JPH03194263A - Lockup clutch connecting force control device for fluid coupling - Google Patents

Abstract

PURPOSE:To suppress a torque shock by suppressing an increase of lockup connecting force when a fluctuation amount of detected turbine torque is increased to a preset value or more, when a lockup clutch of a fluid coupling is connected. CONSTITUTION:When a lockup clutch 28 is put in connection, its connecting force is controlled by a connecting force control means 120, but torque of a turbine 23 in a fluid coupling 2 at this time is detected by a detecting means 121. The level of a fluctuation amount of this detected turbine torque in comparison with a preset value is decided, and when the fluctuation amount is below the preset value, connecting force by the control means 120 is increased as it is by a connecting force suppressing means 122, but when the fluctuation amount of the turbine torque is increased beyond the preset value, tightening force of the lockup clutch 28 is suppressed from increasing. Thus, a shock can be effectively suppressed when the lockup clutch is put in connection.

Description

DETAILED DESCRIPTION OF THE INVENTION (Field of Industrial Application) The present invention relates to an improvement in a control device for controlling the engagement force of a lock-up clutch in a fluid coupling.

(Prior Art) Conventionally, in a torque converter as a fluid coupling disposed between the input shaft of a vehicle automatic transmission and the output shaft of an engine, a lock-up clutch that directly connects the pump and turbine using hydraulic pressure has been used. It is well known that a lock-up clutch is provided to suppress slippage of the torque converter and improve fuel efficiency.

However, in a fluid coupling equipped with a lock-up clutch, when the lock-up clutch is engaged and the pump and turbine are engaged, the turbine torque changes rapidly, and this torque change causes a shock. This is transmitted to the vehicle body, causing discomfort to the occupants.

Therefore, as disclosed in Japanese Patent Application Laid-Open No. 60-1.59466, when the lock-up clutch is engaged, the hydraulic pressure is controlled by a duty solenoid valve, and the duty ratio is sequentially varied from 0 to 100%. A method has been proposed that aims to reduce the torque shock when the lock-up clutch is engaged by adjusting it.

(Problems to be Solved by the Invention) However, in this proposal, although the lock-up clutch is subjected to duty control, the control is open control. Therefore, variations in control cannot be avoided, and in reality, it is difficult to reliably reduce fluctuations in turbine torque when the lock-up clutch is engaged and effectively suppress torque shock.

The present invention has been made in view of the above points, and its purpose is to suppress fluctuations in turbine torque when the lock-up clutch is engaged and ensure torque shock by changing the control configuration when the lock-up clutch is engaged. The aim is to suppress this.

(Means for Solving the Problem) In order to achieve the above object, the invention according to claim (1) detects fluctuations in the actual turbine torque of the fluid coupling, and when the turbine torque changes beyond a set value, , an increase in the engagement force of the lock-up clutch is suppressed.

Specifically, the present invention provides a pump 2 as shown in FIG.
A lock-up clutch 2 that connects the turbine 23 and the lock-up clutch 2
The fluid coupling 2 is configured such that the fastening force of 8 is controlled by the fastening force control means 120. When the engagement force of the clutch 28 increases, the turbine torque "rQ"
Engagement force suppressing means 122 is provided to control the engagement force control means 120 so that an increase in the engagement force of the lock-up clutch 28 is suppressed when the amount of change in T LI R increases beyond a set value.

(Function) With the above configuration, in the present invention, when the lock-up clutch 28 is engaged, the engagement force is controlled by the engagement force control means 120, but at this time, the torque To'ruR of the turbine 23 in the fluid coupling 2 is It is detected by the turbine torque detection means 121. It is determined whether the amount of change in the detected turbine torque 'ro'ruR is greater than or equal to the set value, and when the amount of change is lower than the set value, the fastening force suppressing means 122 controls the fastening force by the fastening force controlling means 120. Although it continues to increase, the turbine torque TQTLJ
When the amount of change in R increases more than the set value, an increase in the engagement force of the lock-up clutch 28 is suppressed. Therefore, when the lock-up clutch 28 is engaged, the turbine torque TQTUR always increases gradually by an amount of increase below a certain level, and its rapid increase is reliably avoided, thereby effectively suppressing the shock when the lock-up clutch 28 is engaged. can do.

(Example) Hereinafter, an example of the present invention will be described based on the drawings from FIG. 2 onwards.

