensation using the selected low eigenvalue compensator.

When the switching operation condition is satisfied, the switching execution unit 805 selects a high eigenvalue compensator when it is determined that the P / T resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is satisfied. Then, the hydraulic control of the primary pulley pressure Ppri and the secondary pulley pressure Psec of the variator 4 is operated by F / B compensation using the selected high eigenvalue compensator (several to several tens of Hz).

[Hydraulic control action on target frequency]
First, for the purpose of # 1 for lower hydraulic pressure eigenvalues, when the target frequency is several Hz or less, as shown in FIG. 9, when determining the region, the hydraulic eigenvalue map is used for calculation. The primary 1st F / B compensator 802b and the primary 2nd F / B compensator 802c are absent (control OFF), the secondary 1st F / B compensator 803g and the secondary 2nd F / B compensator 803h are absent (control OFF). be. Further, for the purpose of # 2 (low temperature) for lowering the hydraulic pressure eigenvalue, when the target frequency is several Hz or less, as shown in FIG. 9, the region is determined by hydraulic pressure. The primary 1st F / B compensator 802b and the primary 2nd F / B compensator 802c are absent (control OFF), the secondary auxiliary 1F / B compensator 803g and the secondary 2nd F / B compensator 803h are absent (control OFF). Is. Here, in the case of aiming at # 1 for the low hydraulic pressure eigenvalue and # 2 for the low hydraulic pressure eigenvalue, the process proceeds from S1 to S9 in the flowchart of FIG.

For the purpose of reducing P / T resonance (for P / T resonance countermeasures), when the target frequency is several Hz (higher frequency than for lower hydraulic pressure eigenvalues), as shown in FIG. Use flags. Then, the primary 1st F / B compensator 802b is used, and the secondary 1st F / B compensator 803g is used. Here, when the purpose is to reduce P / T resonance (for measures against P / T resonance), the process proceeds from S1 → S2 → S4 → S6 in the flowchart of FIG.

For low frequencies other than P / T resonance reduction (for P / T resonance countermeasures), when the target frequency is several Hz (higher frequency than for lower hydraulic eigenvalues), the region is determined as shown in FIG. When the oil pressure eigenvalue map is used for calculation. Then, the primary 1st F / B compensator 802b is used, and the secondary 1st F / B compensator 803g is used. Here, when the purpose is for low frequencies other than P / T resonance reduction (for P / T resonance countermeasures), the process proceeds from S1 → S2 → S3 → S4 → S6 in the flowchart of FIG.

When the target frequency is several to several tens of Hz (higher frequency than for P / T resonance countermeasures) for the purpose of high hydraulic eigenvalue (for high frequency), as shown in FIG. 9, when determining the region, the hydraulic eigenvalue map is used. Calculate. Then, the primary second F / B compensator 802c is used, and the secondary second F / B compensator 803h is used. Here, in the case of aiming at a high hydraulic pressure eigenvalue (for high frequency), the process proceeds from S1 → S2 → S3 → S5 → S6 in the flowchart of FIG.

In the background technology, when an F / B compensator for high frequency is applied to a low frequency controlled object, the hydraulic vibration characteristic of the shifting hydraulic pressure becomes large as shown by the broken line characteristic E in FIG. The vibration characteristics also become unstable.

On the other hand, in the first embodiment, when the low frequency F / B compensator is applied to the low frequency controlled object, the hydraulic vibration characteristic of the shifting hydraulic pressure becomes the hydraulic amplitude as shown in the solid line characteristic F in FIG. Is kept small, and the hydraulic vibration characteristics are stable.

Therefore, in the case of the broken line characteristic E in FIG. 10, the hydraulic vibration has a large amplitude and unstable vibration characteristics in the hydraulic system. On the other hand, in the case of the solid line characteristic F in FIG. 10, the hydraulic vibration amplitude is suppressed to be small in the hydraulic system, the hydraulic vibration characteristic is also stabilized, and the power train resonance of several Hz caused by the torsional fluctuation of the drive system is effectively suppressed. It was proved.

As described above, in the control device of the belt type continuously variable transmission CVT of the first embodiment, the effects listed below can be obtained.

