JPH0621642B2 - Shift hydraulic control device for automatic transmission for vehicle - Google Patents

Description

Description: BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a shift hydraulic pressure control device for an automatic transmission for a vehicle, which supplies proper hydraulic pressure without shift shock when a friction engagement element is engaged.

[Conventional technology]

BACKGROUND ART An automatic transmission for a vehicle must supply hydraulic pressure to hydraulic friction engagement elements such as clutches and brakes, and engage these friction engagement elements with rotary elements such as a rotary drum and gears to select them. Thus, the gear ratio switching (gear shifting) between the input shaft and the output shaft is automatically performed according to the operating state of the vehicle.

An example of a general vehicle automatic transmission will be described with reference to FIG. 7 showing a schematic structure thereof. A crankshaft 4 of an engine 2 which is a power source of the vehicle is a pump 8 of a torque converter 6.
Is directly connected to.

The torque converter 6 includes a pump 8, a turbine 10, a stator 12, and a one-way clutch 14, and the stator 1
2 is coupled to the case 16 via a one-way clutch 14, and the one-way clutch 14 causes the stator 12 to rotate in the same direction as the crankshaft 4, but does not allow rotation in the opposite direction.

The torque transmitted to the turbine 10 is transmitted by an input shaft 20 integrated with the turbine 10 to a gear transmission 22 arranged at the rear of the turbine 10 for achieving four forward gears and one reverse gear.

The transmission 22 includes three sets of clutches 24, 26, 2
Eight, two sets of brakes 30, 32, one set of one-way clutch 34 and one set of Ravigneaux type planetary gear mechanism 36 are configured. The planetary gear mechanism 36 includes a ring gear 38,
Long pinion gear 40, short pinion gear 42,
The front sun gear 44, the rear sun gear 46, and both pinion gears 40 and 42 are rotatably supported and are composed of a carrier 48 which is also rotatable. The ring gear 38 is connected to an output shaft 50, and the front sun gear 44 is a kickdown drum 52. Input shaft 2 via front clutch 24
0, the rear sun gear 46 is connected to the input shaft 20 via the rear clutch 26, and the carrier 48 is connected to the case 16 via the low reverse brake 32 and the one-way clutch 34 which are functionally arranged in parallel. The transmission 22 is connected to the input shaft 20 via a fourth speed clutch (end clutch) 28 arranged at the rear end of the transmission 22. The kick down drum 52 can be fixedly connected to the case 16 by the kick down brake 30.

Then, the torque passing through the planetary gear mechanism 36 is applied to the output shaft 5
The transfer shaft 6 fixed to 0 is transmitted to the driven gear 64 through the idle gear 62 and further fixed to the driven gear 64.
6, via the helical gear 68, it is transmitted to the differential gear unit 72 to which the drive shaft 70 of the drive wheel is connected.

Each of the above-mentioned clutches and brakes, which are frictional engagement elements, is constituted by a frictional engagement device having an engagement piston device or a servo device, and is driven by the engine 2 by being connected to the pump 8 of the torque converter 6. It is operated by the hydraulic pressure generated by an oil pump (not shown). The hydraulic pressure from this oil pump is selectively supplied to each clutch and brake according to the operating states detected by various operating state detecting devices by a hydraulic control device described later, and depending on the combination of the operation of each clutch and brake. A shift speed of four forward gears and one reverse gear is achieved, and the detailed structure and operation thereof are as disclosed in JP-A-58-46258.

Here, for example, in the 1st speed shift stage other than the 1st speed fixed range (so-called D range), the front clutch 2
The fourth and fourth speed clutches 28 are not connected, the kick down brake 30 and the low reverse brake 32 are released, the rear clutch 26 is connected, and the one-way clutch 34 is in a functioning state. . In this case, the upshift to the second speed is achieved only by connecting the kick down brake, and the one-way clutch 34 does not function.

As shown in FIG. 8 showing a part of a hydraulic control device for realizing this speed change system, a 1-2 shift valve 81 is provided in a kick down servo 80 for controlling the operation of the kick down brake 30 via an oil passage 82. A hydraulic control valve 83 and a shift control valve 84 are connected to the 1-2 shift valve 81 via oil passages 85 and 86, respectively.

Since the pair of solenoid valves 87 and 88 that control the operation of the shift control valve 84 in the first speed shift stage release the oil passages 89 and 90, respectively, the spool 9 in the center of the shift control valve 84 is closed.
8 shifts to the left side in FIG. 8 to connect the oil passage 86 to the oil discharge port EX of the shift control valve 84, and the 1-2 shift valve 81
The spool 92 is in the state of being displaced to the left end in FIG. As a result, the oil passage 82 communicates with the oil discharge port EX of the 1-2 shift valve 81, and the piston 94 is returned to the right side in FIG. 8 by the spring force of the compression coil spring 93 of the kickdown servo 80. The engagement of the kick down brake 30 with the kick down drum 52 is released. Further, of the two oil passages 95, 96 connected to the hydraulic control valve 83, the duty ratio of the solenoid valve 98 attached to one oil passage 95 and duty-controlled by the electronic control unit 97 is 1.
The oil pressure is set to 00%, and the oil pressure control valve 83 is provided in the oil passage 95.
Hydraulic pressure to overcome the return spring is not working.
Therefore, the spool 99 of the hydraulic control valve 83 is
It is displaced to the left and the oil passage 85 communicates with the oil discharge port EX of the hydraulic control valve 83.

