JPH0475421B2 - - Google Patents

Description

[Detailed description of the invention]

The present invention relates to a shift hydraulic pressure control method in an automatic transmission for a vehicle, which supplies appropriate hydraulic pressure to frictional engagement elements during gear shifting to always perform gear shifting in a good condition. Automatic transmissions for vehicles supply hydraulic pressure to frictional engagement elements such as clutches and brakes, and
By selecting rotational elements such as gears, gear ratios (shifts) are automatically performed according to the driving conditions of the vehicle, and this frictional engagement is used to protect devices and equipment and maintain a comfortable ride. Delivery of pressure oil to the elements must be done gradually according to certain predetermined characteristics. In other words, if the feed oil pressure is suddenly increased, the frictional engagement elements are suddenly connected and the shock during gear shifting becomes large, which not only places an excessive load on the automatic transmission and engine, but also causes damage to the vehicle. It was also making me feel worse. On the other hand, if the feed oil pressure rises too slowly, the shift time becomes longer and the frictional engagement element slips excessively, shortening the life of the frictional engagement element or preventing the gearshift from being carried out. There was a fear that it would not happen. The present invention has been developed in view of the above circumstances, and uses duty control to achieve proper hydraulic pressure delivery to the frictional engagement elements during gear shifting, thereby preventing excessive gearshift impact and slippage of the frictional engagement elements. The purpose is to achieve good shifting without any problems. A shift hydraulic control method for an automatic transmission for a vehicle according to the present invention that achieves the above object includes duty control of an input shaft into which rotational power of an engine is input, an output shaft that outputs rotational power to drive wheels, and a solenoid valve. A vehicular automatic transmission comprising: a frictional engagement element that operates with pressure oil whose pressure is regulated by a friction engagement element that switches a gear ratio between the input shaft and the output shaft by selecting an arbitrary rotating element; The rotational speed of the rotating element whose rotational speed changes during the gearshift and the rotational speed at the end of the gearshift are set as the target rotational speed at the start of the gearshift, and the rotational speed of the rotary element that is continuously detected from the start of the gearshift and the A duty rate is set every time according to the difference from the target rotation speed, and this duty rate is corrected by a duty rate correction factor corresponding to the rotation change rate of the rotating element. The present invention is characterized in that the oil pressure supplied to the frictional engagement element is controlled by driving a valve so that the rotational speed of the rotary element reaches the target rotational speed at the end of the shift. Hereinafter, an embodiment of the present invention will be described using a shift stage from 1st speed to 2nd speed as an example. First, a Ravigneaux-type automatic transmission with four forward speeds and one reverse speed in which the method of the present invention has been implemented will be explained with reference to FIG. 1, which schematically shows its structure. It is directly connected to the pump 8 of the converter 6. The torque converter 6 includes a pump 8, a turbine 10, a stator 12, and a one-way clutch 1.
4, the stator 12 has a one-way clutch 1
The one-way clutch rotates the stator 12 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 to the input shaft 20
The signal is transmitted to a gear transmission 22 disposed at the rear thereof, which achieves four forward speeds and one reverse speed. The transmission 22 includes three sets of clutches 24, 2
6, 28, two sets of brakes 30, 32, one set of one-way clutch 34, and one set of Ravigneau type planetary gear mechanism 36. The planetary gear mechanism 36 includes a ring gear 38, a long pinion gear 40, a side pinion gear 42, a front sun gear 44, a rear sun gear 46, and both pinion gears 4.
The carrier 48 rotatably supports the carriers 0 and 42 and is also rotatable, and the ring gear 38
is connected to the output shaft 50, and the front sun gear 44
Kickdown drum 52, front clutch 2
4, a rear sun gear 46 is connected to the input shaft 20 via a rear clutch 26, and a carrier 48 connects a low reverse brake 32 and a one-way clutch 34, which are arranged functionally in parallel. The input shaft 20 is connected to the case 16 through the transmission 22, and to the input shaft 20 through a four-speed clutch 28 disposed at the rear end of the transmission 22. The kick-down drum 52 can be fixedly connected to the case 16 by means of a kick-down brake 30. The torque passing through the planetary gear mechanism 36 is transmitted from an output gear 60 fixed to an output shaft 50 to a driven gear 64 via an idle gear 62.
The signal is further transmitted to a differential gear device 72 connected to a drive shaft 70 of the drive wheels via a transfer shaft 66 fixed to the driven gear 64 and a helical gear 68. Each of the above-mentioned clutches and brakes, which are frictional engagement elements, is composed of a frictional engagement device equipped with an engaging piston device or a servo device, etc., and is driven by the engine 2 by being linked to the pump 8 of the torque converter 6. It is operated by hydraulic pressure generated by an oil pump (not shown). The hydraulic pressure is selectively supplied to each clutch and brake according to the operating state detected by various operating state detection devices by an electronic hydraulic control device controlled by a computer, which will be described later. As shown in Table 1, the combination of the following operations achieves four forward speeds and one reverse speed. In the same table, the ◯ mark indicates the engagement state of each clutch or brake, and the 〓 mark indicates that the carrier 48 is engaged by the action of the one-way clutch 34 immediately before the low reverse brake 32 is engaged during gear shifting.
indicates that the rotation is stopped. Incidentally, a computer-based electro-hydraulic control device for achieving the gears shown in Table 1 in the gear transmission 22 described above is disclosed in Japanese Patent Application No. 56-144237 (Japanese Unexamined Patent Publication No. 144237).
No. 58-46258) etc., so the explanation will be omitted.

