JP2021143678A - Estimation device and estimation method - Google Patents

Abstract

To improve estimation accuracy on clutch characteristic.SOLUTION: An estimation device includes: an input frequency sensor 90; output rotation number sensors 93, 94; a clutch control portion 112 executing clutch shifting control to shift clutches 21, 22 from a release state to an engagement state by operating the same under a prescribed hydraulic pressure, from a torque phase generated in a shift progress process to an inertia phase; a differential rotation control portion 120 executing differential rotation control to make a clutch input rotation number agree with a clutch output rotation number in the inertia phase; and an estimation portion 113 acquiring the clutch torque and the hydraulic pressure over a prescribed period of the inertia phase, operating their average value and dispersion, calculating correlation data of the clutch torque and the hydraulic pressure by using the average value in which the operated dispersion is a prescribed threshold value or less, and estimating the clutch characteristic of the clutches 21, 22 on the basis of the calculated correlation data.SELECTED DRAWING: Figure 4

Description

The present disclosure relates to an estimation device and an estimation method, and more particularly to a technique for estimating the clutch characteristics (for example, friction coefficient) of a clutch device that transmits the rotational power of a driving force source to a transmission.

In the power transmission device mounted on the vehicle, the clutch transmission torque transmitted from the driving force source to the transmission via the clutch device during the shift is appropriately controlled to generate a shift shock and extend the shift time. Is effectively suppressed.

For example, Patent Document 1 describes a technique in which the number of rotations of a friction member of a clutch device and the coefficient of friction of the clutch are kept constant, and the acting force on the friction member is determined from the difference between the transmitted torque of the friction member and the target torque. It is disclosed.

Japanese Unexamined Patent Publication No. 9-249051

In the technique described in the above document, the clutch friction coefficient is constant, but the actual clutch friction coefficient changes as the friction member deteriorates over time. Therefore, in order to effectively suppress a shift shock or the like when the clutch is replaced, it is desired to appropriately control the clutch transmission torque by appropriately acquiring the clutch characteristics such as the clutch friction coefficient.

In order to acquire the clutch characteristics, it is conceivable to store the clutch transmission torque and the supply oil pressure at the time of clutch replacement, and to estimate and calculate the clutch characteristics from these correlation data. However, if uncorrelated data including the influence of noise or the like on the clutch transmission torque and the supply oil pressure is used for the estimation, there is a problem that the estimation accuracy of the clutch characteristics is lowered.

The technique of the present disclosure has been made in view of the above circumstances, and an object of the present disclosure is to effectively improve the estimation accuracy of the clutch characteristics.

The apparatus of the present disclosure is an estimation device of the friction coefficient of the clutch device that transmits the rotational power of the driving force source to the transmission, and is an input for acquiring the clutch input rotation speed input from the driving force source to the clutch device. The number of times acquisition means, the output rotation speed acquisition means for acquiring the clutch output rotation speed output from the clutch device to the transmission, and the clutch from the torque phase to the inertia phase generated in the shift progress process of the transmission. The difference between the clutch control means that operates the device with a predetermined supply hydraulic pressure to execute the clutch replacement control for switching from the released state to the engaged state and the clutch input rotation speed that matches the clutch output rotation speed in the inertia phase. The rotation speed control means for executing the rotation control, the clutch torque and the supply hydraulic pressure of the clutch device are acquired over a predetermined period of the inertia phase, and the average value and the dispersion of the acquired clutch torque and the supply hydraulic pressure are obtained. Each calculation is performed, and the correlation data between the clutch torque and the supply hydraulic pressure is calculated using the average value of the calculated dispersion equal to or less than a predetermined threshold value, and the clutch characteristics of the clutch device are calculated based on the calculated correlation data. It is characterized by comprising a clutch characteristic estimation means for estimating.

Further, the rotation speed control means executes constant rotation speed control for maintaining the clutch input rotation speed at a constant rotation speed over the predetermined period of the inertia phase, and the clutch characteristic estimation means performs the rotation speed. It is preferable to acquire the output torque of the driving force source input to the clutch device as the clutch torque during execution of the constant number control.

Further, the clutch characteristic estimation means creates the correlation data by setting the clutch torque on the vertical axis and the supply hydraulic pressure on the horizontal axis, and obtains a slope value of an approximate straight line obtained from the correlation data. It is preferable to estimate and calculate the friction coefficient of the clutch device by dividing the inclination value by the effective clutch radius of the clutch device, the number of clutch plates, and the piston pressure receiving area.

Further, the spring reaction force for estimating and calculating the spring reaction force of the return spring of the clutch device is calculated by obtaining the intercept value of the approximate straight line and dividing the value obtained by multiplying the intercept value by the piston pressure receiving area by the inclination value. It is preferable to further provide a force estimation means.