When an upshift from this state to the second speed is performed, the electronic control unit 97 operates one solenoid valve 87 to close the oil passage 89 depending on the running condition of the vehicle, so that the center spool 91 is closed.
In FIG. 8, the pressure oil from the oil pump (hereinafter, referred to as line pressure) is displaced to the right side in FIG. 8 to pass from the oil passage 100 connected to the shift control valve 84 through the oil passage 86 to 1-2.
It is fed to the shift valve 81. Therefore, the spool 92 of the 1-2 shift valve 81 is displaced to the right end in FIG.
2, 85 communicate with each other through the 1-2 shift valve 81. On the other hand, since the duty ratio of the solenoid valve 98 is reduced by the electronic control unit 97, the oil pressure in the oil passage 95 increases in accordance with the decrease in the duty ratio, and this increased oil pressure is applied to the spool 9
Act on 9. As a result, the spool 99 is
The line pressure from the oil passage 96 is displaced to the right, and the line pressure from the oil passage 96 is reduced by the hydraulic control valve 83 according to the duty ratio and supplied to the oil passage 85. After that, the hydraulic pressure adjusted to a value corresponding to the duty ratio is supplied to the kickdown servo 80 through the oil passages 85 and 82, and the piston 94 is stroked to the left side in FIG. 8 to operate the kickdown brake 30. The kick down drum 52 is tightened.

By the way, at all shift speeds including the shift speed from the 1st speed to the 2nd speed, if the characteristic of change in the hydraulic pressure to be fed to the friction engagement element is inappropriate, the friction engagement element suddenly engages. As a result, gear shift shock may occur, or excessive friction may occur in the friction engagement element to deteriorate the friction engagement element. Therefore, in the shift stage from the first speed to the second speed,
Conventionally, as shown in FIG. 8, a piston stroke sensor 101 such as a potentiometer attached to the piston 94 detects the duty ratio of the solenoid valve 98 that controls the change characteristic of the hydraulic pressure supplied to the kickdown servo 80. The control was performed in the following manner in correspondence with the operation stroke position of 94. That is, depending on the working stroke position of the piston 94 shown in FIG. 9 (c) (operation stroke range of the piston 94 is between the partial pressure ratio of the piston stroke sensor 101 of k ₃ to k _2), 9 ( As shown in a), the solenoid valve 9 is operated by an electric command from the electronic control unit 97.
The duty ratio of No. 8 is changed, and FIG. 9 (b)
As indicated by a dotted line, a high hydraulic pressure is supplied to the kick-down servo 80 by an electric command until the piston 94 strokes toward the engagement side and reaches a preset setting position just before the engagement. When the piston 94 reaches the set position, the hydraulic pressure by the electric command (electric command hydraulic pressure)
Is lowered to a value suitable for gear shifting without gear shifting shock and gear shifting is executed. In FIG. 9 (b), the stroke of the piston 94 changes the hydraulic chamber volume of the kickdown servo 80. Therefore, as shown by the dotted line in the figure, the electric command hydraulic pressure (feed hydraulic pressure) is actually stepped. On the piston 94
The hydraulic pressure that acts on changes as shown by the solid line in the figure.

[Problems to be solved by the invention]

However, in such a conventional shift hydraulic pressure control means, a rapid rise of hydraulic pressure occurs as indicated by a in FIG. 9 (b) due to the hydraulic pressure fluctuation (overshoot) at the completion of the operation stroke of the piston 94. Therefore, there is a problem in that the kick-down brake 30 is suddenly engaged and the occupant is pushed forward to cause a shift shock.

Therefore, as shown in FIGS. 10 (a) and 10 (b), the hydraulic pressure is temporarily reduced immediately before the piston stroke of the friction engagement element is completed (d ₁ + 5% in terms of duty ratio. Here, d ₁ is learned). It is possible to suppress the torque wave when the piston stroke of the friction engagement element is completed by lowering the piston stroke speed and gradually engaging the friction engagement element thereafter. One is also proposed (Japanese Patent Application No. 60-13453).

It should be noted that FIG. 10 (c) is a graph showing changes in the kick down drum pressing force.

However, with such conventional means, d ₁ + 5%
In the low throttle opening range where the oil pressure is low, the piston of the friction engagement element does not make a stroke and b in FIG. 10 (d).
While a feeling of braking is generated in the portion (solid line portion), in the high throttle opening region where the hydraulic pressure corresponding to d ₁ + 5% is high, d ₁ +
At the time of transition from 5% to d ₁ % [reference c in FIGS. 10 (a) and 10 (b)]
[Refer to the section], the hydraulic pressure equivalent to 5% is increased, so that the piston once decelerated is accelerated again, which causes a torque wave in the section e (dashed line) of FIG. 10 (d). As a result, in both cases, the shift feeling becomes poor.

Further, since the hydraulic pressure increase rate β from the command of the initial hydraulic pressure to the start of the actual shift is set to a constant value (for example, 3% / sec) regardless of the throttle opening, this value of β is more than necessary. If it is set to a large value, the torque rises rapidly in the low throttle opening range, causing a shock, while if the value of β is smaller than the required value, the feeling of braking in the high throttle opening range is large. As a result, the shift feeling is deteriorated.

The present invention is intended to solve such a problem, and a shift hydraulic pressure control device for an automatic transmission for a vehicle, which does not generate a brake feeling or a shock at the time of shifting in any throttle opening range. The purpose is to provide.

[Means for solving problems]

Therefore, the shift hydraulic control device for an automatic transmission for a vehicle of the present invention selects an input shaft to which the rotational power of the engine is input, an output shaft for outputting the rotational power to the drive wheels of the vehicle, and an arbitrary rotary element. In an automatic transmission for a vehicle having a hydraulic friction engagement element that selectively switches the transmission ratio between the input shaft and the output shaft by performing the above, an electric signal is supplied to the friction engagement element. Hydraulic pressure adjusting means for adjusting the rotational speed of the rotating element that changes the rotational speed during shifting, and a target change rate for which the change rate of the rotational speed detected by the detecting means is set in advance. Control means for electrically feedback controlling the hydraulic pressure to the friction engagement element so as to follow the above, and an initial hydraulic pressure setting means for setting the value of the initial hydraulic pressure supplied to the friction engagement element at the start of gear shifting. A third electric control for initial hydraulic pressure setting, wherein the initial hydraulic pressure setting means obtains a first electric control value for determining the initial hydraulic pressure from a second electric control value at a required time point of the feedback control. And a fourth electric control value corresponding to the minimum hydraulic pressure for driving the movable member of the friction engagement element.

Further, the shift hydraulic pressure control device in the vehicular automatic transmission of the present invention selects an input shaft to which the rotational power of the engine is input, an output shaft to output the rotational power to the drive wheels of the vehicle, and an arbitrary rotary element. Thus, in a vehicle automatic transmission having a hydraulic friction engagement element that selectively switches the gear ratio between the input shaft and the output shaft, the hydraulic pressure supplied to the friction engagement element is converted into an electric signal. A hydraulic pressure adjusting means for adjusting the rotation speed, a detecting means for detecting the rotation speed of the rotating element whose rotation speed changes during the shift, and a change rate of the rotation speed detected by the detecting means to a preset target change rate. A control means for electrically feedback-controlling the hydraulic pressure to the friction engagement element so as to follow is provided, and an initial hydraulic pressure setting hand for setting the value of the initial hydraulic pressure supplied to the friction engagement element at the start of gear shifting. And the hydraulic pressure changing means for changing the hydraulic pressure at a required ratio after supplying the initial hydraulic pressure set by the initial hydraulic pressure setting means to the friction engagement element. The first electric control value for determining the third electric control value for setting the initial hydraulic pressure obtained from the second electric control value at the required time point of the feedback control, and the minimum for driving the movable member of the friction engagement element. The hydraulic pressure changing means is configured to have a means for setting it to a fourth electric control value corresponding to the hydraulic pressure, and the hydraulic pressure changing means has a hydraulic pressure change rate varying means for changing the hydraulic pressure change rate according to the operating state of the vehicle. It is characterized by being configured.

[Work]

In the shift hydraulic control device for the automatic transmission for a vehicle according to the first aspect of the invention, when an electric signal for shifting is output during shifting, a relatively high hydraulic pressure is first supplied to the friction engagement element, which causes friction. The movable member of the engagement element is driven toward the engagement side, the hydraulic pressure is lowered immediately before the engagement, and the initial hydraulic pressure is set to the initial hydraulic pressure state set by the initial hydraulic pressure setting means. As a result, the movable member of the friction engagement element is decelerated and further driven to the engagement side. At this time, the first for determining this initial hydraulic pressure
The electric control value corresponds to the third electric control value for setting the initial hydraulic pressure obtained from the second electric control value at the required time point of the feedback control, and the fourth electric power corresponding to the minimum hydraulic pressure for driving the movable member of the friction engagement element. Since it is set between the control value and the control value, a feeling of braking or a shock is not caused at the time of shifting due to the magnitude of the throttle opening.

Further, in the shift hydraulic control device for the automatic transmission for a vehicle of the second invention, the movable member of the friction engagement element is driven in the initial hydraulic state, and then the hydraulic pressure is changed to the high pressure side at a required ratio. Although the degree of engagement is increased, the rate of change in hydraulic pressure at this time is changed in accordance with the operating state of the vehicle, so that it is still possible to sufficiently suppress the feeling of braking and shock during shifting.

〔Example〕

Hereinafter, with reference to the drawings, a shift hydraulic control device for an automatic transmission for a vehicle according to an embodiment of the present invention will be described.
FIG. 1 is its schematic configuration diagram, FIG. 2 is a flow chart for explaining its action, and FIGS. 3 (a), (b), (c), (d),
(e) and (f) show the duty ratio change, electric command oil pressure change, kickdown servo oil pressure change, piston stroke change, drum pressing force change, and transmission output shaft torque change, respectively, in the low throttle opening range. The graphs shown in comparison with those in Fig. 4, (a), (b), (c), (d), (e), and (f) are duty ratio changes and electrical commands in the high throttle opening region, respectively. A graph showing changes in hydraulic pressure, changes in kickdown servo hydraulic pressure, changes in piston stroke, changes in drum pressing force, changes in transmission output shaft torque, and FIG. 5 is a graph for explaining the throttle opening zone. A characteristic diagram, FIG. 6 is a graph showing a target change of the turbine rotation speed, and in FIGS. 1 to 6, the same reference numerals as those in FIGS. 7 to 10 indicate almost the same parts.

As shown in FIGS. 1 and 7, the automatic transmission AT equipped with this device has an input shaft 2 to which the rotational power of the engine 2 is input.
0 and an output shaft 50 that outputs rotational power to the driving wheels (front wheels or rear wheels) of the automobile, and three sets of clutches 24, 26, 28, two sets of brakes 30, 32,
The transmission 22 is provided with a set of one-way clutches 34 and a set of Ravigneaux type planetary gear mechanism 36.

Further, the automatic transmission AT includes the clutches 24, 26, 2
8 and the brakes 30, 32 and the like, a hydraulic pressure control circuit O constituting hydraulic pressure adjusting means for adjusting the hydraulic pressure to be supplied in accordance with an electric signal.
This hydraulic control circuit OC is shown in FIG. 8 as a part thereof, and this hydraulic control circuit OC includes a shift valve 81, a hydraulic control valve 83, a shift control valve 84, a manual valve, The hydraulic control solenoid valve 98 and the like are connected by appropriate oil passages, and these oil passages are further connected to the clutches 24, 2
6, 28, brakes 30, 32, etc.

By the way, a controller CL for outputting an electric control signal of a required duty ratio to the solenoid valve 98 of the hydraulic control circuit OC is provided. When the solenoid valve 98 receives an electric signal from this controller CL, for example, the duty ratio 100%.
As the signal oil pressure of the oil passage 95 becomes 0 and the duty ratio becomes smaller, the signal oil pressure of the oil passage 95 becomes higher accordingly, whereby the oil pressure acting on the spool 99 of the oil pressure control valve 83 changes, As a result, the line pressure from the oil passage 96 is reduced by the hydraulic control valve 83 according to the duty ratio and supplied to the oil passage 85.

Focusing on the hydraulic control supplied to the kick-down servo 80, the piston stroke sensor 101 for detecting the stroke amount of the piston 94, the throttle sensor 200 for detecting the throttle opening θt, the turbine rotation speed N _T ,
A rotation speed sensor 201 that detects the transmission output shaft rotation speed N ₀ (this rotation speed sensor 201 constitutes detection means that detects the rotation speed of a rotating element whose rotation speed changes during a gear change) and other engine rotations. Sensors 20 for detecting the number, engine output torque, engine oil temperature, vehicle speed, etc.
2 are provided, and the detection signals from these sensors 200 to 202 are input to the controller CL.

Further, the controller CL is to the control means C1,2 speed for electrically feedback controlled hydraulic pressure to the kickdown servo 80 such that the change rate _T of the turbine rotational speed N _T follows the preset target change rate _S Initial hydraulic pressure setting means IOSM for setting the value of the initial hydraulic pressure to be supplied to the kick down servo 80 at the start of the gear shift, and a required ratio β% after the initial hydraulic pressure set by the initial hydraulic pressure setting means IOSM is supplied to the kick down servo 80. Although it has each function of the oil pressure changing means OCM that changes the oil pressure in / sec, the initial oil pressure setting means IOSM further sets the electric quantity (first electric control value) d ₄ for determining the initial oil pressure. quantity of electricity at the required time of the feedback control quantity of electricity (third electric control value) for the initial oil pressure setting calculated from (the second electrical control values) d 1 _{'d 1} and The amount of electricity corresponding to the minimum hydraulic pressure that drives the piston 94 of the kick-down servo 80 (fourth
Electric control value) m, and the hydraulic pressure change means OCM changes the hydraulic pressure change rate β according to the operating state of the vehicle.
It is configured with a VM. Further, the oil pressure change rate varying means OCVM changes the oil pressure temporally with an inclination that increases as the oil pressure corresponding to the electricity amount d ₄ increases, and increases as the torque increases. It is configured to have means for changing the hydraulic pressure with time by the following inclination.

Specifically, the electric quantity d ₄ for determining the initial hydraulic pressure, that is, the shift initial duty ratio d ₄ is expressed by the following equation.

d ₄ = (d ₁ + 3 × m) / 4 (%) where d ₁ is the initial value calculated from the electric amount at the required time point of the feedback control (that is, the duty ratio corresponding to the previous shift result) d ₁ ′. The amount of electricity for setting the hydraulic pressure (that is, the duty ratio corresponding to the shift start hydraulic pressure obtained from d ₁ ′), and m is the amount of electric power (duty ratio) corresponding to the minimum hydraulic pressure (set value) for making the piston 94 stroke.

From the above equations, the d ₄ is expressed by the weighted expression between d ₁ and m, and d ₄ from this equation d ₁ and m 3: It can be seen that is set to a value obtained by internally dividing the 1.

Further, the hydraulic pressure change rate β is represented by the following equation, for example.

β = 6.8 + {1.1 × (m−d ₁ ) / 2} From the above equation, it can be seen that β is expressed as a linear equation of the difference between d ₁ and m.

Next, the operation of this device will be described. In this device, the hydraulic pressure supplied to the kickdown servo 80 during the shift is controlled according to the flowchart shown in FIG. In addition, this flowchart is based on Japanese Patent Application No. 59-69926 and Japanese Patent Application No. 59-8.
No. 2864, Japanese Patent Application No. 59-204491, etc., the feedback control for controlling the hydraulic pressure to be fed to the kick down servo 80 in the half-engaged state of the kick down brake 30 already proposed, and the automatic transmission are provided for the engine displacement, etc. The shift initial hydraulic pressure setting control for conforming to the standard is also executed, and the part according to the present invention is mainly the part between the * marks in the flowchart.

In the above flow chart, when the controller CL issues a shift start signal from the first speed to the second speed and the shift control solenoid valves 87 and 88 are switched as described above, the duty ratio of the solenoid valve 98 is instructed to be 40%. [See FIG. 3 (a) and FIG. 4 (a)]. As a result, this duty ratio command (electrical command) causes
And, as shown by the solid line in FIG.
%, A relatively high hydraulic pressure (electrically commanded hydraulic pressure) is sent to the kickdown servo 80, and the piston 94 strokes to the engagement side (left side in FIG. 8).

Next, the throttle opening θt of the engine 2 is detected and stored, and this throttle opening θt is divided into four zones A to D in which all throttle openings are divided as shown in FIG. The throttle opening area, the D zone are low throttle opening areas, and the B and C zones are intermediate throttle opening areas).

Now, if it is determined that the throttle opening θt belongs to the D zone (low throttle opening range), then this D
To read the constants θ ₁ ′, d ₁ ′, α in the zone and calculate the duty ratio d ₁ from the arithmetic expression d ₁ = d ₁ ′ −α (θt−θ ₁ ′), and to obtain the effect according to the present invention. Of the gear shift initial duty ratio d ₄ of (d ₁ + 3 × m) / 4%, and the subtraction value (oil pressure change rate) β of the gear shift initial duty ratio d ₄ is {6.8 + 1.1 × (m−d ₁ ) / 2} Δt. Note that Δt is the timer set time.

Next, the stroke position k of the piston 94 is detected by the piston stroke sensor 101, and while this stroke position is k ₂ − (90/8) × 0.8% or more, the duty ratio d
₄ is commanded to the solenoid valve 98 [see FIG. 3 (d)]. As a result, as shown by the solid line in FIG. 3 (b), the electric command hydraulic pressure becomes the electric command hydraulic pressure corresponding to the shift initial duty ratio d ₄ .

At this time, since the throttle opening θt belongs to the D zone (low throttle opening range), the gear shift initial duty ratio d ₄ is shown in FIG. 3 (a) from the calculation result of the calculation formula (d ₁ + 3 × m) / 4. ) Is set to the value shown by the solid line. Since this shift initial duty ratio d ₄ is smaller than the duty ratio d ₁ + 5% set by the conventional means (hereinafter, the conventional means or the conventional one means the technology of Japanese Patent Application No. 60-13453), The hydraulic pressure acting on the piston 94 can be made higher than that of the conventional one, and thus the time from the gear shift command to the start of the gear shift can be made shorter than that of the conventional one. You can eliminate the feeling [see Fig. 3 (f)].

In the present embodiment, the change from the high-pressure electric command hydraulic pressure corresponding to the duty ratio of 40% to the low-voltage electric command hydraulic pressure corresponding to the shift initial duty ratio d ₄ is k ₂ − (90 /8)×0.8%, but the stroke width is not limited to this, and is appropriately determined depending on various conditions of the engine and the automatic transmission. The stroke may not be defined, but the minute time may be defined.

