JPS62255645A - Hydraulic controller for automatic transmission - Google Patents

Abstract

PURPOSE:To reduce the shock on speed change by means of freely reducing the back pressure of an accumulator by adjusting and controlling the line oil pressure and the back pressure for determining the characteristics of an accumulator by respectively independent control means. CONSTITUTION:Electromagnetic proportional valves SA and SL are controlled by a computer 84 into which the signals of a car speed sensor 82, etc. are input. A plurality of electromagnetic proportional valves SA control the back pressure of an accumulator 107 installed into an oil passage communicating to each frictional engaging device, and the electromagnetic proportional valves SL controls the line oil pressure in a hydraulic controller 86 through a primary regulator valve 103. The discharge pressure of a pump 112 acts onto the port 114L of the electromagnetic proportional valve SL, and a line oil pressure PL corresponding to the control oil pressure Ptheta is generated in the valve 103. Therefore, the line oil pressure can be controlled arbitrarily by controlling the load electric current IL supplied into a coil 108L by the instruction of the computer 84.

Description

[Detailed description of the invention]

3.Ay’l111 explanation of the invention

[Industrial Application Field 1] The present invention relates to a hydraulic control device for an automatic transmission, and in particular, a friction engagement device is switched by a line hydraulic pressure regulated by the t51 hydraulic pressure of an oil pump, and a gear stage of an automatic transmission l is switched. The present invention also relates to an improvement in a hydraulic control device for an automatic transmission in which an increase in hydraulic pressure is temporarily filled-DO by an accumulator when the line hydraulic pressure 8 is introduced to the frictional engagement device. [Prior Art] A gear shift BJ structure and a plurality of frictional engagement devices ((keep in mind,
An automatic transmission system for a vehicle 2 that is configured to selectively switch the engagement of the frictional engagement device by operating a hydraulic control device to achieve any one of a plurality of gears! The machine is already widely known. Said l! l! A friction engagement device generally consists of two sets of friction plate elements that are rotatably supported relative to each other and a hydraulic servo device that drives the friction plate elements, and when hydraulic pressure is supplied to the hydraulic servo device, , the two sets of friction plate elements are strongly pressed against each other and are coupled in a relationship that allows 1-lux transmission between them. The working oil pressure for this frictional engagement device is generally called line pressure or line oil pressure. Conventionally, this line oil pressure is usually changed in accordance with a value considered to be representative of the engine load, such as the throttle opening of the engine, and control is performed such that the line oil pressure becomes higher as the engine load becomes larger. Control of this line oil pressure has traditionally been carried out by introducing a throttle oil pressure that changes depending on the throttle pressure into the V control boat of the briny pressure regulator for controlling the line ID+pressure. This throttle oil pressure has generally been generated by a throttle valve that exerts an increasing spring force on its spool in response to depression of the accelerator pedal. In recent years, electronic automatic gearboxes have come to be operated with Di1 transmission, and the main part of the control circuit has come to be constituted by an electronic circuit. Since information regarding the throttle opening is also taken in the form of an electrical signal, a device has also been developed in which the line oil pressure is controlled based on the electrical signal regarding the throttle opening of 1371 degrees (for example, 56-1255
55>. By the way, this line oil pressure is 1! ? When introducing oil pressure to a frictional engagement device, a method is generally adopted in which the increase in oil pressure is temporarily controlled using a 7-cumulator. This accumulator is a type of oil reservoir installed in the line hydraulic oil path to the friction engagement device, and by accumulating oil directed toward the friction engagement device in this oil reservoir, the friction engagement is maintained for a predetermined period. This is to suppress the increase in hydraulic pressure required for alignment. This characteristic at the time of suppression can be changed by controlling the oil pressure (back pressure) applied to the back pressure chamber of the accumulator. Conventionally, this back pressure
The line oil pressure, or the oil pressure adjusted from this line oil pressure, was used, and therefore, the back pressure of the accumulator was determined primarily by the line oil pressure (Japanese Patent Publication No. 1973- 8025>.

Problem to be Solved by the Invention 1 However, if the back pressure of the accumulator is determined depending on the line oil pressure in this way, there is sufficient freedom in setting the oil pressure to the frictional engagement position 1a during gear shifting. There was a problem that it was sometimes not possible to secure the required amount. To be more specific, for example, at power-on upshift, J
It is desirable to control the oil pressure so that it decreases slightly near the end of engagement of the fi friction engagement device. In order to realize this control, if the back pressure of the accumulator depends on the line oil pressure as in the past, the line oil pressure must be lowered. However, since the line oil pressure is determined depending on the throttle opening, it is not possible to freely control the line oil pressure to decrease when changing gears. Furthermore, even if a control means fSQ that can control the line oil pressure in 1st gear is provided, the line pressure is excessively high enough to cause slippage in the frictional engagement device that is already in the engaged state during a particular gear shift. It is not possible to perform any control that would reduce the oil pressure. Generally, for example, during power-on upshift, the engine speed is 3!
! Since the torque decreases, an inertia torque occurs, and therefore, the shared torque of the frictional engagement device, which is already in an engaged state, increases as if the engine torque had increased by the amount of the inertia torque. If this is the case, controlling the line oil pressure to a low level in order to control the acting force of the PR engagement device that is about to be engaged should be suppressed depending on the case. [Object of the Invention] The present invention has been made in view of such conventional problems, and it is an object of the present invention to control the hydraulic pressure to a frictional engagement device that is about to be engaged to a sufficiently low level as intended. An object of the present invention is to provide a hydraulic control device for an automatic transmission that can control υj so that slippage does not occur in a frictional engagement device that is already in an engaged state at that time. [Means for Solving the Problems] The present invention switches the engagement state of a friction engagement device using a line hydraulic pressure regulated from the base hydraulic pressure of an oil pump, and switches the gear stage of an automatic transmission. , J3 is a hydraulic control device for an automatic transmission in which the increase in hydraulic pressure is temporarily controlled by an accumulator when the line hydraulic pressure is introduced to the frictional engagement device, the outline of which is not shown in FIG. Thus, the above object has been achieved by making it possible to regulate and control the line oil pressure and the back pressure that determines the characteristics of the accumulator by independent control means. 1 Effect] In the present invention, since the line oil pressure and the back pressure of the accumulator are controlled independently, there is no slippage in the frictional engagement device that is already engaged, and the slippage IiE friction is maintained even when the frictional engagement device is about to engage. It becomes possible to arbitrarily (lower) control the oil pressure of the engagement device. Preferred implementation! ! ! ! Another feature is that the control means is an electromagnetic proportional valve. Preferably, the control means is a duty valve. As for this electromagnetic proportional valve or the duty valve itself, a conventionally known structure can be adopted. Preferably, the line oil pressure is regulated and controlled by the control means depending on the turbine torque. This makes it possible to generate line oil pressure of necessary and sufficient strength without waste, compared to controlling line oil pressure depending only on engine load such as throttle opening. The advantages of setting the line oil pressure in the hydraulic control device depending on the turbine torque are disclosed in more detail in Japanese Patent Application No. 60-263131. Preferably, the back pressure of the accumulator is adjusted and controlled by changing it before and after a predetermined time during gear shifting. By doing so, it becomes possible to perform detailed control that fully takes into account shift shock. Preferably, the predetermined timing is detected depending on either the engine rotational speed or the rotational speed of a rotating member (including a turbine in the torque converter) of the automatic transmission. This makes it possible to accurately detect the predetermined time, specifically, slightly before the end of the shift. [Embodiment 1] An embodiment of the present invention will be described in detail below with reference to the drawings. First, FIG. 2 shows an outline of a vehicle automatic transmission to which this embodiment is applied. This automatic transmission includes a torque converter 20, an overdrive mechanism 40 as its transmission part,
It includes an underdrive mechanism 60 with three forward stages and one reverse stage. The torque converter 20 includes a pump 21 and a turbine 2.
2, a stator 23, and a lock-up clutch 24. The pump 21 is connected to the crankshaft 10 of the engine 1, and the turbine 22 is connected to the crankshaft 10 of the engine 1.
It is connected to the carrier 41 of the planetary tJ vehicle equipment U in the overdrive machine 4I440 via 2A. In the overdrive mechanism 40, a planetary pinion 42 rotatably supported by the carrier 41 meshes with a sun gear r43 and a ring gear 44. Further, a clutch Co and a one-way clutch Fo are provided between the ring gear 43 and the carrier 41, and a brake Bo is provided between the ring gear 43 and the housing HU.
is provided. The underdrive mechanism 60 is provided with two rows of gears, one on the front side and the other on the gear side. This one-star gear unit has a common sun gear 61, ring gear 62, 63, and rotation gear 64.
65 and gear l rear 66.67. A ring gear 44 of the overdrive R structure 40 is connected to the ring gear 62 via a clutch C1. Further, a clutch C2 is provided between the ring gear 44 and the V gear 61. Further, the carrier 66 is connected to the ring gear 63, and the carrier 66 and the ring gear 63 are connected to the output shaft 70. On the other hand, a brake B3 and a one-way clutch F2 are provided between the carrier 67 and the housing 1-1u, and a brake B3 and a one-way clutch F2 are provided between the sun gear 61 and the housing 14t1. B2 is provided, and between the sun gear 61 and the housing Hu,
A brake B1 is provided. This automatic transmission is equipped with a transmission unit as described above, and a computer ( ECLJ) 84 controls the solenoid valve S in the hydraulic pressure control circuit 86 according to a preset shift map.
+ -52 (for shift valve), electromagnetic proportional valve S^ (for accumulator back pressure control: there are several), and electromagnetic proportional valve SL (for line hydraulic pressure control) are driven and controlled, as shown in Figure 3. A combination of engagements of various clutches, brakes, etc. is performed to control the speed change. In FIG. 3, the O mark indicates the engaged state, and the ◎ mark indicates that the engaged state occurs only during driving. As shown in FIG. 4, the solenoid valve S+ controls the 2-3 shift valve, and the solenoid valve S2 controls the 1-2 shift valve and the 3-4 shift valve.
13 Ull. Then, each shift valve 1-2.2-3 controls the underdrive mechanism 60 from the first gear to the third gear, and the 3-4 shift valve controls the overdrive mechanism 40 from the second gear to the third gear. (speed change between the gear position and the fourth gear position) is performed. Also, the electromagnetic proportional valve S^
There are multiple I! i! The back pressure of the accumulator provided in the oil path to the friction engagement HH is controlled, and the NaA proportional valve SL is configured to control the line oil pressure in the oil pressure control f18 and 86 through the primary regulator valve. (described later). Furthermore, in Fig. 2, reference numeral 90 is a shift position sensor, which is operated by the driver (N, neutral), D.
(Drive), R (Reverse), etc. position detection, 92 is a pattern select switch, E (Economy driving)
, P (power running), etc., and a water temperature sensor 4 that detects the cooling water temperature of the 94 +1 engine.
96.98 indicates a brake switch that detects the operation of the foot brake and handbrake, respectively. In FIG. 5, part 15 of the hydraulic up device 86 is not shown. In the above, S^, SL is the electromagnetic proportional valve, 102 is the pump, 103 is the primary regulator valve, 1
04 is the 1-2 shift/<J rev, S2 is the solenoid valve, and 106 is an accumulator for controlling excessive characteristics when hydraulic pressure is supplied to and discharged from the manual valve brake B2 operated by the driver. are shown respectively. The electromagnetic proportional valves S^ and SL are of a well-known structure, and have spools 109A, 110A, and 1o9, respectively.
L, 110L, coils 108A, 108L, springs 113A, 113L, plungers 111A, 1111, etc. Spool IIOA, 110L and plunger 111A, 1
11L, it is integrally movable in the axial direction and is meshed with 0H. The coil 108A1108L is connected to the plungers 111A and 11 according to the load current I^ and IL from the ECU 84.
1 m, thus exerting forces Fc^ and FcL in the downward direction in the figure on the spools 110A and 110L. On the other hand, spring 113
A, 113 [is the force in the opposite direction Fs^% F! 3
Subu L/lz1 1 0A, 1 1 0L1.
: +. lI't. ? The discharge pressure of the pump 112 acts on the bow 1-114L of the 8-magnetic proportional valve SL. Boat 115L and '
The hydraulic pressure acting on U1 1 6L is Pθ, spool 109
If the face area of land 109L^ of L is A1, then P
θ is determined by equation (1). Pθ-(FsL-FCL)/A1--
<1> Therefore, by controlling the force FCL in the downward direction in the figure generated by the coil 108L, the boat 1
15 to any value between 0 and FsL/A+.
can be controlled. This oil pressure Pθ corresponds to the throttle pressure conventionally generated by a throttle valve whose spool can be mechanically driven via a cam in response to the throttle opening, and is generated by the primary regulator valve 103. It acts on the boat 119 as a hydraulic pressure for controlling the line hydraulic pressure. On the other hand, in the primary regulator valve 103, the line oil pressure PL is generated according to the value of the control oil pressure Pθ by the same operation as the conventional one. As a result, the line oil pressure PL can be arbitrarily controlled by controlling the load current It to the coil 108L in accordance with commands from the ECU 34. , it should be noted that the primary regulator valve 10
The pressure regulation relationship in 3 is shown in equation (2). PL=(Fs++(Bz-83>PR +B2Pθ)/B1 (2) Here, Fs+ is the acting force of the spring 120, 81 to B3
is land 121.122.1 of spool 123.124
25 face area. Also, PR is the land 122 when the manual valve 106 is in the reverse range.
and the line oil pressure applied to 125. On the other hand, with exactly the same configuration, the electromagnetic proportional valve S^ regulates the line oil pressure of the oil passage 153 to a desired constant rFPB according to the command of the load electric current if^ from the ECU 34, and controls the oil passage to the back pressure chamber of the 7th cumulator 107. 154. As a result, transient υ control of the hydraulic pressure PB2 is performed when the brake B2 line hydraulic pressure PL is supplied. The oil pressure 2 days 2 during operation of the accumulator 107 is determined by equation (3) depending on the back pressure 1 days as shown in the following equation. P El 2- F s a + (C+ C2)
P e / Ct (3) Here, F94 is the acting force of the spring 136, C+,
C2 is the face area of the two lands of the accumulator piston 137. From the above equations (1> to (3)), the line oil pressure and the oil pressure PB2 to the brake B2 can be arbitrarily controlled independently by the load current IL11^ commanded from the ECU 34. A few explanations will be given regarding the connection between the brakes and the brake B2.The signal pressure of the solenoid valve S2 acts on the boat 126 of the 1-2 shift valve 104.Therefore, the 1-2 shift The spool 127 of the valve 104 is driven from the right side to the left side in the figure in response to the ON-OFF state of the solenoid valve S2.The force Fss of the spring 128 causes it to slide to the right side. 13
3 and 129 are connected. The line oil pressure PL from the boat 130 of the manual valve 106 is connected to the boat 129.
(drive) It is designed to work on the range. That is, the boats 130, 129, and 133 are connected at the D range selection position of the spool 131 of the manual valve 106. On the other hand, the boat 133 has an oil passage 135,
It is connected to brake B2 via a check valve 134. Therefore, in the D range, the 0 of solenoid valve S2 is
The line hydraulic pressure PL is supplied to and discharged from the brake B2 by N-OFF. FIG. 6 shows a control flow in the apparatus of the above embodiment. In steps 201 to 203, the rotational speed m, No., and throttle opening degree θ of the output shaft 70 of the engine rotating device Ne1 automatic transmission are read, respectively. In step 204, the turbine rotation speed (rotation speed of the turbine shaft 22A) NT is calculated from the product of the current gear ratio i and the output shaft rotation speed NO. Further, in step 205, the engine rotation speed N
The speed ratio e is determined from the quotient of e and the turbine rotational speed NT. Furthermore, in step 206, the torque ratio t is determined from this speed ratio e.
seek. The relationship between the speed ratio e and the torque ratio can be determined, for example, as shown in FIG. 7. In step 207, the engine rotation 311!
l, Ne, throttle opening θ, and engine torque Te.
Search. An example of a map of this engine torque Te is shown in FIG. In step 208, one turbine torque is determined by the product of the engine torque Te and the torque ratio. F in step 209 is a flag for flow υ1 group. In the case of F-0, it is determined in step 210 whether or not to shift,
If the gear is not in the gear shift state, in step 211, the load current 1^3 and ILl to the electromagnetic proportional valve S8%SL are searched based on the gear stage type and the turbine torque 1, respectively. An example of this search relationship is shown in Figure 9. Here, the load currents I^3 and ILl to the electromagnetic proportional valves S^ and SL are set to be equal. On the other hand, if it is determined in step 210 that there is a shift determination (represented here by the case where there is a power-on upshift determination), the shift output is performed in step 220, and the shift type and turbine torque Tt are determined in step 222. Determine I^4 and rL4 depending on . This example is not shown in FIG. Here, ra■^4 and ILl may have the same value. ■^, ILl is set smaller than I^3, ILl, and the oil pressure is increased accordingly. This is a consideration to prevent slipping of the friction engagement device that is already engaged, since an inertia torque occurs when shifting compared to when not shifting. In step 224, the start of the inertia phase (a period during which the rotating members of the automatic transmission change their rotational speeds for gear shifting) is determined by the fact that NT≦No Confirm IL5. In this case, the first
As shown in Figure 0, IL5 here is step 22.
It is preferable to take it equal to ILl in 2. however,
As for 〇 to I, in step 228, it is set to a value that is different from i＾喀. Specifically, 1^4+α(1/
It is best to find it by -1)X■^4. Here, K is a correction coefficient for the turbine rotational speed NT, and can be determined as K-(current turbine rotational speed Nt/turbine rotational speed NTO at the inertia phase start time). Further, α is an N N T correction strength coefficient, which is determined depending on the type of shift, the time during the inertia phase, and the throttle opening degree θ. As represented by the shift from the first speed to the second speed, the step 230 starts after the inertia phase starts.
The values shown in the left column of FIG. 11 are used until the predetermined time detected in , and after the predetermined time, the values shown in the right column of the same figure are used. Then, according to the value of α, fAa in step 228 and fAa in step 234 are determined according to the above-mentioned formula.
The I^f of is determined. Note that the predetermined time in this step 230 is NT≦
It is detected based on whether or not N o Xi l-1+N 1 holds true. Here, iH is the gear ratio of the high gear stage. This predetermined time is qualitatively set to be slightly before the end of the shift. Step 232 has the same content as step 226. Step 236 is a step for detecting the end of the shift based on whether the turbine rotational speed NT becomes smaller than the product of the high gear stage gear ratio tH and the output shaft rotational speed No. Also, step 219 performs l in each process.
This is a step for outputting the ILl 1^ that has been processed. According to this embodiment, since the accumulator back pressure can be set independently of the line oil pressure, the accumulator back pressure is kept low (particularly near the end of the shift), thereby reducing the working oil pressure of the PJ friction engagement device. can be lowered, and the shift shock can be reduced accordingly.Also,
By making the line oil pressure dependent on the turbine torque, completely independent of the I11 control during gear shifting, it is now possible to maintain the oil pressure at the minimum required level at all times (see solid line in Figure 12). In addition, even if an inertia torque occurs during a gear shift, the line oil pressure can be increased by an amount corresponding to the increase in the inertia torque (see the dashed line in the figure). Note that diagram M12 shows a diagram in which the turbine torque at the time of shifting from the first gear to the second gear is converted into clutch CI. As a result, unnecessary power loss during non-shifting can be avoided, and the amount of lubrication and flow rate of the cooler can be optimized. Furthermore, it is possible to freely control the transient hydraulic pressure of the 8 frictional engagement devices that are about to engage without causing slippage in the frictional engagement devices that are currently engaged during gearshifting, thereby reducing shocks during gearshifting. can be made smaller. In the above embodiment, from the first gear to the 23rd! !
Although the explanation has been given using an example of shifting to a gear, it is clear that the present invention is similarly applicable to shifting to other gears. [Effect of R light] As explained above, according to the present invention, the line oil pressure is always controlled to the minimum necessary oil pressure, and the line oil pressure is controlled so that the friction engagement device that is in the engaged state does not slip when changing gears. While it is possible to control the increase in oil pressure, the back pressure of the 7th cumulator can be freely reduced, and as a result, an excellent effect is obtained in that the shock during gear shifting can be reduced by 4 Jli.

[Brief explanation of drawings]

Fig. 1 is a block diagram based on the gist of the present invention;
FIG. 3 is an overall skeleton diagram of a vehicular automatic transmission to which an embodiment of the hydraulic control device for an NJ transmission according to the present invention is applied, and shows the operating state of the frictional engagement position a in the automatic transmission. The diagram, FIG. 4, shows the same <1. Fig. 5 is a diagram showing the input/output relationship of the III system; Fig. 5 is a hydraulic circuit diagram of the main parts of the hydraulic pressure t and +3 control device; Fig. 6 is a flowchart showing the control flow used in the above embodiment; 7 and 8 are diagrams each showing an example of a map of the torque ratio t1 engine torque Te, and FIGS. 9 and 10 are diagrams each showing an example of a map for determining the load current. 12 is a diagram showing an example of a map of the NT correction strength coefficient α, and FIG. 12 is a diagram showing an example of clutch C1 converted torque with respect to throttle opening in order to qualitatively show the effect of the above embodiment. It is. 40... Overdrive mechanism, 60... Underdrive mechanism, 84... ECU. 86... Hydraulic control circuit, S^, SL... Solenoid proportional valve, 103... Primary regulator valve, 107...
··accumulator.

Claims (6)

[Claims] (1) The engagement state of the frictional engagement device is switched by the line hydraulic pressure adjusted from the basic hydraulic pressure of the oil pump, and the gear stage of the automatic transmission is changed, and the line hydraulic pressure is guided to the frictional engagement device. In a hydraulic control device for an automatic transmission in which an increase in hydraulic pressure is temporarily controlled by an accumulator, the line hydraulic pressure and the back pressure that determines the characteristics of the accumulator are controlled by independent control means. Pressure/
A hydraulic control device for an automatic transmission, which is characterized by being able to control the transmission. (2) The hydraulic control device for an automatic transmission according to claim 1, wherein the control means is an electromagnetic proportional valve. (3) The hydraulic control device for an automatic transmission according to claim 1, wherein the control means is a duty valve. (4) The oil pressure control device for an automatic transmission according to any one of claims 1 to 3, wherein the line oil pressure is regulated and controlled by the control means depending on turbine torque. (5) The automatic transmission according to any one of claims 1 to 4, wherein the back pressure of the accumulator is adjusted and controlled by changing the back pressure of the accumulator before and after a predetermined time during gear shifting by the control means. Hydraulic control device. (6) The hydraulic control device for an automatic transmission according to claim 5, wherein the predetermined timing is detected depending on either an engine rotational speed or a rotational speed of a rotating member of the automatic transmission.