JP2897040B2 - Hydraulic control device for automatic transmission - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION [Industrial applications]

The present invention provides an automatic transmission configured to take into account the effect of centrifugal hydraulic pressure generated by rotation of a hydraulic servo of a friction engagement device when controlling a transient engagement pressure of the friction engagement device. To a hydraulic control device.

[Prior art]

Generally, an automatic transmission includes a gear transmission mechanism and a plurality of frictional engagement devices, and by selectively operating the engagement states of the plurality of frictional engagement devices by operating a hydraulic control device, a plurality of frictional engagement devices are provided. It is configured such that any one of the gears is achieved. The engagement of the friction engagement device is performed by supplying a line pressure to a hydraulic servo of the friction engagement device. In this case, in order to reduce the shifting shock, a configuration in which the transient engagement pressure is arbitrarily controlled by real-time feedback control of the back pressure of the accumulator has been widely adopted. The back pressure of the accumulator is controlled by applying a throttle pressure (a hydraulic pressure generated according to the throttle opening) or a line pressure (a basic control hydraulic pressure regulated according to the throttle pressure) to the back pressure chamber of the accumulator. This is performed by inputting the hydraulic pressure adjusted by the control solenoid valve. If the frictional engagement device to be engaged is a clutch, the hydraulic servo of the clutch shifts from a stopped state to a rotating state with the engagement. Even if the engagement pressure is precisely controlled by the feedback control using an accumulator, the centrifugal oil pressure is added and the actual engagement pressure becomes higher than the intended engagement pressure. Occur. In view of such a point, conventionally, in controlling the engagement pressure of the friction engagement device, particularly the clutch, the engagement is performed in consideration of the effect of the centrifugal hydraulic pressure generated by the rotation of the hydraulic servo of the clutch. A technique for performing pressure control has been proposed (Japanese Patent Application Laid-Open No. 60-49158, Japanese Patent Application No. 1-133737).

[Problems to be solved by the invention]

However, in the related art in which the centrifugal oil pressure is taken into consideration, the adjustment is performed by a solenoid valve for adjusting the back pressure of the accumulator. However, this solenoid valve for back pressure adjustment originally adjusted the engagement pressure finely, so even if we tried to consider the effect of centrifugal oil pressure, it could not be adjusted as it was. There is a problem that a solenoid valve for back pressure adjustment (and an accumulator back pressure control valve) having a wide range must be prepared. In other words, when a (large) solenoid valve for adjusting the back pressure that can absorb the influence of the centrifugal oil pressure is used, not only the cost is increased but also the centrifugal oil pressure is originally generated like a brake. In the case of controlling a frictional engagement device that does not perform the adjustment, there is a problem that it is difficult to perform fine and delicate pressure adjustment. SUMMARY OF THE INVENTION The present invention has been made in view of such a conventional problem, and makes it possible to always perform a fine and delicate pressure control by a solenoid valve for adjusting a back pressure of an accumulator, while effectively controlling the influence of centrifugal hydraulic pressure. It is an object of the present invention to provide a hydraulic control device for an automatic transmission which can be eliminated.

[Means for Solving the Problems]

The present invention, as shown in FIG. 1, provides a line pressure to a hydraulic servo of a friction engagement device, and when the friction engagement device is engaged, the transient engagement pressure is reduced. The throttle pressure or the line pressure is controlled by inputting a hydraulic pressure adjusted by a back pressure adjusting solenoid valve to the back pressure chamber of the accumulator, and the hydraulic servo of the friction engagement device is controlled in the control. In a hydraulic control apparatus for an automatic transmission configured to take into account the effect of centrifugal oil pressure generated by rotation, means for detecting the type of shift and whether or not there is an effect of the centrifugal oil pressure on the type of conversion. Only when the speed is determined to be affected by the centrifugal oil pressure, and at least one of the throttle pressure solenoid valve or the line pressure solenoid valve that regulates the throttle pressure or the line pressure itself. Compensate for the effects of centrifugal oil pressure By which is adapted to, in which to achieve the above objects.

[Action]

In the present invention, the effect of the centrifugal oil pressure is corrected not by the back pressure adjusting solenoid valve of the accumulator but by the throttle pressure or the throttle pressure solenoid valve or the line pressure solenoid valve which regulates the line pressure itself. I am trying to do it. That is, in general, the back pressure of the accumulator is adjusted by adjusting the throttle pressure or the line pressure with a solenoid valve for adjusting the back pressure, but the solenoid valve for adjusting the back pressure absorbs all the effects of the centrifugal oil pressure. To do so, as described above, a design change such as making the solenoid valve for adjusting the back pressure considerably large is required. This design change not only results in an increase in weight and cost, but also causes a problem that a minute and delicate pressure adjustment cannot be performed when controlling a friction engagement device that does not generate centrifugal hydraulic pressure. Therefore, the effect of the centrifugal oil pressure is corrected by using a throttle pressure solenoid valve or a line pressure solenoid valve that regulates the throttle pressure or the line pressure itself. Since the throttle pressure or the line pressure can be considerably changed in accordance with the throttle opening of the engine, it is easy to absorb the centrifugal oil pressure by the throttle pressure solenoid valve or the line pressure solenoid valve. . Into the back pressure chamber of the accumulator, the throttle pressure or the line pressure already adjusted for the centrifugal oil pressure and the oil pressure finely adjusted by the back pressure adjusting solenoid valve are input.
Therefore, in the back pressure adjusting solenoid valve, it is not necessary to consider the effect of the centrifugal oil pressure at all, and only the fine adjustment in the original feedback control is sufficient. That is, the control logic can be shared for both the clutch that generates centrifugal hydraulic pressure and the brake that does not generate centrifugal hydraulic pressure, and also enables fine and delicate pressure adjustment. However, if the centrifugal oil pressure is simply corrected by the throttle pressure or the line pressure, an erroneous correction may be performed in a shift that does not need to be corrected. In the case of the present invention, it is determined whether or not the centrifugal oil pressure needs to be corrected depending on the type of the shift, and the correction is performed only in the shift that needs to be corrected. It can prevent that much.