The method of the present disclosure is a method of estimating the friction coefficient of a clutch device that transmits the rotational power of a driving force source to a transmission, and the clutch is from the torque phase to the inertia phase generated in the shifting progress process of the transmission. The clutch replacement control is executed by operating the device with a predetermined supply hydraulic pressure to switch from the released state to the engaged state, and in the inertia phase, the clutch input rotation speed input from the driving force source to the clutch device is set to the clutch. The differential rotation control that matches the clutch output rotation speed output from the device to the transmission is executed, and the clutch torque and the supply hydraulic pressure of the clutch device are acquired over a predetermined period of the inertia phase, and the acquired clutch is obtained. The average value and the dispersion of the torque and the supply hydraulic pressure are calculated, respectively, and the correlation data between the clutch torque and the supply hydraulic pressure is calculated and calculated using the average value of the calculated dispersion being equal to or less than a predetermined threshold value. It is characterized in that the clutch characteristics of the clutch device are estimated based on the correlation data.

According to the technique of the present disclosure, the estimation accuracy of the clutch characteristic can be effectively improved.

It is a schematic block diagram which shows the power transmission device mounted on the vehicle which concerns on this embodiment. It is a schematic diagram which shows an example of the correlation data calculated by the friction coefficient estimation part which concerns on this embodiment. It is a timing chart diagram explaining the clutch friction coefficient and the spring reaction force estimation process at the time of upshifting which switches the 2nd clutch from the disengaged state to the engaged state while changing the 1st clutch from the engaged state to the disengaged state. It is a flowchart explaining the flow of the clutch friction coefficient and the spring reaction force estimation process which concerns on this Embodiment.

Hereinafter, the estimation device and the estimation method according to the present embodiment will be described with reference to the accompanying drawings. The same parts have the same reference numerals, and their names and functions are also the same. Therefore, detailed explanations about them will not be repeated.

[overall structure]
FIG. 1 is a schematic configuration diagram showing a power transmission device mounted on the vehicle 1 according to the present embodiment.

The vehicle 1 is equipped with an engine 10 as an example of a driving force source. The crankshaft 11 of the engine 10 is connected to the first and second transmission input shafts 31 and 32 of the transmission mechanism 30 (transmission) via the dual clutch device 20 (clutch device). A propeller shaft connected to the left and right drive wheels (not shown) via a differential gear device or the like is connected to the transmission output shaft 33 of the transmission mechanism 30.

The dual clutch device 20 has a first clutch 21 and a second clutch 22.

The first clutch 21 is, for example, a wet multi-plate clutch, and is a clutch hub 23 rotatably provided on the crankshaft 11 and a first clutch rotatably provided on the first transmission input shaft 31. The drum 24, the first clutch plate 25 in which a plurality of friction plates and separate plates are alternately arranged, the first piston 26 for pressing the first clutch plate 25, the first hydraulic chamber 26A, and the first return spring 26B. And have. A friction member (not shown) is attached to the friction plate of the first clutch plate 25.

In the first clutch 21, the first piston 26 is on the output side (right in FIG. 1) due to the pressure (hydraulic pressure) of the hydraulic oil supplied from the oil supply circuit 70 to the first hydraulic chamber 26A in response to a command from the control device 100. When the stroke is moved in the (direction) direction, the first clutch plate 25 is pressure-contacted to enter an engaged state (contact state) in which torque is transmitted.

On the other hand, when the hydraulic pressure of the first hydraulic chamber 26A is released in response to a command from the control device 100, the first piston 26 strokes to the input side (to the left in FIG. 1) by the urging force of the first return spring 26B. By moving, the first clutch 21 is in an released state (disengaged state) in which power transmission is cut off.

The second clutch 22 is, for example, a wet multi-plate clutch, which includes a clutch hub 23, a second clutch drum 27 rotatably provided on the second transmission input shaft 32, a plurality of friction plates, and a separate clutch. It includes a second clutch plate 28 in which plates are alternately arranged, a second piston 29 that press-contacts the second clutch plate 28, a second hydraulic chamber 29A, and a second return spring 29B. A friction member (not shown) is attached to the friction plate of the second clutch plate 28.

When the second clutch 22 strokes the second piston 29 to the output side (to the right in FIG. 1) by the hydraulic pressure supplied from the oil supply circuit 70 to the second hydraulic chamber 29A in response to a command from the control device 100. When the second clutch plate 28 is pressed against each other, it is in an engaged state (contact state) in which torque is transmitted.

On the other hand, when the hydraulic pressure of the second hydraulic chamber 29A is released in response to a command from the control device 100, the second piston 29 strokes to the input side (to the left in FIG. 1) by the urging force of the second return spring 29B. By moving, the second clutch 22 is in an released state (disengaged state) in which power transmission is cut off.

The oil supply circuit 70 includes an oil strainer 72 immersed in hydraulic oil in an oil pan 71, a main supply line 73 connected to the oil strainer 72, and first and second supply lines 74 branching from the main supply line 73. , 75 and. Further, the main supply line 73 is provided with an oil pump OP driven by the power of the engine 10.

The first supply line 74 supplies hydraulic oil to the first hydraulic chamber 26A. The first supply line 74 is provided with a first solenoid valve 76 that controls the supply oil pressure to the first hydraulic chamber 26A. The second supply line 75 supplies hydraulic oil to the second hydraulic chamber 29A. The second supply line 75 is provided with a second solenoid valve 77 that controls the supply oil pressure to the second hydraulic chamber 29A. The operation of the first and second solenoid valves 76 and 77 is controlled by being energized in response to a command from the control device 100.

