JP3304658B2 - Control device for hydraulically operated transmission - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION

[0001]

BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a control apparatus for a hydraulically operated transmission, and more particularly, to a hydraulically operated transmission for a vehicle in which co-engagement of a release clutch and an engagement clutch is optimally controlled during gear shifting. The present invention relates to a control device for a hydraulically operated transmission, which can accurately determine whether or not the engine speed has increased.

[0002]

2. Description of the Related Art A hydraulically operated transmission has a plurality of frictional engagement elements such as clutches and brakes, and one of the frictional engagement elements is released and the other is engaged to change the engagement state. The gear is changed by switching. At this time, in order to prevent a so-called shift shock, it is necessary to optimally control the amount of co-engagement between the two.

However, it is extremely difficult to optimally control the co-meshing amount. If the co-meshing amount is large, the output shaft torque of the transmission is greatly reduced (so-called pull-in), giving the driver a feeling of emblem. . Conversely, when the co-meshing amount is small and the torque transmission capacity of the friction engagement element is continuously lower than the transmission input torque, the engine speed rapidly increases, so-called blow-up occurs, and a shock is given to the driver, and the driver is shocked. Had the problem that the performance deteriorated.

Therefore, as disclosed in Japanese Patent Application Laid-Open No. 60-201152, an electromagnetic valve such as a duty solenoid or a linear solenoid is used to optimally control the engagement oil pressure of the friction engagement element during shifting. For example, there is a known type that controls a hydraulic pressure supplied to a friction engagement element such that a rate of change of a rotation speed of a member that changes its rotation during a gear shift, such as a transmission input shaft, follows a target value.

[0005]

Even in such a case, it is necessary to accurately determine whether the co-engagement is in an optimally controlled state, in other words, whether or not the blow-up state has occurred. is there.

Incidentally, the torque transmission capacity of a friction engagement element is proportional to its friction coefficient μ. Further, the friction coefficient μ changes according to the differential rotation Δω between the input side and the output side of the friction engagement element. Generally, in a wet friction engagement element, hydraulic oil (ATF) is interposed between the input and output (between the clutch disc and the clutch plate), and the friction force is based on the shear resistance of the hydraulic oil. Relative rotation speed (difference rotation)
The smaller the value of Δω is, the smaller the value is, and the larger the value of Δω is, the larger the value gradually becomes. When Δω reaches a predetermined value, the characteristic becomes constant.

In a shift in which one friction engagement element is disengaged and the other friction engagement element is engaged, the engagement hydraulic pressure of the released friction engagement element decreases in a torque phase at the beginning of the shift. Even if a differential rotation Δω occurs between input and output, the friction coefficient μ starts to increase as described above, so that the torque transmission capacity of the friction engagement element does not change. However, the differential rotation Δω, that is,
When the input shaft rotation speed increases, especially when the input shaft rotation speed increases and exceeds a predetermined level, there is a disadvantage that the driver feels it as a blow-up. Thereafter, when the engagement hydraulic pressure further decreases, the friction coefficient μ becomes constant, so that the torque transmission capacity of the release-side friction engagement element starts to decrease.

Here, consider a state in which the differential rotation Δω increases, but the torque transmission capacity does not decrease due to the increase in the friction coefficient μ. Then, as an index indicating the toughness with respect to the rising speed of the engine rotation, for example, an index of (engagement side clutch torque transmission capacity + disengagement side clutch torque transmission capacity) / input torque−1 (× 100%) is given (hereinafter, referred to as “index”) This is called the "blow margin"
).

In the initial stage of shifting, the torque transmission capacity of the disengagement (off) side clutch does not decrease, and the torque transmission capacity of the engagement (on) side clutch only gradually increases.
Despite the increase in the differential rotation Δω, the blowing margin ratio hardly changes. Then, when the differential rotation Δω exceeds the blowing determination level at a certain point, it is determined that the blowing margin ratio is low for the first time.