FIG. 2 shows the overall structure of a transmission according to an embodiment in which the present invention is applied to an automobile. The transmission shown in the figure basically consists of a torque converter 2 connected to an output shaft 11 of an on-vehicle engine 1.
, a forward/reverse switching mechanism 3, a continuously variable transmission mechanism 4, a speed reduction mechanism 5, and a differential mechanism 6.

The torque converter 2 constitutes a fluid coupling in the present invention, and as shown in enlarged detail in FIG. engine output shaft 11
A pump impeller 22 rotates integrally with the pump impeller 22 , a turbine launch 23 rotatably provided inside the pump cover 21 to face the pump impeller 22 , and an intervening structure between the turbine runner 23 and the pump impeller 22 . It has a stator 24 that is installed to perform a torque increasing effect, and a turbine shaft 25 that is fixed to a turbine runner 23.

The stator 24 is connected to the transmission case 7 via a one-way clutch 26 and a cylindrical stator shaft 27. The turbine runner 23 and pump cover 21
A lock-up clutch 28 having a lock-up piston 28a slidably attached to the turbine shaft 25 is provided between the lock-up piston 28 and the lock-up clutch 28.
The lock-up clutch 28 is engaged and released by introducing and discharging hydraulic pressure into a lock-up engagement chamber 29a and a lock-up release chamber 29b formed on both sides of the lock-up clutch 28.

The forward/reverse switching mechanism 3 includes a ring gear 35 splined to the turbine shaft 25 of the torque converter 2, and a binion 32□ 32. which meshes with the ring gear 35.・
...and the pinion 32. 32°..., and a sun gear 34 spline-coupled to a primary shaft 411 of a continuously variable transmission mechanism 4, which will be described later, and meshing with the pinion gear 32. A forward clutch 36 is disposed between the ring gear 35 and the carrier 31 to connect and disconnect them, and a reverse clutch 36 is disposed between the carrier 31 and the transmission case 7 to selectively fix the carrier 31 to the transmission case 7. A brake 37 is provided. When the forward clutch 36 is engaged and the reverse brake 37 is released, the ring gear 35 and carrier 31 are connected to the rotating body, and the rotation of the turbine shaft 25 is directly transmitted to the continuously variable transmission mechanism 4. On the other hand, when the reverse brake 37 is engaged and the forward clutch 36 is released, the carrier 31 is fixed to the case 7 so as not to rotate, and the rotation of the ring gear 35 is transmitted to the pinion gear 32. 32. Sun gear 34 via...
The rotation of the turbine shaft 25 is reversed and transmitted to the primary shaft 411 of the continuously variable transmission mechanism 4. In addition, the forward clutch 36 and the reverse brake 3
7 are both opened, a neutral state and a parking state are established in which the driving force of the engine is not transmitted from the turbine shaft 25 to the primary shaft 411 of the continuously variable transmission mechanism 4.

The continuously variable transmission mechanism 4 includes a primary pulley 41 as a driving pulley, a secondary pulley 42 as a driven pulley, and a V-belt 43 wound between these pulleys 41 and 42. The primary pulley 41 includes a primary shaft 411 as a drive shaft disposed coaxially with the turbine shaft 25, and a primary shaft 411 as a drive shaft disposed coaxially with the turbine shaft 25.
A fixed sheave 412 is rotatably fixed to the fixed sheave 412, and a fixed sheave 412 is disposed opposite to the fixed sheave 4]2, and the primary shaft 4
11 and a movable sheave 413 rotatably and slidably supported, and as the movable sheave 413 moves, the pitch line of the V-belt 43 changes, and the effective pitch diameter (effective radius) changes. It looks like this. That is, when the movable sheave 413 approaches the fixed sheave 412, the effective pitch diameter becomes large, and conversely, when the movable sheave 4]3 moves away from the fixed sheave 412, the effective pitch diameter becomes small.

On the other hand, the secondary pulley 42 basically has the same configuration as the primary pulley 41 described above.

That is, a secondary shaft 421 as a driven shaft arranged parallel to the primary shaft 411, and this secondary shaft 42
1 and a movable sheave 423 that is slidably supported.
The effective pitch diameter is changed by the movement of .