[Table] In this embodiment, as can be seen from Table 1, the kick down brake 30 is set to 1st gear - 2nd gear.
As the gear changes from speed (2nd) to third speed (3rd) to fourth speed (4th), the kick-down drum 52 becomes in the rotation-stop-rotation-stop state. This kick-down drum 52 is set as a rotating element that serves as a reference for gear shift hydraulic control.
The rotational speed of the engine is detected by a sensor 142 and input into a computer 80 shown in FIG. FIG. 2 is a drawing showing a hydraulic control device for the kickdown brake 30, which is duty-controlled by a computer 80 provided in the electrohydraulic control device of the gear transmission 22. The hydraulic pressure sent to the down servo piston 30a is adjusted.
This control device includes a reducing valve 81, a solenoid valve 82, and a pressure control valve 83. The reducing valve 81 uses the oil pressure generated by the oil pump (not shown) as a line pressure to a predetermined constant pressure, and further regulates the line pressure to a certain constant pressure to make a reducing pressure, and converts this oil pressure into a reducing pressure through a bifurcated oil path 84. The pressure is supplied to both ends of a tandem spool valve 83a provided in the pressure control valve 83. The solenoid valve 82 is connected to one orifice 8 of the oil passage 84.
5, and a computer 8
It is of a non-energized closed type whose duty is controlled by 0, and adjusts the reducing pressure supplied to the spool valve 83a. The pressure control valve 83 regulates the line pressure by operating the spool valve 83a and controls the kickdown servo piston 3.
0a, and controls the engagement and disengagement of the kickdown band 30b wound around the kickdown drum 52. Here, the duty control is solenoid valve 8.
When one period of the constant frequency pulse current sent to the kick-down servo piston 30a is defined as a, and the energization time is defined as b, this is carried out according to the duty rate expressed as b/a x 100 (%). There is a relationship as shown in FIG. 3 between the supplied hydraulic pressure and the duty rate. The characteristics of duty rate and duty rate change rate as shown in FIG. It will be done. Figure 4a shows the kick-down drum 5 at the end of the gear shift.
Difference between the rotation speed (target rotation speed) Ns of No. 2 and the rotation speed Nd of the kickdown drum 52 during gear shifting N=Nd−
It shows the duty rate V ₁ with respect to Ns, and as shown in Figure 4c, the kickdown drum 5
In the shift stage from the first speed to the second speed where the rotational speed Nd of the engine 2 decreases, _V1 has a negative value and has the characteristic that the rotational speed Nc is the inflection point. Note that the rotational speed Nc is set to, for example, about 1000 rpm, and at higher rotational speeds, _V1 is kept constant to prevent engagement shock of the kick-down brake 30 due to a sudden rate of increase in oil pressure. Fig. 4b shows the differential value of the rotational speed Nd shown in Fig. 4c, that is, the duty rate change rate V2 with respect to the rotational speed change rate N・d of the kickdown drum 52, and this characteristic straight line shows the differential value of the rotational speed Nd shown in _{Fig. 4c} . The slope β is a function of the engine throttle opening degree θt. Shift hydraulic control from the first gear to the second gear is carried out by the computer 80 according to the flowchart shown in FIG. 5 under duty control based on the characteristics shown in FIGS. 4a and 4b. That is, when a shift start signal from 1st speed to 2nd speed is issued and a shift solenoid valve (not shown) of an electro-hydraulic control device controlled by computer 80 is switched, the kick-down drum 52
The target rotation speed Ns is set by the computer 80. This target rotational speed Ns is Orpm because the kick-down drum 52 stops at the end of the shift when changing from 1st to 2nd speed and from 3rd to 4th speed, but when changing from 2nd to 3rd speed, the target rotation speed Ns is Orpm. The value corresponds to the vehicle speed at that time. Then, the engine throttle opening θt and the kick-down drum rotation speed Nd are detected by the detector and inputted to the computer 80, and based on the difference N between this rotation speed Nd and the target rotation speed Ns, the characteristics shown in FIG. 4a are obtained. A duty rate V ₁ is provisionally determined. At the bundle checking stage from 1st speed to 2nd speed, hydraulic pressure determined by the duty rate _V1 , which is a negative value, is sent to the kickdown servo piston 30a, and at the same time the rotational speed of the kickdown drum 52 begins to decrease, this kickdown Drum rotation speed
Nd is detected, the rotation speed change rate N・d is calculated, and the fourth rotation speed change rate N・d and the throttle opening θt are calculated.
Determine the duty rate change rate V ₂ from the characteristics shown in Figure b. In addition, in the gear stage from 1st gear to 2nd gear, as shown in Figure 4 d, the rotation speed change rate N・d is a negative value, so from Figure 4 b, the duty rate change rate V ₂ is a positive value. Become. Then, this duty rate change rate V ₂ is added to the duty rate V ₁ to obtain a corrected duty rate V ₃
is calculated, and the oil pressure supplied to the kickdown servo piston 30a is controlled by this corrected duty rate _V3 . In this way, the above-described process is repeated every calculation cycle of the computer 80 based on the rotational speed of the kick-down drum 52 that is continuously detected from the start of the shift. Note that the corrected duty rate V ₃ is determined by the duty rate V ₁ and the duty rate change rate depending on the characteristics of the vehicle.