The transmission mechanism 30 includes an auxiliary transmission unit 40 arranged on the input side and a main transmission unit 50 arranged on the output side. Further, the transmission mechanism 30 includes a first transmission input shaft 31 and a second transmission input shaft 32 provided in the auxiliary transmission unit 40, a transmission output shaft 33 provided in the main transmission unit 50, and each of these shafts. It is provided with a sub-shaft 34 arranged in parallel with 31 to 33. The first transmission input shaft 31 is inserted into a hollow shaft that penetrates the second transmission input shaft 32 in the axial direction so as to be relatively rotatable.

The auxiliary transmission unit 40 is provided with a first splitter gear pair 41 and a second splitter gear pair 42. The first splitter gear pair 41 is provided with the first input main gear 43 rotatably provided on the first transmission input shaft 31 and the first input main gear 43 rotatably provided on the sub-shaft 34. It is provided with a first input auxiliary gear 44 that constantly bites teeth. The second splitter gear pair 42 is provided with a second input main gear 45 rotatably provided on the second transmission input shaft 32 and a second input main gear 45 rotatably provided on the sub-shaft 34. It is provided with a second input auxiliary gear 46 that constantly bites teeth.

The main transmission 50 is provided with a plurality of output gear pairs 51 and a plurality of synchronization mechanisms 55. Each output gear pair 51 includes an output sub gear 52 that is integrally rotatable on the sub shaft 34 and an output main gear 53 that is provided on the output shaft 33 so as to be relatively rotatable and always meshes with the output sub gear 52. It has. Each synchronization mechanism 55 is configured to include a sleeve, a synchronizer ring, a dog gear, and the like (not shown).

The operation of the synchronization mechanism 55 is controlled by the control device 100, and the speed change shifter 85 shifts the sleeve of the synchronization mechanism 55 according to the running state of the vehicle 1 and the operating state of the engine 10, thereby shifting the transmission. The output shaft 33 and the output main gear 53 are selectively switched to an engaged state (gear-in state) or a non-engaged state (neutral state). The number of output gear pairs 51, the number of synchronization mechanisms 55, the arrangement pattern, and the like are not limited to the illustrated examples, and can be appropriately changed without departing from the gist of the present disclosure.

In the present embodiment, the auxiliary transmission unit 40 is set so that the gear ratio of the first splitter gear to 41 is smaller than that of the second splitter gear to 42. That is, when the second clutch 22 is engaged and the driving force is transmitted from the second splitter gear pair 42 to the main transmission 50, the speed can be set to the low speed side (odd number stage), and the first clutch 21 is engaged. When the driving force is transmitted from the first splitter gear pair 41 to the main transmission unit 50, the high speed side (even number of stages) can be set.

The vehicle 1 is provided with various sensors 90 to 94. The sensor values of these various sensors 90 to 94 are transmitted to the electrically connected control device 100.

The engine rotation speed sensor 90 (an example of the input rotation speed acquisition means) acquires the _{rotation speed of the engine 10 per unit time (hereinafter, engine rotation speed ω e) from the crankshaft 11.} The accelerator opening sensor 91 acquires the fuel injection amount Q (injection instruction value) of the engine 10 according to the depression amount of the accelerator pedal (not shown).

The vehicle speed sensor 92 acquires the vehicle speed V of the vehicle 1 from the transmission output shaft 33 (or the propeller shaft). The vehicle speed sensor 92 may be a wheel speed sensor. The first input shaft rotation speed sensor 93 (an example of the output rotation speed acquisition means) is the rotation speed per unit time of the first transmission input shaft 31 connected to the first clutch 21 (hereinafter, the first clutch output rotation speed). ω ₁ ) is acquired. The second input shaft rotation speed sensor 94 (an example of the output rotation speed acquisition means) is the rotation speed of the second transmission input shaft 32 connected to the second clutch 22 per unit time (hereinafter, the second clutch output rotation speed). ω ₂ ) is acquired.

[Control device]
The control device 100 controls various controls such as the engine 10, the dual clutch device 20, and the transmission mechanism 30, and includes a CPU (Central Processing Unit), a ROM (Read Only Memory), a RAM (Random Access Memory), an input port, and an output. It is configured with a port and the like.

Further, the control device 100 includes an automatic shift control unit 110, a clutch control unit 112 (clutch control means), a difference rotation control unit 120 (rotation speed control means), and a friction coefficient estimation unit 113 (clutch characteristic estimation means). , A spring reaction force estimation unit 114 (spring reaction force estimation means) is provided as a part of the functional elements. Although these functional elements will be described as being included in the control device 100 which is integrated hardware in the present embodiment, any part of them may be provided in separate hardware.

The automatic shift control unit 110 executes automatic shift control for shifting up or down the shift mechanism 30 to an appropriate shift stage based on the operating state of the engine 10 or the running state of the vehicle 1. More specifically, the memory of the control device 100 stores a shift change map (not shown) that is referred to based on the fuel injection amount Q and the vehicle speed V. The automatic shift control unit 110 identifies an appropriate gear stage by referring to the shift change map based on each sensor value transmitted from the accelerator opening sensor 91 and the vehicle speed sensor 92, and operates the shift shifter 85. , The transmission mechanism 30 is automatically shifted to an appropriate gear stage.