(1) Hydraulically control the pulley command pressure signal to the variator 4 having the primary pulley 42 and the secondary pulley 43, which is interposed in the drive force transmission system from the driving drive source (engine 1) to the drive wheel 6. In a speed change hydraulic control device of a continuously variable transmission (belt type continuously variable transmission CVT) including a speed change controller 80 that outputs to a circuit 71.
The speed change controller 80 is
The target hydraulic pressure setting unit 801 that sets the target primary pressure and the target secondary pressure so that the target gear ratio is set based on the operating condition, and
A pulley command pressure signal generator 802,803 that generates a primary pulley command pressure signal and a secondary pulley command pressure signal by control using a feedback compensator that refers to the actual oil pressure based on the target primary pressure and the target secondary pressure.
The characteristic fluctuation factor determination unit 804 for determining the characteristic fluctuation factor that the attenuation characteristic of the feedback compensator does not match due to the change in the control target characteristic including the variator 4 and the characteristic fluctuation factor determination unit 804.
It has a switching execution unit 805 that switches the attenuation characteristic of the feedback compensator according to the control target characteristic when the characteristic fluctuation factor is determined.
Therefore, regardless of the change in the controlled object characteristic, the power train resonance can be suppressed by reducing the amplitude of the hydraulic vibration caused by the torsional fluctuation of the drive system and stabilizing the characteristic.

(2) The pulley command pressure signal generation unit 802, 803 is a first feedback compensator (1st F) having a damping characteristic corresponding to the low frequency hydraulic vibration of the actual hydraulic pressure caused by the torsional fluctuation of the drive system as a feedback compensator. / B compensator 802b, 1st I compensator 803d, 1st F / B compensator 803g) and a 2nd feedback compensator (1st F / B compensator 803g) having damping characteristics according to the high frequency hydraulic vibration of the actual hydraulic pressure caused by the torsional fluctuation of the drive system. It has a second F / B compensator 802c, a second I compensator 803e, and a second F / B compensator 803h).
Therefore, it is possible to reduce the low frequency hydraulic vibration of the actual hydraulic pressure and the high frequency hydraulic vibration of the actual hydraulic pressure by switching the damping characteristic according to the change of the controlled target characteristic.

(3) The first feedback compensator (1st F / B compensator 802b, 1st I compensator 803d, 1st F / B compensator 803g) has a low frequency of several Hz of actual hydraulic pressure due to torsional fluctuation of the drive system. Set the hydraulic damping characteristics to a higher damping characteristic than the second feedback compensator according to the hydraulic vibration.
The second feedback compensator (second F / B compensator 802c, second I compensator 803e, second F / B compensator 803h) is a high-frequency hydraulic pressure of several to several tens of Hz due to torsional fluctuation of the drive system. The hydraulic damping property is set to a damping characteristic lower than that of the first feedback compensator according to the vibration.
Therefore, by switching the damping characteristics according to the change of the controlled target characteristics, the low frequency hydraulic vibration of several Hz actual hydraulic pressure including the shaking vibration and the high frequency hydraulic vibration of several to several tens of Hz actual hydraulic pressure are reduced. be able to.

(4) The characteristic fluctuation factor determination unit 804 determines the switching operation condition, the power train resonance countermeasure region condition, and the hydraulic eigenvalue condition in which the hydraulic eigenvalue is higher than the threshold value.
Therefore, by determining the three conditions, it is possible to determine the characteristic fluctuation factor in which the damping characteristics of the feedback compensator selected due to the change in the controlled object characteristics do not match.

(5) The characteristic fluctuation factor determination unit 804 uses the shift phase lead operation flag to determine the establishment of the power train resonance countermeasure region condition.
When it is determined that the power train resonance countermeasure region condition is satisfied when the switching operation condition is satisfied, the switching execution unit 805 may use the first feedback compensator (1st F / B compensator 802b, 1st I compensator 803d, 1st F /). The hydraulic damping control using the B compensator 803g) and the shift phase lead control for reducing the phase delay of the shift ratio fluctuation are used in combination.
Therefore, the power train resonance can be effectively reduced by using the first feedback compensator having improved hydraulic damping while ensuring the actual pressure response to the command pressure by the shift phase lead control.

(6) The shift phase advance operation flag is a target gear ratio condition in which the target gear ratio is within a predetermined range, a shift speed condition in which the shift speed is within a predetermined range, and a shift ratio difference between the target shift ratio and the actual shift ratio is within a predetermined range. It is assumed that all the conditions of the gear ratio difference condition, the gear ratio non-divergence condition, the lockup operation condition, the traveling range selection condition, the non-fail state condition, and the dither non-operation condition are satisfied.
It is determined that the power train resonance countermeasure region condition is satisfied when the shift phase advance operation flag is set.
Therefore, the characteristic fluctuation factor determination unit 804 can make a determination by reflecting fluctuation factors such as a target gear ratio, a gear ratio, a gear ratio difference between the target gear ratio and the actual gear ratio.

(7) The switching operation conditions are the oil amount balance condition that there is no shortage of oil amount balance, the hydraulic eigenvalue condition that the oil pressure eigenvalue is not lower than the predetermined oil temperature eigenvalue map output by the pulley command pressure and the target stroke speed, and the oil pressure eigenvalue condition. It is judged that the condition is satisfied when all the conditions of the oil temperature condition where the oil temperature is not lower than the predetermined value are satisfied.
When the switching operation condition is not satisfied, the switching execution unit 805 stops the shift hydraulic control and fixes the variator 4 to the lowest gear ratio.
Therefore, the characteristic fluctuation factor determination unit 804 can make a determination by reflecting fluctuation factors such as oil amount balance, oil pressure eigenvalue, oil temperature, etc., and shift hydraulic pressure in the low frequency vibration range where shift hydraulic control is not good. Control can be stopped.

(8) When the switching operation condition is satisfied, the switching execution unit 805 determines that the power train resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is not satisfied, the first feedback compensator (first F / B compensator). 802b, 1st I compensator 803d, 1st F / B compensator 803g) is selected.
When the switching operation condition is satisfied, if the power train resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is determined to be satisfied, the second feedback compensator (second F / B compensator 802c, second I compensator 803e, first 2F / B compensator 803h) is selected.
Therefore, the low-frequency hydraulic vibration other than the swaying vibration and the high-frequency hydraulic vibration can be separated and reduced by the failure / establishment determination of the hydraulic eigenvalue condition.

(9) A continuously variable transmission (belt type continuously variable transmission) provided with a variator 4 having a primary pulley 42 and a secondary pulley 43, which is interposed in a driving force transmission system from a driving drive source (engine 1) to a drive wheel 6. In the speed change hydraulic control method of CVT)
Set the target primary pressure and target secondary pressure so that the target gear ratio is set based on the operating condition,
Based on the target primary pressure and the target secondary pressure, the primary pulley command pressure signal and the secondary pulley command pressure signal are generated by control using a feedback compensator that refers to the actual oil pressure.
The characteristic fluctuation factor that the attenuation characteristic of the feedback compensator does not match due to the change of the controlled object characteristic including the variator 4 is determined.
When the characteristic fluctuation factor is determined, the attenuation characteristic of the feedback compensator is switched according to the control target characteristic.
Therefore, regardless of the change in the controlled object characteristic, the power train resonance can be suppressed by reducing the amplitude of the hydraulic vibration caused by the torsional fluctuation of the drive system and stabilizing the characteristic.

The shift hydraulic control of the continuously variable transmission of the present invention has been described above based on the first embodiment. However, the specific configuration is not limited to the first embodiment, and design changes and additions are permitted as long as the gist of the invention according to each claim is not deviated from the claims.

In Example 1, two types of first feedback compensators and second feedback compensators having different attenuation characteristics are provided as feedback compensators, and an example in which two types of feedback compensators are switched by the switching implementation unit 805 is shown. However, as the feedback compensator, of course, there may be an example in which three or more types of feedback compensators having different damping characteristics are provided according to the type of hydraulic vibration to be reduced. Further, as the feedback compensator, an example may be provided in which one feedback compensator whose damping characteristics can be changed in multiple steps or steplessly is provided.

In the first embodiment, an example of switching the attenuation characteristics of the feedback compensator is shown as the switching implementation unit 805. However, the switching implementation unit may be an example of switching the attenuation characteristics of the feedforward compensator together with the feedback compensator.

In Example 1, an example is shown in which the speed change hydraulic control of the present invention is applied to an engine vehicle equipped with a belt-type continuously variable transmission CVT as an automatic transmission. However, the control of the present invention may be applied to a vehicle or the like equipped with a continuously variable transmission with an auxiliary transmission. Further, the vehicle to be applied is not limited to an engine vehicle, but can also be applied to a hybrid vehicle in which an engine and a motor are mounted on a driving drive source, an electric vehicle in which a motor is mounted on a driving drive source, and the like.