As described above, when the differential rotation Δω increases and exceeds the blow-up determination level, it is determined that the blowing margin ratio is low for the first time. That is, although the differential rotation Δω increases and the toughness with respect to the rise of the engine speed is actually reduced, there occurs a state where the blow margin does not change, and it is unknown how the true blow margin is reduced. Met. Therefore, it is not possible to distinguish whether the blowing margin ratio has a sufficient margin and the engagement hydraulic pressure is higher than necessary, or whether the engagement hydraulic pressure is low and the margin ratio is near the limit. It is not possible to optimally perform the pressure reduction control of the hydraulic pressure of the disengagement side frictional engagement element for controlling according to the value, and it is also impossible to optimally control the release timing, thereby reliably reducing the shift shock. I couldn't do that.

Accordingly, an object of the present invention is to optimize the pressure reduction control or the release timing control of the hydraulic pressure of the release-side friction engagement element by detecting the torque transmission capacity in consideration of the friction coefficient of the friction engagement element. Therefore, it is an object of the present invention to provide a control device for a hydraulically operated transmission which can control the co-meshing according to a target value and thereby surely reduce the shift shock.

[0012]

SUMMARY OF THE INVENTION The above present invention in order to solve the object of the claim 1 wherein, with a plurality of frictional engagement elements engaged is supplied with hydraulic oil, the frictional engagement elements Free up one, <br/> Oite the control device for a hydraulically operated transmission engaging the other performs a shift by switching the engagement state, friction during release of the frictional engagement element on the one coefficient
Is written as a predetermined value equal to or greater than the maximum possible value of the friction coefficient.
And disengagement side frictional coefficient rewriting means for changing the driving force transmission capacity at the release of the previous SL hand friction engagement elements, a release-side driving force transmission capacity detecting means for determining by considering the predetermined value, the other
For determining the driving force transmission capacity when the two friction engagement elements are engaged.
Joint drive force transmission capacity detecting means, and the hydraulically operated transmission
Input driving force detection means for detecting a driving force input to the
Driving force transmission capacity when disengaged and driving force transmission capacity when engaged
And a comparing means for comparing the sum of
Based on the comparison result of the comparing means, and as configured and a rotational fluctuation determination means for determining the degree of quality of the increase of the input rotation in the gear shifting of said hydraulically operated transmission.

According to a second aspect of the present invention, the driving force on the engagement side is provided.
The transmission capacity detecting means is provided when the other frictional engagement element is engaged.
The friction coefficient between the input side of the other friction engagement element and the output
The friction coefficient of the engagement side changes according to the differential rotation of the side.
A driving force based on the changed friction coefficient.
It was configured as Ru determine the transmission capacity.

[0014]

In the hydraulically operated transmission according to the first aspect , the friction coefficient at the time of disengagement of one of the friction engagement elements during the gear shift.
Is written as a predetermined value equal to or greater than the maximum possible value of the friction coefficient.
And disengagement side frictional coefficient rewriting means for changing the driving force (torque) transmission capacity during release of the frictional engagement element of the one, the
Release-side driving force transmission capacity detecting means Ru determined by considering the predetermined value
And the driving force transmission capacity when the other frictional engagement element is engaged.
Engagement-side driving force transmission capacity detection means for determining
Input driving force detection that detects the driving force input to the dynamic transmission
Informing means, drive force transmission capacity at the time of release and drive at the time of engagement
Comparing means for comparing the sum of the force transmission capacities with the input driving force
When, and based-out of a comparison result of the comparing means, the hydraulic
And a rotation fluctuation determining means for determining whether the degree of increase of the input rotation at the time of shifting of the operation type transmission is good or not. Δ
As the torque transmission capacity of the disengagement clutch decreases in proportion to the increase in ω, the blowing margin also decreases, and the differential rotation Δω
Despite the increase, the state in which the blowing margin does not decrease disappears, and it is possible to eliminate the inconvenience that the input rotation increases beyond a predetermined level and the engine rotation speed is felt up. it can. Further, by setting the friction coefficient to the maximum value, it is possible to determine the degree of increase in the input rotation, and it is possible to optimally perform the pressure reduction control of the hydraulic pressure of the disengagement-side friction engagement element or the control of the release timing. Accordingly, the co-meshing can be controlled according to the target value, so that the shift shock can be reliably reduced.