The automatic transmission control unit 110 maintains the currently established power transmission path of the main transmission unit 50 when the current gear stage is shifted up from an odd number stage to an even number stage due to the fulfillment of the shift up request (current gear). (Maintaining the synchronization mechanism 55 corresponding to the step in the engaged state), the clutch control unit 112 is instructed to switch the second clutch 22 from the engaged state to the disengaged state and the first clutch 21 from the disengaged state to the engaged state. Send a signal.

Similarly, the automatic transmission control unit 110 maintains the currently established power transmission path of the main transmission unit 50 when the current gear stage is downshifted from an even numbered stage to an odd numbered stage due to the fulfillment of the downshift request. At the same time, an instruction signal for switching the first clutch 21 from the engaged state to the disengaged state and the second clutch 22 from the disengaged state to the engaged state is transmitted to the clutch control unit 112.

On the other hand, when the automatic shift control unit 110 shifts up the current gear from an even-numbered stage to an odd-numbered stage due to the fulfillment of the shift-up request, the automatic shift control unit 110 engages the synchronization mechanism 55 corresponding to the next gear stage. While pre-shifting the main transmission 50 to establish the power transmission path for the next gear, the clutch control unit 112 engages the first clutch 21 from the engaged state and the second clutch 22 from the released state. Sends an instruction signal to switch to the matching state.

Similarly, when the automatic shift control unit 110 shifts down the current gear from an odd-numbered stage to an even-numbered stage due to the fulfillment of the shift-down request, the automatic shift control unit 110 engages the synchronization mechanism 55 corresponding to the next gear stage. Then, while pre-shifting the main transmission unit 50 to establish the power transmission path of the next gear stage, the clutch control unit 112 is in the disengaged state of the second clutch 22 and in the disengaged state of the first clutch 21. An instruction signal for switching to the engaged state is transmitted.

The clutch control unit 112 performs clutch replacement control for switching engagement / disengagement of the first clutch 21 and the second clutch 22 in response to a command transmitted from the automatic transmission control unit 110. In the present embodiment, the clutch control unit 112 has a first or second hydraulic chamber so that the transmission torque of the first or second clutches 21 and 22 that switch from the released state to the engaged state becomes a desired clutch transmission torque Tc. The supply hydraulic pressure Pc to 26A and 29A (the amount of electricity supplied to the first or second solenoid valves 76 and 77) is controlled.

Specifically, the clutch control unit 112 has a clutch friction coefficient μ read from the friction coefficient map M1 and a spring reaction force map over a period from the start of the torque phase in which the shift request is satisfied to the end of the inertia phase. By substituting the spring reaction force _Friction read from M2 into the following formula (1), the torque transmitted from the first or second clutches 21 and 22 that can be switched to the engaged state to the transmission mechanism 30 is the desired clutch transmission torque. _{The supply hydraulic pressure Pcmd} (indicated value) to the first or second hydraulic chambers 26A and 29A is adjusted so as to have Tc.

Tc = μRN (AP _cmd −F _prompt ) ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ (1)
In formula (1), R is the effective radius of the clutch plates 25 and 28, N is the number of clutch plates 25 and 28, and A is the pressure receiving area of the pistons 26 and 29.

Here, the "torque phase" is one of the shifting processes that occur during the progress of automatic transmission, and the clutches 21 and 22 of the current gear stage gradually shift from the engaged state to the released state, and the next gear. A phase in which the clutches 21 and 22 of the stage gradually shift from the released state to the engaged state. Further, the "inertia phase" is one of the shifting processes that occur during the progress of automatic shifting, and the clutches 21 and 22 on the release side are completely released and the clutches 21 and 22 on the engaging side are in a slipped state. In the meantime, the engine speed ω _e is lowered in the case of upshifting, while the engine speed ω _e is increased in the case of downshifting.

In the inertia phase after the torque phase, the differential rotation control unit 120 has the engine rotation speed ω _{e and the clutch output rotation speeds ω 1} , ω _{2 of} the first or second clutches 21 and 22 that are switched from the released state to the engaged state. The differential rotation control for lowering or increasing the engine rotation speed ω _e is executed so that the actual difference rotation speed Δω (= ω _e −ω ₁ , ω ₂ ) becomes the target difference rotation speed Δω _ref.

Specifically, the differential rotation control unit 120 has _{clutch output rotation speeds ω 1} , ω _{2 of} the first or second clutches 21 and 22 that can switch _{the engine rotation speed ω e from the released state to the engaged state in the inertia phase.} Set the target difference rotation speed Δω _{ref that gradually matches with.} Then, the difference rotation control unit 120 has the target difference rotation speed Δω _{ref with the clutch output rotation speeds ω 1} and ω _{2 of} the first or second clutches 21 and 22 that can be switched to the engaged state with the engine rotation speed ω _e . Feedback control (for example, PID control) is applied to the deviation e (= Δω _ref −Δω) obtained by subtracting the actual difference rotation speed Δω (= ω _e −ω ₁ , ω _{2), and the set engine torque is set.} The fuel injection instruction value for driving the engine 10 by Te is transmitted to an injector (not shown) of the engine 10 to be controlled.