Each of these pulleys 41. , 42, each movable sheave 41
3,423, each movable sheave 413,
Hydraulic cylinder 41.4, 424 sliding 423
is provided. To specifically explain the configuration of the hydraulic cylinder 414 of the primary pulley 41, the hydraulic cylinder 414 has a first oil chamber 14a formed on the back side of the movable sheave 413, and a connecting member at the outer end of the movable sheave 413. Movable piston 1 movable and integrally connected via 14b
4c, and a second oil chamber 14d formed on the back side of the movable piston 14c.
The movable sheave 413 is moved in the right direction in FIG. 3 by supplying hydraulic pressure to both d and d.

On the other hand, the configuration of the hydraulic cylinder 424 of the secondary pulley 42 is similar to that described above, and is connected to the first oil chamber 15a formed on the back side of the movable sheave 423 and the outer end of the movable sheave 423 via the connecting member 15b. movable piston 15c
and a second oil chamber 1.5d formed on the back side of the movable pistons 1, 5c, and compressed into the second oil chamber 15d to urge the movable piston 15c to the left in FIG. 3. Spring 1
5e and 0, and is configured to supply hydraulic pressure to the first and second oil chambers 15a and 1.5d to move the movable sheave 423 to the left in FIG.

The hydraulic cylinder 41 of the primary pulley 41
The total pressure receiving area of the first oil chamber 14a and the second oil chamber 14d of
The area is set to be approximately twice the total pressure receiving area of the oil chamber 15a and the second oil chamber 15d, and when hydraulic pressure is introduced into the hydraulic cylinder 414 of the primary pulley 41, the movable sheave 4
13 moves to the right in FIG. 3 and approaches the fixed sheave 412, the pitch line of the V belt 43 in the primary pulley 41 moves to the outer circumferential side, and the effective pitch diameter of the primary pulley 41 increases. Due to the accompanying displacement of the V-belt 43, the movable sheave 423 of the secondary pulley 42 moves to the right in FIG. The effective pitch diameter of the secondary pulley 42 becomes smaller, and the gear ratio between the primary and secondary shafts 41.1, 421 changes smaller (in the direction of speed increase). Conversely, when the hydraulic pressure is discharged from the hydraulic cylinder 414, the movable sheave 413 of the primary pulley 41 moves to the left in FIG.
12, and its effective pitch diameter becomes smaller, while the movable sheave 423 of the secondary pulley 42 also moves to the right in FIG. 41
The gear ratio between 1.421 and 1.421 changes significantly (in the direction of deceleration).

Further, the speed reduction mechanism 5 and the differential mechanism 6 have a known structure,
Although not described in detail here, the rotation of the secondary shaft 421 is transmitted to the axle 6].

Next, the torque converter 2 in the above-mentioned continuously variable transmission
, the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3, and the primary pulley 41 and secondary pulley 4 of the continuously variable transmission mechanism 4.
A hydraulic circuit for controlling each operation of the motor 2 will be explained based on FIG. 4.

2 The hydraulic circuit shown in the figure includes an oil pump 81 driven by the engine 1. The hydraulic oil discharged from the oil pump 81 is first adjusted to a predetermined line pressure by the line pressure regulating valve 82, and then supplied to the hydraulic cylinder 424 of the secondary pulley 42 via the line 101. The hydraulic cylinder 4 of the primary pulley 41 is finally connected to the hydraulic cylinder 4 of the primary pulley 41 via a line 102 branched from 01.
14.