Claims (9)

It is not equipped with a variator having a primary pulley and a secondary pulley, which is interposed in a drive force transmission system from a driving drive source to a drive wheel, and a speed change controller that outputs a pulley command pressure signal to the variator to a hydraulic control circuit. In the speed change hydraulic control device of the step transmission
The speed change controller
A target hydraulic pressure setting unit that sets the target primary pressure and target secondary pressure so that the target gear ratio is set based on the operating condition,
A pulley command pressure signal generation unit that generates a primary pulley command pressure signal and a secondary pulley command pressure signal by control using a feedback compensator that refers to the actual oil pressure based on the target primary pressure and the target secondary pressure.
A characteristic variation factor determination unit that determines a characteristic variation factor that causes the attenuation characteristics of the feedback compensator to become incompatible due to a change in the controlled object characteristics including the variator.
When the characteristic fluctuation factor is determined, the feedback compensator has a switching implementation unit for switching the attenuation characteristic of the feedback compensator according to the controlled target characteristic.
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for the continuously variable transmission according to claim 1,
As the feedback compensator, the pulley command pressure signal generation unit includes a first feedback compensator having damping characteristics corresponding to low-frequency hydraulic vibration of actual hydraulic pressure caused by torsional fluctuation of the drive system, and torsional fluctuation of the drive system. It has a second feedback compensator, which has damping characteristics according to the high frequency hydraulic vibration of the actual hydraulic pressure due to.
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for the continuously variable transmission according to claim 2.
The first feedback compensator is set to have a damping characteristic higher than that of the second feedback compensator in accordance with the low frequency hydraulic vibration of the actual hydraulic pressure of several Hz caused by the torsional fluctuation of the drive system.
The second feedback compensator sets the hydraulic damping property to a lower damping characteristic than the first feedback compensator in accordance with the high frequency hydraulic vibration of the actual hydraulic pressure of several to several tens of Hz caused by the torsional fluctuation of the drive system. ,
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for the continuously variable transmission according to claim 3.
The characteristic fluctuation factor determination unit determines the switching operation condition, the power train resonance countermeasure region condition, and the hydraulic eigenvalue condition in which the hydraulic eigenvalue is higher than the threshold value.
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for the continuously variable transmission according to claim 4.
The characteristic fluctuation factor determination unit uses the shift phase lead operation flag to determine the establishment of the power train resonance countermeasure region condition.
When it is determined that the power train resonance countermeasure region condition is satisfied when the switching operation condition is satisfied, the switching execution unit performs hydraulic damping control using the first feedback compensator and a phase delay of the gear ratio fluctuation. In combination with reduced shift phase lead control,
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for the continuously variable transmission according to claim 5.
The shift phase advance operation flag has a target gear ratio condition in which the target gear ratio is within a predetermined range, a shift speed condition in which the shift speed is predetermined or less, and a gear ratio difference between the target gear ratio and the actual gear ratio is predetermined or less. It is assumed that all the conditions of the gear ratio difference condition, the gear ratio non-divergence condition, the lockup operation condition, the traveling range selection condition, the non-fail state condition, and the dither non-operation condition are satisfied.
It is determined that the power train resonance countermeasure region condition is satisfied when the shift phase advance operation flag is set.
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for a continuously variable transmission according to any one of claims 4 to 6.
The switching operation conditions include an oil amount balance condition in which there is no shortage of oil amount balance, a hydraulic pressure eigenvalue condition in which the oil pressure eigenvalue map output based on the pulley command pressure and the target stroke speed is corrected by oil temperature is not lower than a predetermined value, and an oil temperature. It is judged that the condition is satisfied when all the conditions of the oil temperature condition which is not lower than the predetermined value are satisfied.
When the switching operation condition is not satisfied, the switching execution unit stops the shift hydraulic control and fixes the variator to the lowest gear ratio.
Speed change hydraulic control device for continuously variable transmission. In the speed change hydraulic control device for a continuously variable transmission according to any one of claims 4 to 7.
When the switching operation condition is satisfied, the switching execution unit selects the first feedback compensator when it is determined that the power train resonance countermeasure region condition is not satisfied but the hydraulic eigenvalue condition is not satisfied.
When the switching operation condition is satisfied, the power train resonance countermeasure region condition is not satisfied, but the hydraulic eigenvalue condition is determined to be satisfied, the second feedback compensator is selected.
Speed change hydraulic control device for continuously variable transmission. In a continuously variable transmission hydraulic control method of a continuously variable transmission equipped with a variator having a primary pulley and a secondary pulley, which is interposed in a driving force transmission system from a driving drive source to a drive wheel.
Set the target primary pressure and target secondary pressure so that the target gear ratio is set based on the operating condition,
Based on the target primary pressure and the target secondary pressure, a primary pulley command pressure signal and a secondary pulley command pressure signal are generated by control using a feedback compensator that refers to the actual hydraulic pressure.
The characteristic fluctuation factor that the attenuation characteristic of the feedback compensator does not match due to the change of the controlled object characteristic including the variator is determined.
When the characteristic fluctuation factor is determined, the damping characteristic of the feedback compensator is switched according to the controlled target characteristic.
How to control the hydraulic pressure of a continuously variable transmission.

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