In the present embodiment, when the predetermined learning condition for estimating the clutch friction coefficient μ, which will be described later, is satisfied, the differential rotation control unit 120 _{maintains the engine rotation speed ω e} at a constant rotation speed for a predetermined period during the differential rotation control. Execute constant rotation speed control (constant rotation speed control).

The friction coefficient estimation unit 113 and the spring reaction force estimation unit 114 have the engine torque Te (= clutch transmission torque Tc) acquired over the period during which the constant rotation speed control is executed during the inertia phase, and the first or first second hydraulic chamber 26A, on the basis of the correlation data between the supply pressure _{P cmd} to 29A, the clutch friction coefficient of the clutch plates 25 and 28 of the first and second clutches 21, 22 mu and respective return springs 26B, 29B The spring reaction force F _rtun of each is estimated.

Specifically, first, the estimation units 113 and 114 obtain the engine torque Te (= clutch transmission torque Tc) acquired during the constant control of the rotational speed in the inertia phase, and the average value and dispersion of _{the supply oil pressure Pcmd, respectively.} Calculate. Further, the estimation units 113 and 114 use the average value as the correlation data when the variance is equal to or less than the predetermined threshold value, while the estimation units 113 and 114 include the influence of noise or the like when the variance is larger than the predetermined threshold value. Therefore, the average value is processed so as not to be used for the correlation data. Further, the estimation units 113 and 114 perform data selection (permission / disapproval) based on the above-mentioned variance for each shift, and sequentially perform a least squares method or the like using the average value of the permitted correlation data. As a result, the clutch characteristics (clutch friction coefficient μ, spring reaction force _Frtun ) are estimated and calculated.

Hereinafter, a more detailed clutch characteristic estimation process will be described. Since the estimation of the clutch friction coefficient μ of the first and second clutches 21 and 22 and _{the estimation of the spring reaction force F rtun} have the same processing contents, the estimation processing of the second clutch 22 will be described below. , The description of the estimation process of the first clutch 21 will be omitted.

In the inertia phase, the relational expression between the engine torque Te and the clutch transmission torque Tc is expressed by the following mathematical expression (2).

Te = Tc-Ie × ωe * ・ ・ ・ ・ ・ ・ (2)
In the formula (2), Ie is the engine inertia, the asterisk (*) attached to the upper right of ωe is the time derivative, and ωe * is the engine rotational acceleration.

When constant rotation speed control is executed during the inertia phase, the engine rotation acceleration ωe * becomes zero in the mathematical formula (2). That is, during constant rotation speed control, the engine torque Te input from the engine 10 to the second clutch 22 and the clutch transmission torque Tc transmitted from the second clutch 22 to the transmission mechanism 30 become equal (Te = Tc). , The above formula (1) is represented by the following formula (3).

Te = Tc = μRN (AP _{cmd- Frtun} )
= ΜRNAP _cmd −μRNF _run
= AP _cmd + b ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ (3)
However, a = μRNA, b = -μRNF _rtun
In formula (3), μ is the clutch friction coefficient of the second clutch 22, P _cmd is the indicated value of the supply oil supply to the second hydraulic chamber 29A (or the indicated value of the amount of electricity supplied to the second solenoid valve 77), and R is the indicated value. The effective radius of the second clutch plate 28, N is the number of the second clutch plates 28, A is the pressure receiving area of the second piston 29, and _Frtrn is the spring reaction force of the second return spring 29B. The supply oil pressure _Pcmd is not limited to the indicated value, and a sensor value or the like of a hydraulic pressure sensor (not shown) provided in the oil supply circuit 70 may be used.

If the equation (3) is regarded as a linear straight line, the inclination "a" and the section "b" of the equation (3) can be obtained by obtaining the correlation between _{the supply oil pressure Pcmd and the engine torque Te (clutch transmission torque Tc).} Furthermore, the clutch friction coefficient μ and the spring reaction force _Frtrn can be calculated, respectively.

The friction coefficient estimation unit 113 processes the engine torque Te (= Tc) obtained during the execution period of the constant rotation speed control for each shift and the supply oil pressure _Pcmd for each accelerator opening, and processes the average value and dispersion of these for each accelerator opening. Calculate. Then, the friction coefficient estimation unit 113 calculates these correlation data using the average value of the variance acquired for each shift and is equal to or less than a predetermined threshold value, and stores the correlation data in the memory of the control device 100.

FIG. 2 is an example of the correlation data D calculated by the friction coefficient estimation unit 113, in which the supply oil pressure _Pcmd is set on the x-axis and the engine torque Te (= clutch transmission torque Tc) is set on the y-axis. The friction coefficient estimation unit 113 _{calculates an approximate straight line L by linearly approximating the correlation data D of the supplied hydraulic pressure Pcmd} and the engine torque Te (= clutch transmission torque Tc) by, for example, the least squares method. The clutch friction coefficient μ of the second clutch 22 is estimated and calculated by obtaining the slope “a” of the approximate straight line L and the section “b” with the y-axis and substituting it into the following equation (4). The correlation data D does not necessarily have to be graphed and may be processed as numerical data. Further, the calculation method of the approximate straight line L is not limited to the least squares method, and various calculation methods such as other linear regression methods can be used.

