JPH0289846A - Speed change control device for automatic transmission for vehicle - Google Patents

Abstract

PURPOSE:To improve driving stability through prevention of the occurrence of a slip by a method wherein when a friction factor detected by a friction factor detecting means is low, the engaging speed of a friction engaging device during gear shifting is lowered. CONSTITUTION:A speed change control device 100 detects the slip state on a running road surface of a tire from the number of revolutions of a front wheel detected by a number of revolutions of front wheel sensor 130 and the number of revolutions of a rear wheel detected by a number of revolutions of rear wheel sensor 140. From the slip state, the friction factor of the running road surface is estimated. A duty ratio of a pulse signal outputted to a duty solenoid valve 82 or 84 is changed so that the more the friction factor is decreased, i.e. the more a difference between the number of revolutions of a front wheel and the number of revolutions of a rear wheel is increased, the more a back pressure in a back pressure chamber 72a of an accumulator 72 or a back pressure chamber 76a of an accumulator 76 is decreased. This constitution prevents a tire from slipping on a running road surface even when a running road surface has a low friction factor, and enables ensurance of excellent driving stability.

Description

[Detailed Description of the Invention] [Industrial Application Field] The present invention relates to a speed change control device for an automatic transmission used in vehicles such as automobiles, and in particular, the present invention relates to a speed change control device for an automatic transmission used in vehicles such as automobiles, and in particular, the present invention relates to a speed change control device for an automatic transmission used in a vehicle such as an automobile. The present invention relates to a shift control device for a vehicle automatic transmission configured to perform gear switching.

[Prior Art] In automatic transmissions used in vehicles such as automobiles, a plurality of rotating elements of a planetary gear unit are brought into a predetermined combination by engaging and disengaging frictional engagement devices such as a plurality of clutches and brakes. In this type of automatic transmission, gears are changed by selective engagement of a frictional engagement device, such as switching between a plurality of gears by interconnecting or fixing the gears. is shown in, for example, Japanese Patent Laid-Open Nos. 57-134057 and 61-130653.

In the conventional automatic transmission for vehicles as described above, the engagement speed of the frictional engagement device during gear shifting, that is, the transmission torque capacity (
The rate of increase in clutch capacity ff1) or brake capacity changes depending on the output of the prime mover found from the throttle opening and the like, but does not change depending on other requirements.

[Problem to be solved by the invention] When a vehicle is running on a normal road surface, even if there is a relatively large variation in the torque transmitted to the tires during gear shifting, the tires will cause slip on the road surface. However, due to the low coefficient of friction of roads that are wet due to rainwater, etc. or frozen,
Torque fluctuations during gear shifting may also cause the tires to slip on the road surface.

The present invention provides a speed change for an automatic transmission for a vehicle that prevents the tires from slipping on the road surface even if the speed is changed while the vehicle is traveling on a road surface with a low coefficient of friction, thereby ensuring excellent handling stability. The purpose is to provide a control device.

[Means for Solving the Problems] According to the present invention, the above-mentioned object is to provide an automatic transmission for a vehicle, which is configured to change gears by selectively engaging a frictional engagement device. In the transmission control device, the friction coefficient detection means detects the friction coefficient of the road surface on which the vehicle travels, and when the friction coefficient detected by the friction coefficient detection means is low, the friction engagement device engages during gear shifting. This is achieved by a speed change control device characterized by having an engagement speed control means for reducing speed.

[Operations and Effects of the Invention] According to the configuration as described above, when the coefficient of friction of the road surface on which the vehicle travels is low, the engagement speed of the friction engagement device for gear shifting decreases;
By increasing the shift time, the rate of torque fluctuation during gear changes becomes smaller, and the torque transmitted to the tires during gear changes does not change suddenly. The vehicle will no longer slip on the road surface, ensuring excellent steering stability.

[Example] The present invention will be described in detail below with reference to the accompanying drawings.