The line pressure regulating valve 82 has a spool 8 composed of a main spool 821 and a sub spool 822 arranged in series.
It has 20. A main spool 821 and a sub-spool 822 constituting the spool 820 are connected such that one end of the sub-spool 822 is brought into contact with one end of the main spool 821. At the other end of the sub spool 822,
A large diameter portion 822a having a larger cross-sectional area than the contact area (cross-sectional area of the connecting portion) with the main spool 82 is provided. At a position corresponding to the center of the main spool 821, there are provided a pressure regulating port 823 through which oil discharged from the oil pump 81 is introduced, and a drain boat 824 communicating with the suction side of the oil pump 81. If it moves to the left side in the figure, the pressure regulating port 823 and the drain port 824 are cut off, and the main spool 821
If it moves to the right side in the diagram, the pressure regulating port 823 and drain boat 8
24 is cut off, and when the main spool 821 moves to the right side in the figure, the pressure regulating port 823 and the drain port 824 are communicated with each other. A first
A pilot chamber 825 is formed, and a spring 826 is interposed in this first pilot chamber 825 to urge the main spool 821 to the left in the figure. In addition, the sub spool 8
A second pilot chamber 827 communicating with the first pilot chamber 825 is formed in the large diameter portion 822a of the second pilot chamber 822. These first pilot chamber 825 and second pilot chamber 82
7, the hydraulic oil is reduced to a predetermined pressure by the reducing valve 83 while passing through the line 103 after branching from the line 102. While passing through the pilot passage 4 103a, the first duty solenoid valve 9
It is designed to be introduced as a pilot pressure adjusted in step 1. While this pilot pressure acts in the same direction as the biasing force of the spring 826, the hydraulic pressure in the line 101 acts on the other end of the main spool 821 to counteract the biasing force and the pilot pressure. By moving the spool 820 depending on the force relationship and communicating and cutting off the pressure regulating port 823 and the drain port 824, the line pressure is regulated by the first duty solenoid valve 91 to a value corresponding to the pilot pressure. It is now controlled by.

A speed ratio control valve 85 is provided in the line 102 . This gear ratio control valve 85 includes a spool 851 and a spring 8 that biases the spool 851 rightward in the figure.
52, a line pressure port 853 connected to the upstream part of the line 102, a drain boat 854, and a spring 85.
2 to the installation side and connected to the shift valve 8 via line 104.
Reverse port 855 connected to 7 and spring 8
5. It has a pilot chamber 856 formed on the opposite side to the installation side of 52 and into which pilot pressure is introduced. Pilot room 8
56 is connected via a pitot valve 86 to a second duty solenoid valve 92 and a pitot pressure generating means 90 that generates a pitot pressure corresponding to the rotational speed of the engine 1. Therefore, the pitot pressure generated by the pitot pressure generating means 90 and the pressure adjusted by the second duty solenoid valve 92 can be selectively introduced into the pilot chamber 856 as pilot pressure by the pitot valve 86. Even if the second duty solenoid valve 92 fails, the pitot pressure can be introduced from the pitot pressure generating means 90 into the pilot chamber 856 as a pilot pressure.

When the gear ratio control valve 85 is moving forward (when the shift valve 87 is in any of the shift positions 2.1)
Since hydraulic pressure is drained from the reverse port 855 via the shift valve 87, the spool 851 moves due to the force relationship between the pilot pressure introduced into the pilot chamber 856 and the biasing force of the spring 852, and the line The pressure port 853 and the drain boat 854 are selectively communicated with the hydraulic cylinder 414 of the primary pulley 41. In this way, when moving forward, the supply and discharge of hydraulic pressure to and from the hydraulic cylinder 414 of the primary pulley 41 is controlled in accordance with the pilot pressure introduced into the pilot chamber 856, so that the primary pulley 41 of the continuously variable transmission mechanism 4 It is configured to variably adjust the gear ratio between it and the secondary pulley 42.

On the other hand, when traveling in reverse (when the shift valve 87 is in the R shift position), hydraulic pressure (operating pressure to be described later) is introduced from the reverse port 855, and this operating pressure causes the spool 85 to
1 is fixed in a state where it is pressed to the right side in the figure. Therefore, when the vehicle moves backward, the hydraulic cylinder 414 of the primary pulley 41 and the drain boat 854 are constantly communicated with each other, and the gear ratio is fixedly maintained at the maximum gear ratio.

Furthermore, even in neutral and parking states (when the shift valve 87 is in the N and P shift positions), in which the driving force of the engine 7 is not transmitted to the axle 61 by the forward/reverse switching mechanism 3, the same state as in reverse is maintained. become.

The hydraulic oil whose pressure is regulated by the line pressure regulating valve 82 is
In addition to line 101, it is also sent out to line 105. The hydraulic fluid sent to line 105 is adjusted to a predetermined operating pressure by an operating pressure regulating valve 88, and then supplied to line 106 and line 1.07.