μ = a / RNA ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ ・ (4)
In the formula (4), R is the effective radius of the second clutch plate 28, N is the number of the second clutch plates 28, and A is the pressure receiving area of the second piston 29.

The friction coefficient estimation unit 113 transmits the estimated calculated clutch friction coefficient μ to the friction coefficient map M1 and sequentially updates the friction coefficient map M1 based on the clutch friction coefficient μ.

The spring reaction force estimation unit 114 substitutes the slope "a" and the intercept "b" of the approximate straight line L obtained by linearly approximating the above-mentioned correlation data D into the following equation (5). 2 The spring reaction force F _rtun of the return spring 29B is estimated and calculated.

F _rtun = -b / μRN = -Ab / a ... (5)
In the formula (5), R is the effective radius of the second clutch plate 28, N is the number of the second clutch plates 28, and A is the pressure receiving area of the second piston 29.

The spring reaction force estimation unit 114 _{transmits the estimated calculated spring reaction force F rtun} to the spring reaction force map M2, and sequentially updates the spring reaction force map M2 based on the spring reaction _{force F rtun.}

_{FIG. 3 shows} the estimation processing of the clutch friction coefficient μ and the spring reaction force Frtrn at the time of upshifting when the second clutch 22 is switched from the disengaged state to the engaged state while the first clutch 21 is released from the engaged state. It is a timing chart diagram to be described.

In FIG. 3, (A) shows _{changes in the engine rotation speed ω e} , the first clutch output rotation speed ω _1, and the second clutch output rotation speed ω ₂ , and (B) shows the clutch friction coefficient μ and the spring reaction. The changes in the engine torque Te and the driver required torque Td when the force F _rtrn is estimated, and (C) is the engine when the clutch friction coefficient μ and the spring reaction force F _rtrn are not estimated. The changes in the torque Te and the driver required torque Td are shown, respectively. Further, the times T0 to T1 indicate the torque phase, the times T1 to T4 indicate the inertia phase, and the times T2 to T3 indicate the period during which the constant rotation speed control is executed during the inertia phase.

When the shift-up request is satisfied at time T0, the clutch replacement control starts releasing the first clutch 21 corresponding to the current gear stage and starts engaging the second clutch 22 corresponding to the next gear stage. Is started.

When the torque phase shifts to the inertia phase at time T1, the actual difference rotation speed Δω (= ω _e −ω _{2 ) between the engine rotation speed ω e} and the clutch output rotation speed ω _{2 of} the second clutch 22 becomes the target difference rotation. The differential rotation control for lowering the engine rotation speed ω _e is started so that the speed Δω _{ref is obtained.}

When a predetermined learning condition is satisfied at time T2, constant rotation speed control for maintaining the engine rotation speed ω _{e at a predetermined value is started in order to estimate the clutch friction coefficient μ and the spring reaction force Frtrn.} .. The predetermined learning condition is, for example, satisfied when a difference ΔTe occurs in the current engine torque Te with respect _{to the reference engine torque Te _ST} acquired in advance in the initial state before the clutch friction coefficient μ changes. It should be.

When the constant rotation speed control is started at time T2, the friction coefficient estimation unit 113 sets the engine torque Te (= Tc) obtained during the execution period of the constant rotation speed control and the supply hydraulic pressure _Pcmd for each accelerator opening. It processes and stores each torque required by the driver.

At time T3, the rotational speed constant control is finished, the friction coefficient estimation unit 113, the engine torque Te obtained over a time T2 to T3 (= Tc) and calculates the average value and the variance of the supply pressure _{P cmd} .. Further, if the variance is equal to or less than a predetermined threshold value, the friction coefficient estimation unit 113 uses the average value as the correlation data, and uses the slope “a” and the section “b” of the approximate straight line L obtained from the correlation data D as the correlation data. By substituting into the above mathematical formula (4), the coefficient of friction μ of the second clutch 22 is estimated and calculated. Further, in parallel with this, the spring reaction force estimation unit 114 substitutes the inclination "a" and the intercept "b" into the above equation (5), so that the spring reaction force _{Frtun of the second return spring 29B} Is estimated.

At time T4, when the engine rotation speed ω _e and the clutch output rotation speed ω ₂ match, the shift control is terminated.

As shown in FIG. 3C, when the clutch friction coefficient μ and the spring reaction force _{Tort are not estimated during the inertia phase, the engine speed ω e} and the clutch output rotation are performed at time T4. After the speed ω ₂ matches, the engine torque Te deviates from the driver required torque Td (torque step), which causes a shift shock. In addition, the driver may cause an unintended change in acceleration.

On the other hand, in the present embodiment in which the clutch friction coefficient μ and the spring reaction force _{Frtrn are estimated during the inertia phase, the engine rotation speed ω e} is the clutch output as shown at time T4 in FIG. 3 (B). When the rotation speed ω ₂ is matched, the engine torque Te also substantially matches the driver required torque Td. That is, it is possible to effectively suppress the occurrence of shift shock.