According to a second aspect of the present invention, the driving force on the engagement side is provided.
The transmission capacity detecting means is provided when the other frictional engagement element is engaged.
The friction coefficient between the input side of the other friction engagement element and the output
The friction coefficient of the engagement side changes according to the differential rotation of the side.
A driving force based on the changed friction coefficient.
Since it is configured as Ru determined transmission capacity, it is possible to eliminate a disadvantage that more effectively, input rotation is experience racing of the engine speed by increasing above a predetermined level, the release side This makes it possible to optimally control the pressure reduction or release timing of the hydraulic pressure of the frictional engagement element, thereby controlling the co-meshing according to the target value and reliably reducing the shift shock.

In the above, the rotation fluctuation determining means includes:
More specifically, the determination is made by obtaining an index of (engagement side clutch torque transmission capacity + disengagement side clutch torque transmission capacity) / input torque−1 (× 100%).

[0017]

Embodiments of the present invention will be described below with reference to the accompanying drawings.

FIG. 1 is a schematic diagram generally showing a control device for a hydraulically operated transmission according to the present invention.

In the following, an automatic transmission T for a vehicle will be described.
A main shaft MS connected to a crankshaft 1 of an internal combustion engine E via a torque converter 2 having a lock-up mechanism L, and a counter shaft CS connected to the main shaft MS via a plurality of gear trains. Prepare.

On the main shaft MS, a main first gear 3, a main second gear 4, a main third gear 5, a main fourth gear 6, and a main reverse gear 7 are supported. On the counter shaft CS, a counter first gear 8 meshing with the main first gear 3, a counter second gear 9 meshing with the main second gear 4, a counter third gear 10 meshing with the main third gear 5, A counter fourth gear 11 meshing with the main fourth gear 6 and a counter reverse gear 12 connected to the main reverse gear 7 via a reverse idle gear 13 are supported.

In the above, when the main first speed gear 3 rotatably supported by the main shaft MS is connected to the main shaft MS by the first speed hydraulic clutch C1, the first speed is established. 1st speed hydraulic clutch C1 is 2nd to 4th
Since the engaged state is maintained even when the gear stage is established, the first-speed counter gear 8 is supported via the one-way clutch COW. When the main second speed gear 4 rotatably supported by the main shaft MS is coupled to the main shaft MS by the second speed hydraulic clutch C2, the second speed is established. When the counter third speed gear 10 rotatably supported by the counter shaft CS is coupled to the counter shaft CS by a third speed hydraulic clutch C3, a third speed is established.

In a state where the counter fourth speed gear 11 rotatably supported on the counter shaft CS is connected to the counter shaft CS by the selector gear SG, the main fourth speed gear 6 rotatably supported on the main shaft MS is connected to the fourth gear.
Speed-reverse hydraulic clutch C4R with main shaft M
When engaged with S, the fourth gear is established. With the counter reverse gear 12 relatively rotatably supported on the counter shaft CS connected to the counter shaft CS by the selector gear SG, the main reverse gear 7 rotatably supported on the main shaft MS is connected to the fourth speed-reverse gear. When the hydraulic clutch C4R is engaged with the main shaft MS, the reverse gear is established.

In the above, the clutches C1, C2, C3, C4
R corresponds to the friction engagement element. The friction engagement element in the illustrated example is a wet multi-plate clutch, and more specifically, has a structure in which friction materials (not shown) are adhered to both surfaces with n clutch disks (not shown). The friction material is made of, for example, a paper material, and the characteristic of the friction coefficient μ increases as the sliding speed (difference rotation Δω) increases at 0.08 in a stationary state as shown in FIG. 2 and reaches a maximum at 0.12. . In the embodiment, the clutches C1, C2, C3 and C4R are constituted by rotary hydraulic clutches, and the pistons are provided with check valves (not shown) for centrifugal hydraulic discharge.