By the way, after instructing the shift initial duty ratio d ₄ , the timer is set and the shift is further executed. After the timer is set, the rotation speed N _T of the turbine 10 and the rotation speed of the output shaft corresponding to the input shaft rotation speed. The rotational speed N ₀ of the transfer drive gear 60 corresponding to the above is detected, and in the present embodiment, it is determined whether or not the out-of-synchronization of the first speed defined by N _T ≦ N ₀ × 2.7 is achieved. If the result of this determination is that out-of-synchronization has not been achieved, the operation of subtracting the previously calculated subtraction value β from the initial shift duty ratio d ₄ so as to gradually increase the electrical command hydraulic pressure is repeated, Achieve out of sync.

At this time, since the throttle opening θt belongs to the D zone (low throttle opening region), the subtraction value β is obtained from the calculation result of the calculation formula {6.8 + 1.1 × (m−d ₁ ) / 2} Δt. Is set to a value that results in the solid line inclined portion in FIG. Since this value is smaller than the subtraction value β (= 3) of the related art, it is possible to moderate the rise of torque at the time of gear shifting, which improves the feeling of gear shifting.

When the first speed out of synchronization is achieved as described above, as shown in FIG. 6, the target change rate _S of the rotational speed of the turbine 10 corresponding to the throttle opening θt during this shift is calculated, The feedback control described above is performed. That is, the hydraulic pressure is supplied to the kickdown servo 80 at an optimum degree and the actual change rate _T of the rotational speed calculated from the actual rotational speed N _T of the turbine 10 to cause an engagement shock of the kickdown brake 30 or an excessive amount. When a shift is achieved without slippage or the like, the rotational speed change rate _S (target change rate) indicated by the turbine 10 is compared to calculate a deviation Δ between the actual change rate _T and the target change rate _S. To do. Then, a gain γ = bΔ, where b is a constant, for the deviation Δ is calculated to calculate a duty ratio correction amount Δd = γΔ, and the duty ratio d _{1 obtained} by correcting the duty ratio d ₁ with the duty ratio correction amount Δd is calculated. Command ₀ and solenoid valve 9
Drive eight. And the actual rotation speed N of the turbine 10
_T and the rotational speed N ₀ of the transfer drive gear 60
In the present embodiment, it is determined whether or not the synchronization to the second speed specified by N _T ≦ N ₀ × 1.7 has been achieved. As a result, when the synchronization is not achieved, the duty ratio d ₀ is used as the duty ratio d ₁ to perform the same correction as described above, and the feedback control is repeatedly performed to achieve the synchronization. That is, according to the above-mentioned feedback control, the hydraulic pressure sent to the kick-down servo has the deviation Δ
Is adjusted so that the actual rotational speed change of the turbine 10 shown by the dotted line in FIG. 6 follows the target change rate _S shown by the solid line in FIG. Therefore, the hydraulic pressure supplied to the kick-down servo 80 is optimally controlled, and the gear shift shock-free shift is achieved here as well.

When the synchronization is achieved as described above, the duty ratio d ₀ at this point is held for 0.1 second in this embodiment, and then the duty ratio is commanded to 100% to complete the shift to the second speed. During this operation, a constant α for shifting in the next D zone is calculated from the duty ratio d ₀ and the throttle opening θ ₀ by an arithmetic expression α = (d ₁ ′ −d ₀ ) / (θ ₀ −θ ₁ ′). This is stored in memory. As a result, in the next shift in the D zone, the constant α for calculating the duty ratio d ₁ is corrected by the feedback control, and the more appropriate duty ratio d _{1 and} thus the shift initial duty ratio d
₄ can be set.

By the way, considering the case where it is determined that the throttle opening θt belongs to the A zone (high throttle opening region), the constants θ ₁ ′, d ₁ ′, α in this A zone are also considered in this case.
Is read, and the duty ratio d ₁ is calculated by an arithmetic expression d ₁ = d ₁ ′ −
The shift initial duty ratio d ₄ for calculating the effect of the present invention is calculated from α (θt−θ ₁ ′) (d ₁
+ 3 × m) / 4%, and the subtraction value β of the shift initial duty ratio d ₄ is further calculated as {6.8 + 1.1 × (m−d ₁ ) / 2} Δt.

Then, also in this case, the stroke position k of the piston 94 is detected by the sensor 101, and this stroke position is k ₂
The duty ratio d ₄ is instructed to the solenoid valve 98 while − (90/8) × 0.8% or more [see FIG. 4 (d)]. As a result, the fourth
As shown by the solid line in FIG. 6B, the electric command hydraulic pressure becomes the electric command hydraulic pressure corresponding to the shift initial duty ratio d ₄ .

At this time, since the throttle opening θt belongs to the zone A (high throttle opening region), the shift initial duty ratio d ₄ is shown in FIG. 4 (a) from the calculation result of the calculation formula (d ₁ + 3 × m) / 4. ) Is set to the value shown by the solid line. The shift initial duty ratio d ₄ is set to the duty ratio d ₁ + 5% to d ₁ set by the conventional means.
Since it is smaller than the amount corresponding to 5% at the time of shifting to%, the change in hydraulic pressure acting on the piston 94 can be made smaller than that of the conventional one. 4 (f), see the dotted line section] can be eliminated. [Refer to the solid line in FIG. 4 (f)].

Even in the control in the A zone, the change from the high-pressure electric command hydraulic pressure corresponding to the duty ratio of 40% to the low-voltage electric command hydraulic pressure corresponding to the shift initial duty ratio d ₄ is performed by the division ratio k of the sensor 101. _2- (90/8) × 0.8
However, the stroke width is not limited to this, but may be appropriately determined depending on various conditions of the engine, the automatic transmission, and the like, and is not specified by the stroke of the piston 94 as described above. You may make it stipulate in a certain minute time.

By the way, after instructing the shift initial duty ratio d ₄ , the timer is set and the shift is further executed. After the timer is set, the rotation speed N _T of the turbine 10 and the rotation speed of the output shaft corresponding to the input shaft rotation speed. The rotational speed N ₀ of the transfer drive gear 60 corresponding to is detected, and in this case as well, it is determined whether or not the out-of-synchronization of the first speed defined by N _T ≦ N ₀ × 2.7 is achieved. If the result of this determination is that out-of-synchronization has not been achieved, the operation of subtracting the previously calculated subtraction value β from the initial shift duty ratio d ₄ so as to gradually increase the electrical command hydraulic pressure is repeated, Achieve out of sync.

At this time, since the throttle opening θt belongs to the A zone (high throttle opening region), the subtraction value β is obtained from the calculation result of the calculation formula {6.8 + 1.1 × (m−d ₁ ) / 2} Δt. Is set to a value that results in the solid line inclined portion in FIG. 4 (a). Since this value is larger than the subtraction value β (= 3) of the related art, it is possible to almost eliminate the feeling of braking at the time of shifting, thereby improving the shifting feeling.

The process of completing the shift to the second speed and the process of storing the constant α after the out-of-synchronization of the first speed is achieved are the same as those in the D zone (low throttle opening range).

Also, when the throttle opening θt belongs to the B zone or the C zone (intermediate throttle opening area), the above-mentioned arithmetic expression d ₄ = (d ₁ + 3 × m) / 4 and β = {6.8 + 1. 1 × (m−d ₁ ) / 2} Δt to d
₄ and β are calculated, and the shift hydraulic pressure control is performed using this calculation result. In this case as well, a smooth shift can be performed without generating a feeling of braking or a torque wave when shifting.

In this way, according to the shift initial duty ratio for the next shift based on the data of the previous shift, that is, the shift initial hydraulic pressure setting control that brings the shift initial hydraulic pressure closer to the optimum value, the exhaust amount, the output torque amount, etc. Even if an automatic transmission that does not conform to the standard engine is combined with this engine, the automatic transmission can be adapted to an engine that originally does not conform in the course of use. That is, if the initial hydraulic pressure for shifting is too high or too low, as soon as the shift start signal is transmitted, the frictional engagement elements become engaged, causing a great shift shock, or the shifting operation takes a long time. However, according to the above control, these problems are eliminated in the course of use, and the gear shift operation can be performed in any throttle opening range without causing a gear shift shock or a feeling of braking. That is, it is possible to optimize the torque capacity at the moment when the friction engagement elements are connected, and the shift feeling becomes extremely good.

3 (c) and 4 (c) are graphs showing changes in kick down servo hydraulic pressure in the low throttle opening region and the high throttle opening region, respectively. FIG. 6E is a graph showing changes in the kickdown drum pressing force in the low throttle opening region and the high throttle opening region, respectively.

Further, in the above example, the description has been given by taking the shift speed from the first speed to the second speed as an example, but it goes without saying that the shift speeds other than the above can be similarly implemented.

Further, in the calculation of the subtraction value β, the hydraulic pressure corresponding to the initial shift duty ratio d ₄ is increased as the hydraulic pressure is increased, and is increased as the engine output torque is increased. Well, subtraction value β
May be changed according to the driving state of another vehicle (for example, engine speed or vehicle speed).

It should be noted that the initial shift duty ratio d ₄ does not depend on the arithmetic expression (d ₁ + 3 × m) / 4, but other arithmetic expressions such as (ld ₁ + ym) / (x + y)% [where x, y
Is an arbitrary number] or other formula, and d ₄ may be set between d ₁ and m.

〔The invention's effect〕

As described above in detail, according to the shift hydraulic control device for the automatic transmission for a vehicle of the present invention, it is possible to supply the optimum hydraulic pressure to the friction engagement element when the friction engagement element is actuated to the engagement side. As a result, the conventional brake feeling is generated in one throttle opening range and the shift shock based on the torque wave is generated in the other throttle opening range. There is an advantage that a smooth shift feeling can be achieved without causing a shift shock.

[Brief description of drawings]

1 to 6 show a shift hydraulic pressure control device in an automatic transmission for a vehicle as an embodiment of the present invention. FIG. 1 is a schematic configuration diagram thereof, and FIG. 2 is a flow chart for explaining its operation. Fig. 3 compares the duty ratio change, electric command oil pressure change, kickdown servo oil pressure change, piston stroke change, drum pressing force change, and transmission output shaft torque change in the low throttle opening range with the conventional one. The graph shown in Fig. 4 shows changes in duty ratio, electric command oil pressure change, kickdown servo oil pressure change, piston stroke change, drum pressing force change, transmission output shaft torque change in the high throttle opening range. FIG. 5 is a graph for comparison with FIG. 5, FIG. 5 is a characteristic diagram for explaining the throttle opening zone, and FIG. 6 is a target change of the turbine rotation speed. 7 to 9 are graphs showing a shift hydraulic control device in a conventional vehicular automatic transmission, and FIG. 7 is a schematic configuration diagram showing an example of the vehicular automatic transmission, and FIG. FIG. 9 is a schematic configuration diagram showing a part of the hydraulic control circuit, FIG. 9 is a graph showing the duty ratio change, kickdown servo hydraulic pressure change, and piston stroke change, and FIG. 10 is another conventional automatic transmission for vehicles. Change of duty ratio when using variable speed hydraulic control device in
5 is a graph showing changes in kick down servo hydraulic pressure, changes in drum pressing force, and changes in transmission output shaft torque. 2 ... engine, 10 ... turbine, 20 ... input shaft,
30 ... kick down brake, 50 ... output shaft, 80
...... Kick down servo, 94 ...... Piston, 98 ......
Solenoid valve, 101 ... Piston stroke sensor, 200
...... Throttle sensor, 201 ...... Revolution speed sensor, 20
2 ... Sensors, AT ... Automatic transmission, C1 ... Control means, CL ... Controller, IOSM ... Initial hydraulic pressure setting means, OC ... Hydraulic pressure control circuit (hydraulic pressure adjusting means), OCM
...... Hydraulic pressure changing means, OCVM ...... Hydraulic pressure change rate varying means.

Claims (4)

[Claims] Claim: What is claimed is: 1. An input shaft to which rotational power of an engine is input, an output shaft to output rotational power to a drive wheel of a vehicle, and an optional rotating element to select between the input shaft and the output shaft. In a vehicular automatic transmission having a hydraulic friction engagement element for selectively switching the gear ratio of, a hydraulic pressure adjusting means for adjusting the hydraulic pressure supplied to the friction engagement element according to an electric signal, and A detection unit that detects the rotation speed of the rotating element whose rotation speed changes, and a hydraulic pressure to the friction engagement element so that the change rate of the rotation speed detected by the detection unit follows a preset target change rate. The initial hydraulic pressure setting means is provided with a control means for electrically performing feedback control, and an initial hydraulic pressure setting means for setting a value of the initial hydraulic pressure to be supplied to the friction engagement element at the start of gear shifting. The first electric control value for determining the third electric control value for setting the initial hydraulic pressure obtained from the second electric control value at the required time point of the feedback control, and the minimum hydraulic pressure for driving the movable member of the friction engagement element. A hydraulic pressure control device for a vehicle automatic transmission, characterized in that it is provided with a means for setting it to a fourth electric control value corresponding to. 2. An input shaft for inputting rotational power of an engine, an output shaft for outputting rotational power to driving wheels of a vehicle, and an optional rotating element to select between the input shaft and the output shaft. In a vehicular automatic transmission having a hydraulic friction engagement element for selectively switching the gear ratio of, a hydraulic pressure adjusting means for adjusting the hydraulic pressure supplied to the friction engagement element according to an electric signal, and A detection unit that detects the rotation speed of the rotating element whose rotation speed changes, and a hydraulic pressure to the friction engagement element so that the change rate of the rotation speed detected by the detection unit follows a preset target change rate. An initial hydraulic pressure setting means for setting a value of an initial hydraulic pressure to be supplied to the friction engagement element at the start of a gear shift, and an initial hydraulic pressure set by the initial hydraulic pressure setting means. The hydraulic pressure changing means for changing the hydraulic pressure at a desired ratio after supplying the electric pressure to the friction engagement element, and the initial hydraulic pressure setting means feeds back the first electric control value for determining the initial hydraulic pressure. Between the third electric control value for setting the initial hydraulic pressure obtained from the second electric control value at the required time of control and the fourth electric control value corresponding to the minimum hydraulic pressure for driving the movable member of the friction engagement element. A vehicle automatic transmission, characterized in that it is provided with setting means, and the hydraulic pressure changing means is provided with hydraulic pressure change rate varying means for changing the hydraulic pressure change rate according to the operating state of the vehicle. Shift hydraulic control device for aircraft. 3. The hydraulic pressure change rate varying means is provided with means for temporally changing the hydraulic pressure with an inclination that increases as the hydraulic pressure corresponding to the first electric control value increases. A shift hydraulic control device for an automatic transmission for a vehicle according to claim 2. 4. The hydraulic pressure change rate varying means is provided with means for temporally varying the hydraulic pressure with an inclination that increases as the output torque of the engine increases. 13. A shift hydraulic control device for an automatic transmission for a vehicle according to claim 15.