It is also possible to obtain it from V ₂ using other arithmetic expressions. That is, in this way, the corrected duty rate V ₃
By controlling with
V ₂ increases, and the rate of increase in the oil pressure sent to the kick-down servo piston 30a is generally slowed down to prevent a shift shock caused by sudden engagement, and conversely, the rate of decrease in the rotational speed Nd is reduced. If it is dull and the absolute value of the rotational speed change rate N・d is small, the duty rate change rate V ₂ decreases, and the rise rate of the oil pressure sent to the kickdown servo piston 30a is generally increased to increase the frictional engagement element. prevent excessive slippage.
Therefore, the oil pressure supplied to the frictional engagement element that cooperates with the kick-down brake 30 at each gear stage becomes appropriate, and the gear shifting operation of the automatic transmission proceeds smoothly. Then, it is determined whether or not the rotational speed difference N has become zero, and if the rotational speed difference N is not zero, the duty control of the oil pressure supplied to the kick-down servo piston 30a as described above is changed to the rotational speed difference N. This process is repeated until the end of the shift when N becomes zero. Furthermore, as the speed change approaches the end, the rotational speed Nd decreases, so the duty ratio _V1 decreases as shown in FIG. reduced. In addition, near the end of the shift, the increase in the oil pressure supplied to the kick-down servo piston 30a is small, so it is possible that the gear will jump up again to 1st gear due to insufficient engagement force, but the rotation speed Nd of the kick-down drum 52 As the rotational speed change rate N·d increases to a positive value, the duty rate _V1 and the duty rate change rate _V2 change, increasing the degree of increase in the oil pressure supplied to the kick-down servo piston 30a. Prevents speed up to 1st gear. As a result of such duty control, the hydraulic pressure of the kick-down servo piston 30a increases, as shown by the dotted line in FIG. 6, compared to the conventional system not using the above method, which is shown by the solid line in the same figure. The rotational speed of the kick-down drum 52 is reduced smoothly, and rapid fluctuations in the output shaft torque (marked with a circle in the figure) are suppressed. In the above embodiment, the gear stage from 1st gear to 2nd gear was explained, but the gear stage from 2nd gear to 3rd gear and from 3rd gear to 4th gear can also be carried out in the same manner as above. In addition, the Lavigneaux type has 3 forward speeds and 1 reverse speed.
The same applies when applied to an automatic transmission. In addition, in the above embodiment, since the rotational element to be detected is a kick-down drum that switches between a rotating state and a stopped state at all gears, the rotation speed detection device is simple; A rotating element that switches between a stopped state and a stopped state, such as rear sun gear 4
6. It is also possible to use a carrier 48 or the like as appropriate. The method of the present invention is also effective when applied to the initial oil pressure setting method for automatic transmissions for vehicles, which has already been proposed in Japanese Patent Application No. 59-69926. In other words, the line pressure of the automatic transmission that is combined with the engine is determined by the displacement, output torque, etc., so if the specifications of the engine and automatic transmission do not match, frictional engagement may occur at the beginning of gear shifting. There is a problem in that excessive or insufficient hydraulic pressure supplied to the elements may cause a large shift impact or a prolonged shift completion time. However, according to the method of the present invention, even if the specifications of the engine and the automatic transmission do not match, it is possible to deliver an appropriate hydraulic pressure to the frictional engagement element to achieve a good gear shift, and it is possible to achieve a good gear shift by supplying the appropriate hydraulic pressure to the frictional engagement element. It is possible to adapt the automatic transmission to engines of different standards in a shorter time than in the above-mentioned existing proposal, which calculates and sets the initial shift oil pressure based on the oil pressure of the automatic transmission. As explained above, according to the present invention, in order to achieve a good shift, the rotation speed at the end of the shift of the rotating element whose rotation speed changes during the shift is set as the target rotation speed at the start of the shift, and the rotation speed at the end of the shift is set at the start of the shift. The duty rate is set each time according to the difference between the rotational speed of the rotating element that is continuously detected from time to time and the target rotational speed, and this duty rate is corrected with a duty rate correction factor that corresponds to the rotational change rate of the rotating element. However, since the solenoid valve is driven by this corrected duty rate to control the hydraulic pressure supplied to the frictional engagement element, it is possible to prevent excessive shift impact or excessive slippage of the frictional engagement element. It is possible to achieve good gear shifting.