FIG. 1 shows an example of a Jffl gear type transmission of a vehicle automatic transmission to which a hydraulic control device according to the present invention is applied. In the figure, 10 is a first sun gear, 12 is a first ring gear concentric with the first sun gear 10, 14 is a first planetary pinion that meshes with the first sun gear 10 and the first ring gear 12, and 16 is a first planetary pinion. 20 is a second sun gear; 22 is a second ring gear concentric with the second sun gear 20; 24 is meshed with the second sun gear 20 and the second ring gear 22; 26 indicates a second carrier rotatably supporting the second planetary pinion 24. The first ring gear 12 is connected to the second carrier 26 by a connecting element 30, and the first carrier 16 is connected to the second ring gear 22 by a connecting element 32.

Note that the simple planetary gear mechanism constituted by the first sun gear 10, the first ring gear 12, the first planetary pinion 14, and the first carrier 16 is referred to as the first row planetary gear mechanism, and the second sun gear 20 A simple planetary gear mechanism constituted by the second ring gear 22, the second planetary pinion 24, and the second carrier 26 is referred to as a second row planetary gear mechanism.

A first carrier 16 and a second ring gear 22 connected to the first carrier 16 by a connecting element 32 and a housing 5
0, a first one-way clutch 34 and a second one-way clutch 36 are provided in series with each other. In this case, the first one-way clutch 34 is connected to the first carrier 1
6, and a second one-way clutch 36 is provided on the housing 50 side. More specifically, the first one-way clutch 34 is connected to the first carrier 16 through its inner race 34a, and its outer race 34b is connected to the inner race 36a of the second one-way clutch 36 through the connecting member 31. Outer race 36b is connected to housing 50.

The second carrier 26 is connected to the output gear 54 so that it always acts as an output member.

The first one-way clutch 34 is in an engaged state when the outer race 34b is about to rotate beyond the rotational speed of the inner race 34a during engine drive, and is in a slipping state when it is the opposite, and the second one-way clutch 36 is in an engaged state when the inner race 36a is reversed with respect to the outer race 36b during engine drive, and is in a sliding state when the inner race 36a is reversed to the outer race 36b.

A first clutch 38 is provided between the second sun gear 20 and the input shaft 52 to selectively connect the two to each other.

A second clutch 40 is provided between the first carrier 16 and the input shaft 52 to selectively connect the two to each other.

A third clutch 42 is provided between the first sun gear 10 and the input shaft 52 to selectively connect the two to each other.

A fourth clutch 44 is provided between the first sun gear 10 and the connecting member 31 to selectively connect the two to each other.

A connecting member 31 is provided between the connecting member 31 and the housing 50.
A first brake 46 is provided for selectively securing the housing 50 to the housing 50.

A second brake 48 is provided between the second ring gear 22 and the housing 50 to selectively fix the second ring gear 22 to the housing 50.

The manner in which the first gear, second gear, third gear (directly coupled gear), fourth gear (increasing gear), and reverse gear are achieved by the planetary gear type transmission configured as described above is as follows. As shown in Table 1. In this table, a 0 mark indicates that the relevant clutch, brake, or one-way clutch is engaged in the engine drive state, and (0) indicates that the relevant clutch or brake is engaged, the gear shift is This indicates that engine braking can be applied during this stage.

If the ratio of the number of teeth of the first sun gear 10 to the number of teeth of the first ring gear 12 is ρ1, and the ratio of the number of teeth of the second sun gear 20 to the number of teeth of the second ring gear 22 is ρ, then the shift of each gear stage The ratios are shown in Table 2.

Second gear stage ((1+ρ:)/ρ:)-+1/ FIG. 2 shows an embodiment of the speed change control device according to the present invention. In FIG. 2, 46a indicates a hydraulic servo chamber of the first brake 46, and 48a indicates a hydraulic servo chamber of the second brake 48. When hydraulic pressure is supplied to the hydraulic servo chamber 46a, the first brake 46 is engaged, and the hydraulic servo chamber 48a
The second brake 48 is engaged when hydraulic pressure is supplied to the second brake 48 .

Line hydraulic pressure is selectively supplied to the hydraulic servo chamber 46a from the oil passage 62 by switching the shift valve 60. The shift valve 60 is operated to switch by opening and closing the solenoid valve 64, and when the solenoid valve 64 is in the on state, the shift valve 60 is switched to the line hydraulic pressure supply position as shown in the figure in response to the opening of the solenoid valve 64; When 64 is in the OFF state, it is switched to the oil drain position in response to the valve closing.

Therefore, when the solenoid valve 60 is in the on state, the first brake 46 is engaged by supplying line hydraulic pressure to the hydraulic servo chamber 40a thereof, whereas when the solenoid valve 64 is in the off state, the first brake 46 is engaged. The hydraulic pressure of 46a is discharged and the state is released.

Line hydraulic pressure is selectively supplied to the hydraulic servo chamber 48a from the oil passage 62 by switching the shift valve 66. The shift valve 66 is switched by opening and closing the solenoid valve 68, and when the solenoid valve 68 is in the on state, the shift valve 66 is switched to the line oil pressure supply position as shown in the figure when the solenoid valve 68 is opened. When the valve is in the OFF state, it is switched to the oil drain position in response to the valve closing.

Therefore, when the solenoid valve 68 is in the on state, the second brake 48 is engaged by supplying the line hydraulic pressure to the hydraulic servo chamber 48a, whereas when the solenoid valve 68 is in the off state, the second brake 48 is engaged by the line hydraulic pressure in the hydraulic servo chamber 48a. It is expelled and becomes liberated.

A hydraulic delay circuit 70 and an accumulator 72 are provided in the middle of the oil path leading from the shift valve 60 to the hydraulic servo chamber 46a. Also, from the shift valve 66, the hydraulic servo chamber 48a
A hydraulic delay circuit 74 and an accumulator 76 are provided in the middle of the oil path leading to.

The accumulators 72 and 76 have back pressure chambers 72a and 76, respectively.
The accumulator operates with characteristics according to the oil pressure applied to the brake 6a, and the brake capacities of the first brake 46 and the second brake 48 are each optimized during gear shifting.

Hydraulic pressure regulated by an accumulator control valve 78 is supplied to the back pressure chamber 72a of the accumulator 72 for the first brake 46, and the back pressure chamber 72a of the accumulator 76 for the second brake 48 is supplied with hydraulic pressure regulated by an accumulator control valve 78.
Hydraulic pressure regulated by an accumulator control valve 80 is supplied to a.

The pressure regulation value of the accumulator control valve 78 is determined by the pressure regulation value by the duty solenoid valve 82, and this pressure regulation value is determined by the duty ratio of the pulse signal given to the duty solenoid valve 82.

The pressure regulation value of the other accumulator control valve 80 is determined by the pressure regulation value by the duty solenoid valve 84, and the pressure regulation value of the duty solenoid valve 84 is determined by the duty ratio of the pulse signal given to it. .

Note that dampers 86 and 88 are installed to absorb hydraulic pulsations caused by the repeated opening and closing of duty solenoid valves 82 and 84.
and is provided. Further, in order to prevent the oil pressure control state by the duty ratio control of the duty solenoid valves 82 and 84 from changing due to a change in the line oil pressure, a modulating valve 9 is provided for these duty solenoid valves 82 and 84.
0, a constant oil pressure is supplied.

According to the above configuration, the duty solenoid valve 82
By controlling the duty ratio of the pulse signal given to each of the accumulators 72 and 84,
The back pressure of each of the accumulators 6 is controlled, and the brake capacities of the first brake 46 and the second brake 48 during gear shifting can be arbitrarily set accordingly.

Solenoid valve 64 for switching shift valves 60 and 66
and 68 and the duty ratio control of the duty solenoid valves 82 and 84 for back pressure control of the accumulators 72 and 76 are performed by electrical signals from an electronic control device 100 including a microcomputer of a general structure. It is becoming more and more popular.

The electronic control device 100 receives information regarding the vehicle speed from the vehicle speed sensor 110, information regarding the throttle opening of the internal combustion engine from the throttle opening sensor 120, and information regarding the throttle opening of the internal combustion engine from the front wheel rotation speed sensor 1.
30 and the rear wheel rotation speed sensor 140, respectively, and the speed change control and the back pressure side iH of the accumulator are performed in accordance with these information.

In the transmission control device according to the present invention, the slip situation of the tires on the road surface can be detected from the difference between the front wheel rotation speed detected by the front wheel rotation speed sensor 130 and the rear wheel rotation speed detected by the rear wheel rotation speed sensor 140. The friction coefficient of the running road surface is estimated from this, and when the friction coefficient is low, that is, when the difference between the front wheel rotation speed and the rear wheel rotation speed is large, the back pressure chamber 72a of the accumulator 72 or the accumulator 76 is detected.
The duty ratio of the pulse signal output to the duty solenoid valve 82 or 84 is changed so that the back pressure in the back pressure chamber 76a is reduced.

The first brake 46 is engaged during a downshift to the first gear that requires engine braking, such as in the L range, and the second brake 48 is engaged during a downshift to the first gear that requires engine braking, such as in the S range. It is engaged when downshifting to the next step, and the hydraulic servo chamber 4 of these brakes
Back pressure chambers 72a, 76 of accumulators 72, 76 that control the rising speed of the servo oil pressure applied to 6a, 48a
The pressure a becomes lower when the vehicle is traveling on a road with a low friction coefficient due to the duty ratio control as described above, and accordingly, the rise in the servo oil pressure in the oil pressure servo chamber 46a or 48a becomes slower. This causes the brake 46 or 48 to
The engagement speed of the brakes decreases, and the capacity of these brakes gradually increases.

As a result, during gear shifting, the torque applied to the tires changes gradually without sudden changes, and the tires do not slip on the road surface even if the friction coefficient of the road surface is low.

In addition, in the above-mentioned embodiment, the accumulator control valve 7
8 and 80 are duty solenoid valves 82 and 84, respectively.
The pressure adjustment value is controlled by the oil pressure from the duty solenoid valve, but even though the pressure adjustment is performed according to the oil pressure according to the throttle opening in addition to the oil pressure from the duty solenoid valve. good. The back pressure characteristics of the accumulator in this case are as shown in FIG.

The front wheel rotation speed sensor 130 and the rear wheel rotation speed sensor 140 may be used in an anti-skid control system or an anti-lock brake system in a vehicle braking system.

Furthermore, in the above-described embodiment, the friction coefficient of the running road surface is detected based on the difference between the front wheel rotation speed and the rear wheel rotation speed, but this can be calculated from, for example, the drive wheel rotation speed. It may be detected based on the wheel acceleration determined by
Further, it may be carried out using other appropriate means, for example, a Doppler speedometer.

Furthermore, in the above embodiment, the engagement speed of the frictional engagement device is controlled by back pressure control of the accumulator, but this is due to the servo oil pressure itself applied to the hydraulic servo chamber of the frictional engagement device. The control may be performed under the control of

[Brief explanation of drawings]

FIG. 1 is a skeleton diagram showing an example of a planetary gear type transmission of a vehicle automatic transmission to which a hydraulic control device according to the present invention is applied, and FIG. FIG. 3 is a graph showing the back pressure control characteristics of the accumulator in the transmission control device according to the present invention. 10...first sangya, 12...first ring gear. 14... First planetary pinion, 16... First carrier, 20... Second sun gear, 22... Second ring gear, 24... Second planetary pinion, 26...
・Second carrier, 30.31.32...Connection element, 3
4...First one-way clutch, 36...Second one-way clutch, 38...First clutch, 40...
・Second clutch, 42...Third clutch, 44...
Fourth clutch, 46...first brake, 48...second brake. 50...Housing, 52...Output shaft, 54...
Output gear, 60...Shift valve, 62...Oil passage, 64
...Solenoid valve, 66...Shift valve, 68-...
Solenoid valve. 70... Hydraulic delay valve, 72... Accumulator,
74...Hydraulic delay valve, 76...Accumulator,
78.80...Accumulator control valve, 82.84
...Duty solenoid valve, 86.8g...Damper, 90...Modulating valve

Claims (1)

[Claims] Friction coefficient detection for detecting the friction coefficient of the road surface on which the vehicle is running, in a shift control device for an automatic transmission for a vehicle that is configured to switch gears by selectively engaging a friction engagement device. and engagement speed control means for reducing the engagement speed of the frictional engagement device during shifting when the friction coefficient detected by the friction coefficient detection means is low. Control device.