The rotation of the counter shaft CS is
The power is transmitted to the differential D via the final drive gear 14 and the final driven gear 15, and then transmitted to the drive wheels W, W via the left and right drive shafts 16, 16.

Here, the intake passage of the internal combustion engine E (not shown)
A throttle opening sensor S1 for detecting the opening θTH is provided in the vicinity of a throttle valve (not shown) disposed at the position. In the vicinity of the final driven gear 15,
A vehicle speed sensor S2 for detecting the vehicle speed V from the rotation speed of the final driven gear 15 is provided. Further, near the crankshaft 1, a crank angle sensor S3 for detecting the engine speed Ne from the rotation thereof is provided.

An input shaft speed sensor S4 for detecting the input shaft speed NM of the transmission through its rotation is provided near the main shaft MS, and near the counter shaft CS through its rotation. An output shaft speed sensor S5 for detecting the output shaft speed NC is provided. Further, P, R, N, D4, D3, and P3 are located near a shift lever (not shown) mounted on the floor of the vehicle driver's seat.
A shift lever position sensor S6 for detecting a position selected by the driver among the seven positions 2, 1 is provided.

In the vicinity of the drive shaft 16, a torque meter S for detecting the driving force (driving torque) TDS is provided.
7 are provided. The output of these sensors S1 etc. is EC
U (electronic control unit).

The ECU is a CPU 17, a ROM 18, a RAM
19, a microcomputer comprising an input circuit 20 and an output circuit 21. Outputs of the above-mentioned sensor S1 and the like are input into the microcomputer via the input circuit 20. In the microcomputer, the CPU 17 determines a shift position (gear position), and the shift solenoid SL of the hydraulic control circuit O through the output circuit 21.
A shift valve (not shown) is switched by exciting / de-energizing 1, SL2, and a hydraulic clutch of a predetermined gear stage is released / engaged.

Reference numerals SL3 and SL4 denote an ON / OFF control solenoid and a capacity control solenoid of the lock-up mechanism L of the torque converter 2. Also, the symbol S
L5 is a linear solenoid for clutch hydraulic pressure control. Reference numeral S8 generically indicates three pressure heads that detect clutch oil pressures of the clutches C2 to C4R.

FIG. 3 is a flowchart showing the operation of the control device for the hydraulically operated transmission according to the embodiment.

FIG. 3 is a flow chart showing the above-described blowing margin ratio, (engagement side clutch torque transmission capacity TON (t) + disengagement side clutch torque transmission capacity TOFF (t)) / input torque TE (t) -1.
The procedure for calculating (× 100%) will be described. Here, t indicates time. This program is started, for example, every 20 ms. FIG. 4 is a timing chart for explaining the operation.

First, in step S10, the input torque TE is calculated.

The input torque TE is obtained by multiplying a value retrieved according to predetermined characteristics from the detected engine speed and the engine load such as the intake pressure or the throttle opening by the torque ratio of the torque converter 2 to obtain the transmission at the start of the torque phase. It is obtained by calculating the input torque. Note that the input torque TE after that point is determined by linear interpolation as shown in the figure. The input torque TE may be calculated by multiplying the torque TDS acting on the drive shaft 16 detected through the torque meter S7 by a gear ratio.

Then, the program proceeds to S12 in which the torque (driving force) transmission capacity TON of the engagement (on) side clutch is obtained as a value converted on the main shaft MS.

In general, the torque transmission capacity is calculated by detecting a hydraulic pressure acting on the clutch and the like by using a well-known formula described later. When the input shaft torque TT (indicated by the symbol a) is reversed, the torque transmission capacity T of the on-coming clutch is determined.
Focusing on a point similar to ON (indicated by the symbol b), that is, based on the recognition that the pull-in of the output torque of the transmission is caused by an increase in the engagement force of the friction engagement element, the torque phase capacity is determined. The estimation was made without using the oil pressure. Note that the capacity of the inertia phase is calculated using a known formula described later, and the discontinuity due to the difference in the calculation method between the two phases is matched. In the timing chart of FIG. 4, these three zones are indicated by a, b, and c.

Thus, in the torque phase, the torque transmission capacity of the on-coming clutch is determined by the input torque T
E and main shaft M of drive shaft torque TDS
The following calculation is made from the difference between the converted torques TT on S. TON (t) = {TE (t) −TT (t)} × ioff / (ioff−ion). . Here, t is the time as described above, ioff is the gear ratio on the release side, and ion is the gear ratio on the engagement side. That is, the input shaft torque TE before the shift is started and the input shaft torque TT during the shift are obtained, and the torque transmission capacity of the frictional engagement element on the engagement side is obtained in consideration of the gear ratio difference.

To further expand on this, the torque TOUT on the output shaft is obtained as follows. FIG. 5 illustrates the parameters used. TOUT = TOFF × ioff + TON × ion = (TIN−TON) × ioff + TON × ion = (TIN × ioff) −TON × (ioff−ion) where, TOFF: torque transmission capacity of the frictional engagement element on the release side, ioff: gear ratio of the current speed stage, ion: gear ratio of the destination speed stage, TIN: torque acting on the transmission input shaft (main shaft MS).

Further, the following relationship is established. TON = (TIN × ioff-TOUT) / (ioff-ion). . TOUT = TT (t) × ioff. . TIN = TE (t). . Therefore, the following is derived from the equation.

Incidentally, the input shaft torque TT (t) during the shift is obtained as follows. TT = 2.times.TDS (t) / (ifal.times.ioff.times..eta.) Where TDS (t): torque acting on the drive shaft 16 at the time t via the torque meter S7, ifinal: final gear Ratio, η: Transmission efficiency (for example, 0.9 in consideration of various losses such as stirring resistance). It should be noted that multiplying by 2 in the equation is for obtaining the torque value on the assumption that the torque output from the transmission is traveling straight ahead transmitted equally to the two drive shafts on the left and right.

Next, the calculation of the torque transmission capacity of the engagement-side clutch in the inertia phase will be described.

The torque transmission capacity in the inertia phase is obtained by the following equation. Although the hydraulic fluid (ATF) is not fully charged in the piston chamber of the clutch at the beginning of the inertia phase, it is assumed that the hydraulic fluid is fully charged for the sake of calculation. TON (t) = μ × 2n × Rm × {PON (t) × Apis + Fε−Frtn}. . . Here, μ: coefficient of friction, n: number of clutch disks (multiplied by 2 to use a structure in which friction material is stuck on both sides), Rm: effective radius of clutch disk, PON: hydraulic pressure of the engaging clutch, Apis : Piston receiving area, Frt
n: Indicates the return spring load.

Fε indicates the pressure due to the centrifugal force acting on the hydraulic oil in the piston chamber, and is obtained by the following equation. Fε = (πρ / 4g) × ω ² × Rout where ρ: ATF density, Rout: piston outer diameter.

The rotation speed ω is obtained as follows. That is, the value when the clutch is on the main shaft MS is obtained by ω = Ne (t) × ETR (t). The value when the clutch is on the countershaft CS is obtained by ω = Ne (TINT) × ETR (TINT) / ioff.

The value of ω on the counter shaft CS is
The engine speed Ne (TINT) at the start of the inertia phase is used instead of Ne (t) that changes with time because it is considered to be substantially constant because it is directly connected to the drive wheel W.
Here, TINT: the moment of inertia start, and ETR: the torque converter slip ratio.

Further, the differential rotation Δω is obtained by Δω = Ne (t) × ETR (t) × {1-ECL2 (t)}. Here, ELC2 indicates the slip ratio of the on-coming clutch which starts the calculation at the same time as the shift command.

Here, a supplementary explanation of the friction coefficient μ will be given.

The clutches C1, C2, C used in the embodiment
3, the characteristics of the friction coefficient μ of C4R are as shown in FIG. 2. However, as described above, the friction coefficient μ increases according to the differential rotation Δω, and becomes constant when the differential rotation Δω reaches a predetermined value. . This means that the friction coefficient μ changes from large to small immediately before the end of the gear shift (end of the inertia phase), and the torque transmission capacity decreases. Accordingly, in this embodiment, as shown in FIG. 4, the friction coefficient μ is forcibly set near the end of the inertia phase, and the maximum value of the clutch to be used (specifically, 0.12) is used.
To a minimum value (specifically, 0.08) in a stepwise manner (three steps in the embodiment).

The switching start point is determined by a parameter corresponding to the vicinity of the end of the shift, for example, the predicted engine speed Ne at the end of the shift obtained from the engine speed at the start of the shift and the current gear ratio, or the main shaft speed. NM or the comparison value of the input / output rotational speed of the clutch (slip ratio E
CL, differential rotation) or the like, but in this embodiment, the switching start point is changed based on the maximum value of the engine speed Ne or the main shaft speed NM at the time of shift transition.

The reason is that the characteristic of the friction coefficient μ with respect to the differential rotation Δω differs depending on the temperature, and becomes lower as the temperature becomes higher.

That is, if the vehicle speed is the same, in other words, if the countershaft rotation speed NC is the same, the higher the transmission input shaft rotation speed (main shaft rotation speed NM), the larger the amount of heat generated by friction until the shift is completed. , The temperature rise becomes large. That is, as the engine speed Ne (or the main shaft speed NM) is higher, the timing of the decrease of the friction coefficient μ is earlier as shown in FIG. 6, so that the friction coefficient μ used in the equation is changed from the maximum value to the minimum value. Changed the timing of changing hands.

Since the engine speed Ne (or the main shaft speed NM) increases in proportion to the throttle opening θTH, the switching start point is replaced with the engine speed Ne (or the main shaft speed NM) instead of the throttle speed. It may be possible to change the opening degree θTH.

Next, a description will be given of the torque phase-inertia phase matching. As described above, this is an operation for correcting values obtained by different methods for the torque phase and the inertia phase so as to be smoothly connected. is there. Specifically, FIG.
As shown by C, the values of the torque phase and the inertia phase are matched by linear interpolation.

More specifically, since the phase shifts from the torque phase to the inertia phase at the point where the torque is most likely to be pulled in, FIG.
As shown in the timing chart, the time point at which the drive shaft torque TDS becomes minimum is detected, and the time point at which both phases shift is determined. Then, an increase in the hydraulic pressure PON at time t1 after the transition is estimated. This is performed by detecting the rate of increase in oil pressure from the output of the pressure head S8 and estimating the amount of increase in oil pressure after time t1 (may be estimated from the command value to the linear solenoid SL5).

Then, the estimated hydraulic pressure PON is substituted into the above equation to estimate the torque transmission capacity of the inertia phase after time t1, and a straight line is connected to the torque transmission capacity at the transition. That is, the torque transmission capacity during the time t1 is obtained by linear interpolation during that time.

The correction time t1 is changed according to the clutch rotational speed (ω described above). When the clutch rotation speed is high, the check valve for centrifugal pressure discharge normally provided in the clutch is not closed until the pressure becomes high. Therefore, when the clutch rotation speed is high, the correction time t1 is lengthened. Conversely, if the clutch rotation speed is low, in other words, the check valve is closed at a low pressure, the correction time t1 is shortened. The correction time t1 may be changed, for example, by the clutch pressure in addition to the clutch rotation speed. The higher the clutch pressure, the sooner the check valve closes. Therefore, the higher the clutch pressure, the shorter the correction time t1. I just need.

[0056] In the above, since the sought capacity without using the formula in the torque phase, it can be determined with high accuracy capacitance without being affected by the centrifugal hydraulic pressure in the rotary hydraulic clutch, also using equation in the torque phase Needless to say, the capacity may be obtained by using the above method, and in this case, the matching operation is not required.

In the flow chart of FIG.
Proceeding to 14, the friction coefficient coefficient μ of the release side clutch is rewritten to a predetermined value μFIX, and the flow proceeds to S16 to calculate the torque transmission capacity of the release side clutch. The capacity on the release side is also obtained as a value converted on the main shaft MS, similarly to the capacity on the engagement side. FIG. 7 shows the release-side capacitance TOFF.

The torque transmission capacity TOFF of the release clutch
Is calculated by the same formula as the capacity of the inertia phase on the engagement side described above. More specifically, TOFF (t) = μ × 2n × Rm × ｛POFF (t) × Apis + Fε
−Frtn Ｐ Here, POFF indicates the hydraulic pressure of the release clutch, and the clutch is switched at the clutch speed. As described above, the friction coefficient μ is a predetermined value μFIX (specifically, a value increased by 30% of the maximum value). The remainder is not different from the previous one.

Then, the program proceeds to S18, in which a blowing margin is calculated from the above-mentioned equation. If the calculated value is, for example, about 10%, it can be determined that there is almost no toughness for the blow-up, and it is necessary to delay the release timing as described later.

In this embodiment, as described above, the friction coefficient μ is set to a predetermined value in the equation, so that the above-mentioned blowing margin ratio is accurately obtained, so that the co-meshing can be optimally controlled, and the shift shock is reduced. It becomes possible to reduce.

That is, the clutches C1 and C1 used in the embodiment
When the friction coefficient μ of the friction material of the clutch disk surface of C2, C3, C4R has the characteristic as shown in FIG. 2, as shown by the imaginary line in FIG.
FIX, more specifically, increased by 30% to 0.16.

That is, as described earlier, since the friction coefficient μ increases with an increase in the differential rotation Δω, even if the engagement hydraulic pressure POFF of the release side clutch decreases and the differential rotation Δω increases, the release side clutch increases. Does not substantially change. However, since the differential rotation Δω increases, in other words, the input shaft rotation speed increases, when the rotation speed exceeds the blowing determination level, there is a disadvantage that the driver feels as a blow-up.

Therefore, by setting the friction coefficient μ to a constant value, the torque transmission capacity is reduced with a decrease in the engagement oil pressure POFF of the disengagement side clutch, and the blow margin is also reduced. In other words, the differential rotation Δ due to the decrease in the engagement oil pressure POFF
The blowing margin decreases as ω increases. Therefore, it is possible to detect from the change in the blow-off margin rate that the differential rotation Δω increases, in other words, the engine rotation speed or the transmission input shaft rotation speed increases and exceeds the blow determination level, so that the driver experiences a blow-up. Can be prevented.

Further, when obtaining the blow margin, it is necessary to well obtain the transition of the blow margin in the torque phase, including the period until the engine speed or the transmission input shaft speed increases and exceeds the blow determination level. For example, "Even if you release it earlier, no blow-up will occur." On the contrary, "Because there is little toughness against the blow-up, there is a risk of a blow-up if you do not release slowly." Should be released slowly ". Thereby, for example, in the hydraulic control circuit, it becomes possible to correct the set values of the oil discharge control valve for the clutch, the throttle of the oil passage, and the like.

In the embodiment, the coefficient of friction μ is not the maximum value of 0.12 of the frictional engagement element but a value increased by about 30%. This makes it possible to absorb the variation in the friction coefficient μ due to the manufacturing variation of the frictional engagement element, and when the frictional engagement element is fixed to the maximum value of 0.12, the point in time when the blowing margin becomes 0% is also detected.
5, there is no allowance for setting the engagement hydraulic pressure and setting the hydraulic control valve, and there is no safety margin.