[Brief explanation of the drawing]

FIG. 1 is a schematic configuration diagram of a vehicle automatic transmission to which an embodiment of the present invention is applied, and FIG. 2 is a schematic configuration diagram showing an example of a hydraulic control device for a kick-down brake.
Figure 3 is a graph showing the relationship between the duty rate and oil pressure, Figure 4a is a graph showing the relationship between the rotational speed of the kickdown drum and the duty rate, and Figure 4b is a graph showing the relationship between the rotational speed of the kickdown drum and the rotational speed. Graphs showing the relationship with the duty rate change rate, Figures 4c and d are graphs showing changes in the rotational speed and rotational speed change rate of the kick-down drum during gear shifting, respectively, and Figure 5 is a graph showing the change in rotational speed of the kick-down drum during gear shifting. The flowchart, FIG. 6, is a graph showing the relationship among the oil pressure supplied to the kick-down brake, the output shaft torque, and the rotational speed of the kick-down drum during gear shifting. In the drawing, 2 is the engine, 20 is the input shaft, 30 is the kick-down brake, 50 is the output shaft, 52 is the kick-down drum, Ns is the target rotation speed, V ₁ is the duty rate, V ₂ is the duty rate change rate, V ₃ is the corrected duty rate.

Claims (1)

[Claims] 1. An input shaft into which the rotational power of the engine is input;
An output shaft that outputs rotational power to the drive wheels, and a gear ratio between the input shaft and the output shaft by operating with pressure oil whose pressure is regulated by duty control of a solenoid valve and selecting an arbitrary rotating element. In a vehicle automatic transmission equipped with a frictional engagement element that changes the rotation speed during a shift, the rotation speed at the end of the shift of the rotating element that changes during the shift is set as the target rotation speed at the start of the shift, and from the start of the shift A duty rate is set every time according to the difference between the continuously detected rotational speed of the rotary element and the target rotational speed, and this duty rate is set by a duty rate correction factor corresponding to the rotational change rate of the rotary element. the solenoid valve is driven by the corrected duty rate to control the hydraulic pressure supplied to the frictional engagement element, and the rotational speed of the rotary element reaches the target rotational speed at the end of the shift. 1. A method for controlling shift hydraulic pressure of an automatic transmission for a vehicle, characterized in that: