JPH03134366A - Controller of automatic transmission - Google Patents

Abstract

PURPOSE:To stabilize the changeover timing of a hydraulic friction element by controlling the quantity of oil to be supplied into a piston cylinder to the maximum value during the detection of a coupling starting point at an engine mount displacement speed, at the time of coupling of a hydraulic friction element. CONSTITUTION:To a control unit 5, rotation pulses 51-53 of the impeller, turbine, and output shaft of a torque converter 2, a shift range signal 51, a throttle opening degree signal 55, and the mount displacement signal 56 of an engine 1 corresponding to the reaction to the driving of the output shaft 32 of an auxiliary gearbox 3 are inputted, and based on the mount displacement speed value, it detects the coupling starting point of a hydraulic friction element. And, the discharge of the electro-hydraulic conversion value of a hydraulic controller 4 is set to the maximum value until the hydraulic friction element reaches the coupling starting point since the issue of a speed change instruction. In this way, the invalid stroke time of a piston accompanying the coupling operation of the hydraulic friction element is controlled shortest, and changeover timing can be done stably.

Description

DETAILED DESCRIPTION OF THE INVENTION [Field of Industrial Application] The present invention relates to a control device for an automatic transmission for smoothing the switching of hydraulic friction elements during gear shifting of the automatic transmission.

[Prior art] As a conventional automatic transmission control device, for example,
There is one disclosed in Japanese Patent No. 2-177953.

This disclosed device provides pressure sensors on the input boat and the output boat of an electro-hydraulic conversion valve arranged in a hydraulic pressure supply circuit of a hydraulic friction element, and supplies pressure from the electro-hydraulic conversion valve to the hydraulic friction element. The oil volume is measured from the pressure difference, and the discharge amount of the electro-hydraulic conversion valve is output at the maximum value until the friction element exceeds the allowable stroke volume until the friction element is substantially engaged. This shortens the time it takes to complete the fastening operation.

[Problem to be solved by the invention]

Conventional automatic transmission control devices are configured as described above, so in the case of a configuration in which an electro-hydraulic conversion valve is arranged for each of multiple hydraulic friction elements, a large number of pressure sensors are required, resulting in high costs. was there. In addition, it is not possible to shorten the fastening operation time due to changes over time and individual differences in the allowable stroke volume of hydraulic friction elements, and when measuring the flow rate from the pressure difference, the flow rate coefficient changes depending on the oil temperature. There were problems such as easy measurement errors and unstable switching timing of hydraulic friction elements.

This invention was made to solve the above-mentioned problems, and it is possible to shorten the piston invalid stroke time associated with the coupling operation of the hydraulic friction elements during gear shifting, and to stabilize the switching timing of the hydraulic friction elements. The purpose of this invention is to obtain a control device for an automatic transmission that can be obtained by using the following methods.

[Means to solve the problem]

The automatic transmission control device according to the present invention includes a mount displacement detecting means for detecting a mount displacement of an engine, a mount displacement speed calculating means for calculating a displacement speed of the mount displacement, and a calculated value from the mount displacement calculating means. , a coupling start point detection means for detecting the coupling start state of the hydraulic friction elements, an electrohydraulic conversion valve for controlling the oil pressure supplied to the hydraulic friction elements, and a coupling start point detection means for detecting the coupling start state of the hydraulic friction elements The electrohydraulic conversion valve control means sets the discharge amount of the electrohydraulic conversion valve to the maximum value until the coupling start point is detected.

[For production]

In the transmission control device according to the present invention, the connection start state of the hydraulic friction element is determined by the connection start point detection means based on the mount displacement speed value obtained from the engine mount displacement detection means via the mount displacement speed calculation means. By detecting this, the electro-hydraulic conversion valve control means sets the discharge amount of the electro-hydraulic conversion valve to the maximum value and performs the coupling operation of the hydraulic friction elements from the time of the shift command until the hydraulic friction elements reach the coupling start point. The piston's invalid stroke time associated with the piston is controlled to be the shortest possible time, and the switching timing of the hydraulic friction element is stably performed.

〔Example〕

An embodiment of the present invention will be described below with reference to the drawings. 1st
The figure is a block diagram showing the general configuration of the main parts of this invention.
In the figure, 1 is an engine, and 2 is an input element (pump impeller) 21 driven by the engine 1 that rotates internal hydraulic oil, and this hydraulic oil drives an output element (turbine launch) 22 under a reaction force from a stator (not shown). This is a torque converter that rotates while increasing torque.

Reference numeral 3 denotes an auxiliary transmission, in which the shaft connected to the turbine runner 22 of the torque converter 2 is used as a transmission input shaft 31, and the shaft that transmits power to the tires is used as a transmission output shaft 32, and provides a gear ratio according to the running condition of the vehicle. It is.

Reference numeral 4 denotes a hydraulic control section for controlling the hydraulic pressure supplied to a hydraulic friction element (not shown) in the auxiliary transmission 3.

Reference numeral 5 indicates input signals such as a pump impeller rotation pulse 51, a turbine runner rotation pulse 52, and an auxiliary transmission output shaft rotation pulse 53 of the torque converter 2 detected from gears and an electromagnetic pickup (not shown) and at a selected position of the shift lever. It has a plurality of switches to close, and is connected to a shift range signal 54 for detecting the position of the switch in the open/closed state and to the accelerator pedal, and a potentiometer is used to detect the voltage value of the engine throttle opening, which changes in response to the amount of pedal depression. The throttle opening signal 55 output from the throttle opening sensor, which is converted into and detected by the throttle opening sensor, and the driving reaction force of the auxiliary transmission output shaft act on the engine mount as rolling rotation of the engine, and in response to the driving reaction force, An engine mount displacement signal 56 output from a mount displacement detection means that detects the changing amount of engine mount displacement by converting it into a voltage value using a potentiometer is inputted, and based on this information, the engine mount displacement signal 56 is detected by converting the amount of displacement of the engine mount into a voltage value using a potentiometer. Hydraulic control signal 57 necessary for determining the gear position and shifting the gear position
This is a control unit whose main component is a microcomputer that outputs the signal as an output signal to a solenoid valve (not shown) in the hydraulic control device 4.

Figure 2 shows double pinion planetary gear set 1.
FIG. 2 is a schematic diagram of an auxiliary transmission with three forward speeds using a set of gears. In FIG. 2, 20 is an engine crankshaft driven by the engine and rotates together with the pump impeller 21 of the torque converter 2, and 31 rotates together with the turbine runner 22 of the torque converter 2 to provide driving force to the auxiliary transmission 3. It is also the output shaft of the torque converter that transmits the power, and the input shaft of the auxiliary transmission.

33 is a double pinion type planetary gear cent.
This planetary gear set 33 is the forward sun gear 3
4, a reverse sun gear 35, a ring gear 36, a short pinion 37, a long pinion 38, and a carrier 39.

The forward sun gear 34 is connected to the rear clutch C2, and the reverse sun gear 35 is connected to the front clutch C1.
The output shafts 31 of the torque converter 2 are connected to the output shafts 31 of the torque converter 2, respectively.

Also, Bl is the low reverse brake, B2 is the second brake, 3C is the transmission case, and the low reverse brake is
The reverse brake B1 fixes and releases the carrier 39 to the case 3C, and the second brake B2 fixes and releases the reverse sun gear 35 to the case 3C. The ring gear 36 is coupled to the output shaft 32 of the auxiliary transmission 3.

FIG. 3 is a block diagram of the hydraulic control circuit of this automatic transmission, and 40 in the figure is an oil pump provided on the crankshaft 20 of the engine l.

41 is a pressure regulator valve for regulating the oil pressure generated by the oil pump 40 as line pressure, 42 is a shift control valve that is directly connected to the shift lever, and 43 to 45 are a low reverse brake B1 and a second brake B2. , is a duty solenoid valve for controlling the hydraulic pressure supplied to the front clutch C1, for example, by duty control.

C2 is a rear clutch, and when moving forward, hydraulic pressure is supplied through the shift control valve and the engine is engaged.

The combination of the coupling operations of the friction elements in each gear stage is such that the rear clutch C2 is engaged in forward motion, the low reverse brake B1 is engaged in 1st gear, the second brake B2 is in 2nd gear, and the front clutch C1 is engaged in 3rd gear. Become. When reversing the vehicle, the front clutch C1 and low reverse brake B1 are engaged.

FIG. 4 shows the torque capacity coefficient C( of the torque ratio t r(-T t/T "'T F/NP") as the speed ratio e
1 is a torque converter performance curve expressed as a function of . The speed ratio e≦1.0 is a region where the engine is on the driving side, and e
>1.0 is the driven side.

FIG. 5 shows the relationship between the control hydraulic pressure P and the duty ratio (valve opening/closing ratio) when the hydraulic pressure supplied to the clutch or brake is controlled by a duty solenoid valve.

Figure 6 shows that the transmission torque Te of a wet multi-disc clutch or brake is expressed as Tc = P (-A-R, n ・u, and since the transmission torque T is proportional to the control hydraulic pressure PC, The relationship between the transmission torque Tc and the duty ratio of the duty solenoid valve is shown based on the characteristics shown in the figure.

Here, PC is the control oil pressure, A-R-n is the piston pressure-receiving area of the clutch or brake, the average radius of the friction surfaces, and the number of friction surfaces, and μ is the friction coefficient.

If the number of teeth of the forward sun gear 34 of the planetary gear set 33 is zps, the number of teeth of the reverse sun gear 35 is Z□, and the number of teeth of the ring gear 36 is Z++c, the reduction ratio i at 1st speed is I + = Z JIG/ Z rs The reduction ratio 12 at 2nd speed is i z = Z Ill; (Z FS ÷
It is expressed as Z*s)/Zps (Z++s + Z*c).

The reaction torque TBI of the low reverse brake B1 and the reaction torque TB2 of the second brake B2 have the following relationship with respect to the turbine torque Tt.

TBI = (it-1)・Tz-(it-1) /
(it-1)・TB2 ... (1) TB2-(it
-1)・T, ...(2) Next, when the reaction torques TBI and TB2 are expressed by the duty ratios DB,, DB, which are the control signals of the duty solenoid valves, from the relationship shown in Fig. 6, TB1= α・DB,,TB2
=β・DB2...It can be expressed by the approximate expression (3).

Therefore, the brake torques TBI and TB2 are converted into a duty ratio DB, which is a control signal for the duty solenoid valve.
, DB, is expressed as below, where the control duty rate DB + of the duty solenoid valve 43 for the low reverse brake B1 is the control duty rate of the duty solenoid valve 44 for the turbine torque TL and the second brake B2. D B
Calculate from z and output.

The control duty rate DB of the duty solenoid valve 44 for the second brake B2 is calculated from the turbine torque Tt and outputted, so that, for example, both engagements in the friction element switching process during gear shifting with the engine on the drive side are controlled. It also fully tracks sudden changes in load caused by accelerator operation during gear changes.

Next, the operation will be explained using an upshift from 1st speed to 2nd speed as an example. In the speed state, the rear clutch C2 and the low reverse brake B1 are connected.

For example, when the engine is on the drive side, shifting from 1 to 2,
From this state, by engaging the second brake B2 while releasing the low reverse brake B1, the friction elements are smoothly switched, and the brake hydraulic pressure is adjusted to achieve 2nd gear while keeping output shaft torque fluctuations small. control.

A flowchart of this control is shown in FIG. Phase 0.1°23 used in this flowchart shows the gear shifting process.

Phase 0: A state before shifting.

Phase 1: The low reverse brake B1 is connected with a reaction force meter corresponding to the turbine torque T, and the second brake B2 is in a released state with the piston cylinder filled with oil and the torque being zero. .

Phase 2: The second brake B2 starts to be engaged, and the low reverse brake B1 is gradually released accordingly (indicating the so-called torque phase (a state where the speed ratio is 1st gear and only the torque ratio shifts to 2nd gear). .

Phase 3: The engagement of the second brake B2 further increases and the reaction force of the low reverse brake B1 becomes zero, indicating a so-called inertia phase (state B in which the speed ratio shifts to second speed) where gear shifting begins.

FIG. 7(a) is the main flowchart, in which step fI
At step ft, the throttle opening amount and engine mount displacement amount are read from the throttle opening signal 55 and engine mount displacement signal 56, and at step ft, the output of the input shaft 20 of the torque converter 2, the output horn shaft 31, and the auxiliary transmission 3 is read. The rotation pulses 51 to 53 of each shaft detected from the configuration of the gear (not shown) provided on the shaft 32 and the electromagnetic pink "up" are measured by the control unit 5 to measure the average pulse period at each fixed sampling period, and the rotation of the pump of the torque converter 2 is determined by the control unit 5. Speed (engine rotation speed a) N F 7 Turbine rotation speed N
t, output shaft rotational speed of the auxiliary transmission 3 (equivalent to vehicle speed) N.

Calculate.

In step f3, the turbine torque Tt is calculated based on the rotational speeds NA and NY of the torque converter 2 obtained in step f2. The torque calculation flowchart is shown in FIG.

In this FIG. 7(b), in step f6, the speed ratio e (=NT/NP) of the torque converter 2 is calculated, and in step ft, the operating state of the torque converter 2 is determined whether the engine is on the driving side or the driven side. The speed ratio e
Judgment is based on the value of .

As a result of this determination, if the engine is on the drive side (e≦1), the torque ratio 1.0, which is a function of the speed ratio e, is determined from the torque converter performance curve map of FIG. 4 preset in step f. ) Read the torque capacity coefficient C, and assume that the turbine torque T is a positive torque: T t = t r (Il
Calculate by l・C,,)・N,.

On the other hand, when the engine is on the driven side (e>1), the turbine torque TL is calculated as a negative torque in step r in the same manner as in step r.

Subsequently, in step f of the main flowchart in FIG. 7(a), the throttle opening amount obtained in step f1 and the transmission output shaft rotational speed N+1 (
(equivalent to vehicle speed) and compared with the shift pattern characteristics (not shown).
When it is determined that a shift from 1st speed to 2nd speed is necessary, the process proceeds to step f, where the shift is executed.

An execution flowchart of the speed change control is shown in FIG. 7(C).

In step flG, the speed change process is determined based on the phase value, and processing is performed for each phase. When a shift command is issued, the phase is 0, and steps f to f14 are performed only once.

In step [11, the duty rate DB of the duty solenoid valve 43 for the low reverse brake Bl,
Since the second brake B2 is in the released state, DB,
-0 and is calculated from the entire right side of equation (4) above and output.

In step r+z, in order to fill the piston cylinder of the second brake B2 with oil by setting the discharge amount of the valve to the maximum value, the duty rate DB! of the duty solenoid valve 44 is increased. Output as 1=100%.

In step f13, when the engine load is small, the action on the engine mount due to the drive reaction force of the auxiliary transmission output shaft is small, and the mount displacement speed is minute, it is difficult to detect the connection start point of the second brake B2. In order to set the oil filling time of the second brake B2 to a predetermined value, the timer is cleared and started. In step r14, the shift process is set to phase 1 and the process returns to the main routine.

Next, when the shift process moves to phase 1, the process proceeds to step f15, and in step "5., first, the second brake B2
In step f, check whether the engine is under a load condition that allows the connection start point of the engine mount to be reliably detected at a high displacement speed of the engine mount.
The turbine torque value T obtained at , and the predetermined torque value T. This is done by comparing with. If T, ≧T1゜,
The mount displacement speed large is calculated by calculating the difference between the engine mount displacement amounts obtained in step f1, and the mount displacement speed large is determined in advance. It is determined whether the second brake B2 has reached the coupling start point.

5c, < sentence,. If , go to step f and write the sentence,
In the case of 2 sentences, the timer value at this time is used as the second brake B.
The learning value for the oil filling time in step 2 is set to , and the process proceeds to step rat. If T t < T t o, determine whether the timer value has reached the set value Lt of the oil filling time of the second brake B2, and if 1 < 11, proceed to step f16. is step f17
Proceed to.

In step f+a, the duty rate DB++ of the duty solenoid pulp 43 is calculated and output in the same manner as described above, and the hydraulic pressure of the low/reverse brake B1 is controlled, while the duty solenoid valve 44 is set to the duty rate DBz+=1.
00% and fills the piston cylinder of the second brake B2 with oil.

At step f+t, the phase is set to 2 to set the transition of the gear shifting process. In step rlII, second brake B
The filling of oil into the piston cylinder of No. 2 has been completed, and the connection is now possible, and the duty rate of the duty solenoid pulp 44 for the second brake B2 is
z is calculated from equation (5), and DBzt is gradually increased from 0% and output so as to reach this DBt within a predetermined time.

In step tlq, the duty rate DB of the solenoid valve 43 is calculated and output from equation (4) using the value of the duty rate DBzz, and the process returns to the main routine.

When the gear shifting process moves to phase 2, the so-called torque phase, the process proceeds to step tzo, and in step fzo, as the engagement of the second brake B2 increases, the reaction torque TB1 of the low reverse brake B1 decreases, and the reaction torque TB1 decreases. When it reaches zero, the gear shift starts, so the turbine rotational speed N〒 becomes 1 dislodging rotational speed N t+ (-i 1
・Determine whether it has become below (NO).

When the turbine rotation speed N Y =N T + is still in the process of the torque phase, steps [18 and flq] are performed.

When N t < N t + is detected, the process proceeds to step f, and in step rz+, the phase is set to 3 in order to shift to the inertia phase.

In step f2□, in order to keep the low reverse brake B1 in a released state (hydraulic pressure zero), the duty ratio DB of the solenoid valve 43,! =0% and output.

In step ris, the turbine rotational speed N? The duty solenoid valve 4 includes a term for feedback controlling the rate of decrease in the turbine torque T to a specified target value and a term for correcting the change ΔTt in the turbine torque T.
608g3=-Ea+(it 1)・Δ as correction duty rate ΔD B z x of 4
Tt...(6) Calculate from β, add to duty rate DB, 3, and output.

In this equation (6), K is the control gain, and El is the turbine rotation speed N? is the deviation of the rate of decline from the target value.

When the speed change process shifts to phase-3, so-called inertia phase, the process proceeds to step ft4, and in step rz4, the turbine rotational speed N7 and the second gear synchronous rotational speed Nyg
(= 1t-No) and N t > N t !
If , step f. Then, in step rzs, the above-mentioned inertia phase processing is performed.

N? -N T H, proceed to step rzs;
Duty rate DB of duty solenoid valve 44,
4-100%, keep the second brake B2 in the connected state, clear the phase in step f0, and set it to 1-100%.
2-speed control is completed.

Figure 8 shows a time chart of the 1-2 shift process, and Figure 8 (a) shows the duty solenoid valve 4.
3, the duty rate DBI, and FIG. 8 (c) is the control oil pressure PB of the low reverse brake B1.

Fig. 8(C) shows the duty ratio DB of the duty solenoid valve 44, and Fig. 8(d) shows the second brake B2.
The control oil pressure PB,, Fig. 8(e) shows the turbine rotation speed N?
, Fig. 8(f) is the engine mount displacement X, , Fig. 8(f) is the engine mount displacement
g) shows the shift control phase.

Note that in the above embodiment, engine mount displacement is detected by a potentiometer, but it may be detected by a non-contact displacement sensor or a differential transformer. Further, although the driving reaction force of the auxiliary transmission output shaft is detected as the amount of displacement of the engine mount, the same effect as in the above embodiment can be obtained even if this is detected as the strain stress acting on the engine mount.

〔Effect of the invention〕

As described above, according to the present invention, when a hydraulic friction element is coupled, the amount of oil supplied into the piston cylinder is changed to the discharge amount of the electro-hydraulic conversion valve until the coupling start point is detected based on the displacement speed of the engine mount. Since the configuration is configured to control by setting the value to the maximum value, the invalid stroke of the piston due to the coupling operation is controlled against individual differences in allowable stroke volume of hydraulic friction elements, changes over time, changes in valve discharge flow rate due to oil temperature, etc. This has excellent effects such as minimizing time and stabilizing the switching timing of the hydraulic FR friction element.

[Brief explanation of the drawing]

Fig. 1 is a block diagram showing the main part configuration of an automatic transmission control device according to an embodiment of the present invention, and Fig. 2 is a forward drive gear shift system using a double pinion type planetary gear set applied to the above embodiment. Fig. 3 is a block diagram of the hydraulic control circuit of the automatic transmission applied to the above embodiment, Fig. 4 is a performance curve diagram of the torque converter in the above embodiment, and Fig. 5 is a schematic diagram of the auxiliary transmission. A characteristic diagram of the control hydraulic pressure of the duty solenoid valve applied to the above embodiment, Fig. 6 is a characteristic diagram of the transmission torque of the hydraulic friction element and a control signal of the solenoid valve, and Fig. 7 (a) is the main characteristic of the above embodiment. Figure 7 (C) is a diagram showing a flowchart of the torque calculation according to the above embodiment, Figure 7 (C) is a flowchart showing a shift execution flowchart of the embodiment above, and Figure 8 is a time chart when performing a shift. FIG. 1...Engine, 2...Torque converter, 3...
- Auxiliary transmission, 4... Hydraulic control section, 5... Control unit, 40... Oil pump, 43-45... Duty solenoid valve, 56... Mount displacement signal, B1...・Low reverse brake, B2...Second brake, C1...Front clutch, C2...
...Rear clutch. Note that the same symbols in the figures indicate the same or equivalent parts.

Claims (1)

[Claims] By releasing one hydraulic friction element and connecting another hydraulic friction element, shifting from a gear position corresponding to the one hydraulic friction element to a gear position corresponding to the another hydraulic friction element. A control device for an automatic transmission configured to perform the following: mount displacement detection means for detecting mount displacement of the engine; mount displacement speed calculation means for calculating the displacement speed of the mount displacement; coupling start point detection means for detecting a coupling start state of the another hydraulic friction element based on the calculated value;
The discharge amount of the electro-hydraulic conversion valve is set to a maximum value until a coupling start point between the electro-hydraulic conversion valve that controls the hydraulic pressure supplied to the other hydraulic friction element and the other hydraulic friction element is detected. 1. A control device for an automatic transmission, characterized by comprising an electro-hydraulic conversion valve control means.