FIG. 4 is a flowchart illustrating a flow of estimation processing of the clutch friction coefficient μ and the spring reaction force _{Frtrn according to the present embodiment.} This routine is executed at the start of shift control (see time T0 in FIG. 3).

In the torque phase (see time T0 to T1 in FIG. 3) from the start of shift control to the transition to the inertia phase, the first clutch 21 corresponding to the current gear stage is shifted to the released state and the next gear. Clutch replacement control for shifting the second clutch 22 corresponding to the stage to the engaged state is executed.

In step S100, it is determined whether or not the torque phase has shifted to the inertia phase. This determination may be performed, for example, by detecting slip based on the difference in input / output rotation speed of the speed change mechanism 30 and the like. If it is determined in step S100 that the phase has shifted to the inertia phase (Yes), the process proceeds to step S110. On the other hand, when it is determined that the inertia phase has not been started (No), the determination process of step S100 is repeated.

In step S110, the target difference rotation speed Δω _ref is set, and the actual difference rotation speed Δω (= ω _e −ω ₂ ) between the engine rotation speed ω _e and the second clutch output rotation speed ω ₂ is the target difference rotation speed Δω. _The difference rotation control that controls the drive of the engine 10 is started based on these deviations e (= Δω _{ref −Δω) so as to be ref.}

In step S120, it is determined whether or not a predetermined learning condition for estimating the clutch friction coefficient μ is satisfied. The predetermined learning conditions include, for example, a case where a difference ΔTe occurs in the current engine torque Te with respect _{to the reference engine torque Te _ST} acquired in the initial state before the clutch friction coefficient μ changes, or the previous learning. If a predetermined time has passed since then, or if it is determined that the shift shock is large based on the acceleration or the size of the jerk, it may be established. When the learning condition is satisfied (Yes), this control proceeds to step S130. On the other hand, if the learning condition is not satisfied (No), this control proceeds to step S200.

In step S130, the constant rotation speed control is started, and in step S140, the engine torque Te (= Tc) and the supply oil pressure _Pcmd are acquired and stored during the execution period of the constant rotation speed control.

In step S150, it is determined whether or not the constant rotation speed control is completed. When the constant rotation speed control is completed, this control proceeds to step S160, and calculates the engine torque Te (= Tc) acquired during the execution of the constant rotation speed control, and the average value and variance of _{the supply oil pressure Pcmd, respectively.}

In step S170, it is determined whether or not the variance is equal to or less than a predetermined threshold value. When the variance is equal to or less than a predetermined threshold value (Yes), the control proceeds to step S180, and the average value calculated in step S160 is stored as the correlation data D. On the other hand, when the variance is larger than the predetermined threshold value (No), this control proceeds to step S200 without using the average value calculated in step S160 for the correlation data D.

In step S190, the slope “a” and the intercept “b” of the approximate straight line L of the correlation data D are calculated and substituted into the above mathematical formula (4) to estimate the clutch friction coefficient μ of the second clutch 22. Further, by substituting the slope "a" and the intercept "b" into the above mathematical formula (5), the spring reaction force _Frtun of the second return spring 29B is estimated and calculated.

In step S195, the maps M1 and M2 are updated based on the estimated clutch friction coefficient μ and the spring reaction force _Frtun.

In step S200, the differential rotation control is continuously executed while using the updated maps M1 and M2, and in step S210, it is determined whether or not _{the engine rotation speed ω e} matches the second clutch output rotation speed ω _2. judge. When the engine rotation speed ω _e does not match the second clutch output rotation speed ω ₂ (No), this control repeats the processes of steps S200 and S210. On the other hand, when the engine rotation speed ω _e matches the second clutch output rotation speed ω ₂ (Yes), this control proceeds to step S220, ends the differential rotation control, and then returns.

According to the present embodiment described in detail above, during the inertia phase, the engine torque Te equal to the clutch transmission torque Tc and the supply hydraulic pressure _Pcmd are acquired by constant rotation speed control, and the average value and dispersion of these are obtained, respectively. The slope "a" of the approximate straight line L obtained from the correlation data D by calculating the correlation data between the _{engine torque Te (Tc) and the supply hydraulic pressure Pcmd} by calculating and using the average value whose dispersion is equal to or less than a predetermined threshold value. The clutch friction coefficient μ and the spring reaction force _Frtun are estimated and calculated based on the section “b”. _{As a result, the clutch friction coefficient μ and the spring reaction force F rtun} can be effectively estimated based on the data in which the dispersion that does not include (or has a small effect) the influence of noise or the like is equal to or less than a predetermined threshold value. It is possible to surely improve the estimation accuracy of the clutch characteristics. In addition, by improving the estimation accuracy of the clutch characteristics, the controllability of the engine torque Te and the clutch transmission torque Tc will surely be improved, and the occurrence of shift shock and the extension of the shift time will be effectively prevented. Will also be possible. Further, by performing the estimation process when the engine rotation speed ω _e is constant, it is possible to remove the error of the engine inertia and the like.

The present disclosure is not limited to the above-described embodiment, and can be appropriately modified and implemented without departing from the spirit of the present disclosure.

For example, in the above embodiment, it has been described that the engine torque Te equal to the clutch transmission torque Tc is acquired during the execution period of the constant rotation speed control, but when the engine inertia Ie of the above formula (2) is obtained, , The clutch transmission torque Tc may be calculated and acquired from the above equation (2) without executing the constant rotation speed control.

Further, in the above embodiment, the clutch for connecting and disconnecting the power between the engine 10 and the transmission mechanism 30 has been described by taking the dual clutch device 20 as an example, but the clutch device is a single clutch device and an AT having a plurality of clutches and brakes. It may be a device.

Further, although the clutch output rotation speeds ω ₁ and ω ₂ of the clutches 21 and 22 have been described as being acquired by the input shaft rotation speed sensors 93 and 94, the gear ratio of the speed change mechanism 30 is added to the sensor value of the vehicle speed sensor 92. It may be obtained by multiplying.

Further, although the vehicle 1 has been described as having the engine 10 as a driving force source, it may be a vehicle having a driving force source other than the engine 10 such as a hybrid vehicle in which the engine 10 and the motor are used in combination.

Further, in the above-described embodiments, the estimated clutch friction coefficient μ and the spring reaction force F _Rtun to an example, not limited to this, the correlation of the clutch characteristic (e.g., a clutch transmission torque Tc and the supply pressure P _cmd It may be configured to estimate the relationship).

1 vehicle 10 engine (driving force source)
11 Crankshaft 20 Dual clutch device (clutch device)
21 1st clutch 22 2nd clutch 30 Transmission mechanism (transmission)
31 1st transmission input shaft 32 2nd transmission input shaft 90 Engine speed sensor (input speed acquisition means)
91 Accelerator opening sensor 92 Vehicle speed sensor 93 First input shaft rotation speed sensor (output rotation speed acquisition means)
94 Second input shaft rotation speed sensor (output rotation speed acquisition means)
100 Control device 110 Automatic transmission control unit 112 Clutch control unit (clutch control means)
113 Friction coefficient estimation unit (clutch characteristic estimation means)
114 Spring reaction force estimation unit (spring reaction force estimation means)
120 Differential rotation control unit (rotation speed control means)

Claims (5)

It is an estimation device of the friction coefficient of the clutch device that transmits the rotational power of the driving force source to the transmission.
An input number acquisition means for acquiring the clutch input rotation speed input to the clutch device from the driving force source, and
An output rotation speed acquisition means for acquiring the clutch output rotation speed output from the clutch device to the transmission, and
A clutch control means for executing clutch replacement control for switching from an released state to an engaged state by operating the clutch device with a predetermined supply oil pressure from a torque phase to an inertia phase generated in the shifting progress process of the transmission.
In the inertia phase, a rotation speed control means for executing differential rotation control for matching the clutch input rotation speed with the clutch output rotation speed, and
The clutch torque and the supply oil pressure of the clutch device are acquired over a predetermined period of the inertia phase, and the average value and the dispersion of the acquired clutch torque and the supply oil pressure are calculated respectively, and the calculated dispersion is a predetermined value. A clutch characteristic estimation means for calculating the correlation data between the clutch torque and the supply oil pressure using an average value equal to or less than a threshold value and estimating the clutch characteristics of the clutch device based on the calculated correlation data is provided. An estimation device characterized by. The rotation speed control means executes constant rotation speed control for maintaining the clutch input rotation speed at a constant rotation speed over the predetermined period of the inertia phase.
The estimation device according to claim 1, wherein the clutch characteristic estimation means acquires the output torque of the driving force source input to the clutch device as the clutch torque during execution of the constant rotation speed control. The clutch characteristic estimation means creates the correlation data by setting the clutch torque on the vertical axis and the supply hydraulic pressure on the horizontal axis, obtains a slope value of an approximate straight line obtained from the correlation data, and obtains the slope value. The estimation device according to claim 1 or 2, wherein the friction coefficient of the clutch device is estimated and calculated by dividing the clutch effective radius of the clutch device, the number of clutch plates, and the piston pressure receiving area. Spring reaction force estimation that estimates and calculates the spring reaction force of the return spring of the clutch device by obtaining the intercept value of the approximate straight line and dividing the value obtained by multiplying the intercept value by the piston pressure receiving area by the inclination value. The estimation device according to claim 3, further comprising means. It is a method of estimating the friction coefficient of the clutch device that transmits the rotational power of the driving force source to the transmission.
From the torque phase to the inertia phase generated in the shift progress process of the transmission, the clutch replacement control is executed by operating the clutch device with a predetermined supply hydraulic pressure to switch from the released state to the engaged state, and in the inertia phase. , The differential rotation control for matching the clutch input rotation speed input from the driving force source to the clutch device with the clutch output rotation speed output from the clutch device to the transmission is executed, and during a predetermined period of the inertia phase. The clutch torque and the supplied hydraulic pressure of the clutch device are acquired over the same period, and the average value and the dispersion of the acquired clutch torque and the supplied hydraulic pressure are calculated, respectively, and the calculated average value of the dispersion is equal to or less than a predetermined threshold value is used. An estimation method characterized in that the correlation data between the clutch torque and the supply hydraulic pressure is calculated, and the clutch characteristics of the clutch device are estimated based on the calculated correlation data.

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