The operating pressure regulating valve 88 has a spool 881 and a spool 88.
a pilot chamber 882 formed on one end side of 1; a spring 883 interposed in this pilot chamber 882;
It has a first pressure regulation port 884 connected to the line 105, a second pressure regulation port 885 connected to the line 107, and a drain boat 886. The pilot chamber 882 is connected to the first duty solenoid valve 91 via the pilot passage 103a, and therefore, the hydraulic fluid whose pressure is regulated by the first duty solenoid valve 91 is supplied to the pilot chamber 882. It is now being introduced as pressure. While this pilot pressure acts in the same direction as the biasing force of the spring 883, the hydraulic pressure in the line 105 acts on the other end of the spool 881 to counteract the biasing force and the pilot pressure. The spool 881 moves depending on the relationship, and the first and second pressure regulating ports 884, 885 and the drain port 886 are communicated with and disconnected from each other, so that the operating pressure of the forward clutch 36 and the reverse brake 37 is adjusted to the first output. It is controlled to a value according to the pilot pressure regulated by the utility solenoid valve 91.

The hydraulic oil supplied to the line 106 is transferred to the shift valve 87
Gari, 2. When in the shift position 1, the forward clutch 3 of the forward/reverse switching mechanism 3 is connected via the line 109.
When the shift valve 87 is in the R shift position, it is supplied to the hydraulic chamber 37a of the reverse brake 37 of the forward/reverse switching mechanism 3 via the line 108, It is supplied to the reverse port 855 of the gear ratio control valve 85 via the.

On the other hand, when the shift valve 87 is in the R, N, or P shift position, the hydraulic fluid in the hydraulic chambers 36a, 37a of the forward clutch 36 and reverse brake 37 of the forward/reverse switching mechanism 3 flows through the lines 109, 108. It is designed to be discharged through the Therefore, the forward clutch 36 of the forward/reverse switching mechanism 3
The reverse brake 37 is engaged and released according to the shift position of the shift valve 87, and the continuously variable transmission mechanism 4 is engaged and released at the R, N, and P shift positions as described above.
The gear ratio is held fixed at the maximum gear ratio.

Further, the hydraulic oil supplied to the line 107 is supplied to the lockup engagement chamber 29a or lockup release chamber 29b of the lockup clutch 28 via the lockup control valve 89. The lock-up control valve 89 is configured such that the operation of the spool 891 is controlled by the pilot pressure regulated by the third duty solenoid valve 93. That is, when the pilot pressure becomes lower, the spool 891 moves to the right in the figure, and hydraulic oil is supplied from the line 107 to the lockup engagement chamber 29a, and at the same time, the lockup opening chamber 29b
The hydraulic oil inside is now drained, and the lock-up clutch 28 is engaged.

On the other hand, when the pilot pressure increases, the spool 891
moves to the left in the figure, hydraulic oil is supplied from the line 107 to the lock-up opening chamber 29b, the hydraulic oil in the lock-up engagement chamber 29a is drained, and the lock-up clutch 28 is disengaged. In this embodiment, the lock-up control valve 89 and the third
Engagement force control means 120 that controls the engagement force of the lock-up clutch 28 by the duty solenoid valve 93
is configured.

In addition, in the figure, 94 is the first duty solenoid valve 91.
Accumulator valve to prevent the pilot pressure in the pilot passage 103a from pulsating when becomes 0N10FF, 95.96 is one of the forward clutch 36 and reverse brake 37, respectively. Accumulator to alleviate the shock when engaged, 97 is a relief. It's a valve. Further, 98 is a pressure holding valve that maintains a predetermined low constant pressure when draining the pressure oil in the hydraulic cylinder 414 of the primary pulley 41.

FIG. 5 shows an electric control circuit for the above-mentioned continuously variable transmission. In this figure, a control unit 110 containing a microcomputer etc. receives a shift position signal from a shift position sensor 111 that detects a shift position (D, 1, 2 degrees R, N, P) operated by a driver. The primary pulley rotation speed signal from the primary rotation speed sensor 112 that detects the rotation speed Np of the primary shaft 4110, the secondary pulley rotation speed signal from the secondary rotation speed sensor 113 that detects the rotation speed NS of the secondary shaft 421, and the A throttle valve opening signal from a throttle opening sensor 114 that detects the throttle valve opening TvO, an engine rotational speed signal from an engine rotational speed sensor 115 that detects the rotational speed Ne of the engine 1, and a signal from the torque converter 2 A turbine rotation speed signal from a turbine rotation speed sensor 116 that detects the rotation speed Nt of the bottle shaft 25 is input.

The control unit 110 controls the first to third duty solenoid valves 9 based on these input signals.
1, 92, and 93 are duty-controlled, thereby adjusting each pilot pressure introduced into the line pressure regulating valve 82, the operating pressure regulating valve 88, the gear ratio control valve 85, and the lock-up control valve 89. .

The lock-up control procedure performed in the control units +-1 and 10 will be explained based on the control flow shown in FIG. Lock-up control is performed when the vehicle speed and the speed ratio of the torque converter 2 meet predetermined conditions. That is, as shown in FIG. 7, the secondary rotation speed N of the continuously variable transmission mechanism 4
Ss throttle opening TVo and target primary rotation speed N
pt, the upper limit line L1 of the secondary rotation speed NS and the upper and lower limit lines L2 . Lock-up control is executed in an operating region surrounded by L3, an inline L5 parallel to the 3-Low line L4, and an overdrive line L6. First, in step S1, lock-up torque LTQ is calculated. This lock-up torque LTQ becomes the engagement force of the lock-up clutch 28, and is calculated, for example, from a map set to increase as time passes. Then,
In step S2, the turbine torque TQTLIR applied to the turbine launch 23 is calculated based on the throttle opening degree Tvo and the engine speed Ne using the following equations (1) and (2).

In other words, engine speed Ne and throttle opening Tvo
The engine torque ENTQ, the engine speed Ne, and the lock-up torque LTQ, which are determined according to ), lock-up torque LTQ, and turbine torque TQTLJR ■
It is related to the formula.

4 ENT Q (Ne, Tvo)-IB
x2. vx (d, Ne /d t)−, c (
e) X, (Ne /1000) 2+LTQ
...■TQTUR-kc (e) X (Ne/1000) 2 Xk7 (e)+
LTQ... ■However, Ie
: Engine inertia NTUR : Turbine rotation speed Then, find the engine speed NO at which the above equation 0 holds, and substitute this engine speed Ne into equation (2) to find the turbine torque TQTUR.

Next, in step S3, the turbine calculated above]
It is determined whether the difference TQTUR-TQTUR(i-1) between the current value TQTUR and the previous value TQTUR(i-1) is less than or equal to a predetermined value. Here, [TQTUR-TQTUR(i
-1)≦predetermined value'', step S
Proceed to step 4, determine the final lockup torque LTQ as the lockup torque LTQ obtained in step S1 above, convert the lockup torque LTQ into a duty ratio, and apply the duty signal to the third duty solenoid valve. Output to 93. On the other hand, the determination in step S3 is rT
oruR-TQTUR(i-1)>predetermined value", the process proceeds to step S5, where the final lock-up torque LTQ is set lower than the lock-up torque LTQ obtained in step S1, and then the process proceeds to step S2. Return to ■ above based on the lockup torque LTQ.

(2) Calculate the turbine torque TQTUR again using the formula, and proceed to step S3.

In this embodiment, a turbine torque detection means 121 is configured to detect the torque TQ T U R of the turbine runner 23 in the torque converter 2 from step S2 in the above flow.

Also, step S3. Ss receives the output of the turbine torque detection means 121, and when the engagement force of the lock-up clutch 28 increases, when the amount of change in the turbine torque TQTUR increases beyond the set value 6, the engagement force of the lock-up clutch 28 increases. In order to suppress the increase, a fastening force suppressing means 122 is configured to control the third duty solenoid valve 93 and eight lock-up control valves (fastening force controlling means 120).

Therefore, in the above embodiment, when changing the gear ratio small, the second duty solenoid valve 9
2, the pilot pressure introduced into the pilot chamber 856 of the gear ratio control valve 85 increases and the spool 851 moves to the left in FIG. Line pressure flows into the hydraulic cylinder 414 of the primary pulley 41, and the amount of the inflow increases, so that the hydraulic pressure acting on the hydraulic cylinder 414 increases. Therefore, the effective pitch diameter of the primary pulley 41 becomes large, and the effective pitch diameter of the secondary pulley 42 becomes small. On the other hand, when changing the gear ratio to a larger value, contrary to the above, the spool 851 moves to the right in FIG. Therefore, the effective pitch diameter of the primary pulley 41 becomes smaller, and the secondary pulley 42
The effective pitch diameter becomes larger.

When the rotation speeds of the primary shaft 411 and the secondary shaft 421 in the continuously variable transmission mechanism 4 are in the lock-up region shown in FIG. 891 moves to the right in FIG. 4, hydraulic oil is supplied to the lockup engagement chamber 29a, the lockup clutch 28 is engaged, and the lockup piston 28a is engaged.
In other words, the turbine shaft 25 and the pump cover 21, that is, the engine output shaft 11 are directly connected.

In the engagement operation of the lock-up clutch 28, the engagement force of the lock-up clutch 28 is controlled by the operation of the third duty solenoid valve 93 and the lock-up control valve 89. Torque TQ of turbine runner 23
TUR is calculated at regular intervals in the control unit 110 by arithmetic operation 8, and this calculated turbine torque TO
The amount of change in TUR (difference from the previously calculated value) is compared with the set value. As shown in FIG. 8, when the amount of change in the turbine torque TQTUR is low, below the set value, the engagement force of the lock-up clutch 28 is controlled to increase as time passes, but the turbine torque TQTLJR When the amount of change in the lock-up clutch 28 increases more than the set value, the increase in the engagement force (lock-up torque LTQ) of the lock-up clutch 28 is suppressed, as shown by the broken line in FIG. As a result, when the lock-up clutch 28 is engaged, the turbine torque TQ T LI R always increases gradually with an amount of increase below a certain level, and its rapid increase is reliably avoided.
8. Shock when fastened can be effectively suppressed.

Incidentally, when the inventor of the present invention determined the change in the turbine torque in the torque converter when starting the vehicle with the throttle opening fully open, the results were as shown in FIG. 9(a). The broken line in the figure shows 9 conventional examples. Further, the change in the longitudinal acceleration of the vehicle when the same full development progresses is as shown in FIG. 4(b). From this figure, it is clear that according to the configuration of the present invention, the rotational fluctuation of the turbine torque is small when the lock-up clutch is engaged.
It turns out that the shock can also be reduced.

It goes without saying that the present invention can also be applied to fluid couplings other than torque converters.

(Effects of the Invention) As explained above, according to the invention according to claim (1), when the lock-up clutch of the fluid coupling is engaged, the actual fluctuation of the turbine torque is detected, and the fluctuation amount of the turbine torque is set. By suppressing the increase in the lock-up engagement force when it increases beyond the value, it is possible to effectively suppress the torque shock that accompanies the engagement of the lock-up clutch, and in particular eliminates the discomfort felt by the vehicle occupants. , comfort can be improved.

[Brief explanation of drawings]

FIG. 1 is a diagram showing the configuration of the present invention. Fig. 2 0 The following drawings show embodiments of the present invention, Fig. 2 is a skeleton diagram showing the overall configuration of the continuously variable transmission, Fig. 3 is a sectional view showing the detailed structure, and Fig. 4 is hydraulic control. Circuit diagram, Figure 5 is a block diagram of the electrical control system, Figure 6 is a flowchart for lock-up control in the control unit, and Figure 7 is a block diagram of the electrical control system.
The figure shows the characteristic diagram of the map showing the lock-up control operation area.
FIG. 8 is a characteristic diagram showing changes in lock-up torque when the lock-up clutch is engaged, and FIG. 9 is a characteristic diagram showing fluctuation characteristics of turbine torque and acceleration characteristics of the vehicle. 4]...Primary pulley (drive pulley) 42...
Secondary pulley (driven pulley) 43... Belt 411... Primary shaft (drive shaft) 421... Secondary shaft (driven wheel) 414.424... Hydraulic cylinder 120... Fastening force control means 121... - Turbine torque detection means 122... fastening force suppression means 1 JP-A-3 194263 (10) 12345 → Time (general) Fig. 9 (a) 2345 → 8 red 藺 (millet'!i') Fig. 9 (b) )

Claims (1)

[Claims] (1) In a fluid coupling in which the engagement force of a lock-up clutch that engages a pump and a turbine is controlled by a engagement force control means, there is provided a turbine torque detection means for detecting the torque of the turbine, and an output of the detection means. and controlling the engagement force control means to suppress an increase in the engagement force of the lockup clutch when the amount of change in turbine torque increases beyond a set value when the engagement force of the lockup clutch increases. A lock-up clutch engagement force control device for a fluid coupling, characterized in that a force suppressing means is provided.