On the other hand, by fixing to about 0.16, which is an increase of about 30%, the point at which the blowing margin becomes 0% is actually a state where there is a margin of 30%. As a result, the blowing margin becomes 30%. % Can be detected.
The setting on hardware such as an electromagnetic valve or the setting on software such as its duty ratio may have a margin with respect to the blowing margin. In other words, what percentage of the maximum value of the friction coefficient μ should be fixed is
What is necessary is just to determine from the margin of setting.

[0067]

According to the control apparatus for a hydraulically operated transmission according to the first aspect, it is possible to eliminate the inconvenience that the engine rotation speed is increased due to the input rotation exceeding a predetermined level. In addition to this, it is possible to determine the degree of increase in the input rotation, and it is possible to optimally perform the pressure reduction control of the hydraulic pressure of the disengagement side frictional engagement element or the control of the release timing. By performing control according to the target value, the shift shock can be reliably reduced.

According to the second aspect of the invention, it is possible to more effectively eliminate the inconvenience of causing the engine speed to rise due to the input speed rising beyond a predetermined level. It is possible to optimally perform the pressure reduction control of the hydraulic pressure of the disengagement side frictional engagement element or the control of the disengagement timing, so that the co-meshing can be controlled to the target value and the shift shock can be reduced reliably.

[Brief description of the drawings]

FIG. 1 is an explanatory diagram generally showing a control device for a hydraulically operated transmission according to the present invention.

FIG. 2 is an explanatory diagram showing characteristics of a friction coefficient μ of the friction engagement element of FIG.

FIG. 3 is a flowchart showing the operation of the apparatus in FIG. 1;

FIG. 4 is a timing chart for explaining a calculation operation of the flow chart of FIG. 3;

FIG. 5 is an explanatory diagram showing parameters used for calculating an engagement-side torque transmission capacity in the calculation of the flowchart of FIG. 3;

FIG. 6 is an explanatory diagram showing a characteristic of a friction coefficient μ with respect to a rotation speed in an operation of calculating an engagement side torque transmission capacity in the calculation of the flow chart of FIG. 3;

FIG. 7 is a timing chart showing the torque transmission capacity on the release side in the calculation of the flow chart of FIG. 3;

[Explanation of symbols]

E Internal combustion engine T Transmission O Hydraulic control circuit C1, C2, C3, C4R Clutch (friction engagement element)

Continuation of front page (58) Field surveyed (Int.Cl. ⁷ , DB name) F16H 61/00-61/24

Claims (2)

(57) [Claims] [Claim 1 further comprising a plurality of frictional engagement elements engaged is supplied with hydraulic oil, to free up one of the frictional engagement elements, a gear shift by switching an engagement state by engagement of the other A control device for a hydraulically operated transmission that performs: a. The friction coefficient at the time of release of the one friction engagement element is
It rewrites the maximum value over a predetermined value that can be taken of the serial friction coefficient
Releasing side friction coefficient rewriting means; b . The driving force transmission capacity at the release of the frictional engagement element before Symbol hand, a release-side driving force transmission capacity detecting means for determining by considering the predetermined value, c. Driving force transmission capacity when the other friction engagement element is engaged
Engaging side driving force transduction capacity detecting means for obtaining, d. Detecting a driving force input to the hydraulically operated transmission;
Input driving force detecting means ; e . Driving force transmission capacity when disengaged and driving force transmission when engaged
The sum of the capacity, and comparing means for comparing the input driving force, and f. Hydraulic, characterized in that the basis of the comparison result of the comparison means, with a, a rotational fluctuation determination means for determining the degree of quality of the increase of the input rotation in the gear shifting of said hydraulically operated <br/> transmission Control device for transmission. 2. The engagement-side driving force transmission capacity detecting means , comprising: g . The friction coefficient at the time of engagement of the other friction engagement element is
According to the differential rotation between the input and output sides of the other friction engagement element
Wherein Mochikaeru engagement side frictional coefficient dimensional worlds means, Te, before
The control device for a hydraulically operated transmission according to claim 1, wherein the driving force transmission capacity is obtained based on the newly held friction coefficient .