【Example】

Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. FIG. 2 shows an overall outline of an automatic transmission for a vehicle to which this embodiment is applied. The power of the engine 1 is transmitted to the automatic transmission via the shaft 10. This automatic transmission includes a torque converter section 20, an overdrive mechanism section 40, and an underdrive mechanism section 60 having three forward stages and one reverse stage as its transmission section. The torque converter section 20 includes a pump 21 and a turbine 2
2. It is a well-known device having a stator 23 and a lock-up clutch 24. The overdrive mechanism section 40 includes a sun gear 43, ring gear 44, planetary pinions 42, and includes a pair of planetary gear unit consisting of Kiyariya 41, clutch C ₀ the rotation state of the planetary gear device, the brake B _0, the one-way by the clutch F ₀ are connexion control. The underdrive mechanism 60 includes a common sun gear 6
1, two sets of planetary gear units including ring gears 62 and 63, planetary pinions 64 and 65, and carriers 66 and 67 are provided, and the rotational state of the two sets of planetary gear units and the connection state with the overdrive mechanism are clutched. C _1, C _2, the brake B ₁
.About.B _3, and are by connexion controlled freewheel F _1, F _2. Since the transmission unit is well known per se, the specific connection state of each component is only shown in the skeleton in FIG. 2 and detailed description is omitted. This automatic transmission includes the transmission unit and the computer 84 as described above. The computer 84 includes a throttle sensor 80 for detecting the throttle opening θ for reflecting the output (torque) of the engine 1, a vehicle speed sensor (rotation speed sensor for the output shaft 70) 82 for detecting the vehicle speed n _0, and a centrifugal oil pressure. detecting the rotational speed of the clutch C ₂ in order to correct
Each signal such as NT sensor 99 for detecting the rotational speed NT of the turbine 22 of the automatic transmission to reflect C ₂ sensor 98, and a shift transient state is entered. The computer 84 drives and controls the solenoid valves S ₁ and S ₂ (for the shift valve) and SL (for the lock-up clutch) in the hydraulic control circuit 86 according to a preset throttle opening-vehicle speed shift map. , Third
The shift is executed by performing a combination of engagement of each clutch, brake and the like as shown in the figure. In FIG. 2, S _LN is an electromagnetic proportional valve for controlling the back pressure of the accumulator, and S _LT is an electromagnetic valve for controlling the throttle pressure (hence, the line pressure). FIG. 4 shows a main part of the hydraulic control circuit 86. In FIG, S _LT said Surotsutoru pressure control proportional solenoid valves, 102 pumps, 103 primary the regulation Yoo regulator valve, 104 1-2 shift valve, is Maniyuaru valve 106 is by connexion operated driver . Proportional solenoid valve S _LT is of itself well known, the spool 109, a coil 108, a spring 113, plunger
It is composed of 111 etc. Spool 110 and plunger 11
1 is meshed so as to be integrally movable in the axial direction. The coil 108 exerts a downward force F _C on the plunger 111 and thus the spool 110 according to the load current I _{P (} considering the centrifugal oil pressure from the computer 84). On the other hand, the spring 113 exerts a force F _S or the like in the opposite direction on the spool 110. The discharge pressure of the pump 102 acts on the port 114. The oil pressure acting on ports 115 and 116 is Pθ, the spool
The Fueisu area of the land 109A of 109 When A _1, Pθ is obtained by (1). _{Pθ = (F S -F C)} / A 1 ...... (1) accordance connexion, force drawing down direction by connexion generated in the coil 108 F
By controlling _C , Pθ generated at port 115 is reduced to 0.
~ (F _S / A ₁ ) can be controlled to any value. Conventionally, the hydraulic pressure Pθ is equivalent to a throttle pressure generated by a throttle valve in which a spool can be mechanically driven in accordance with a throttle opening degree via a cam. The port 119 acts on the port 119 as a control oil pressure for the line pressure generated by the valve 103. In the primary regulator valve 103, the line pressure is adjusted according to the value of the control oil pressure
Generate PL. As a result, eventually by controlling an additional current I _P to Yotsute coil 108 with a command computer 84, it becomes possible to arbitrarily control the line pressure PL. Load current I _P of the coil 108 is conventionally electrically detects Surotsutoru opening, it has been assumed in accordance with the detected Surotsutoru opening. In this embodiment, the load current _IP is determined in consideration of the centrifugal oil pressure in addition to the above operation. The degree of the influence may be determined in advance depending on the type of the shift (the type of the clutch), or the actual rotational speed of the clutch may be detected as described later, and this detection may be performed. The correction may be made in real time based on the rotation speed. Note that the pressure regulation relational expression in the primary regulator valve 103 is shown in Expression (2). PL = {F _S2 + (D ₂ −D ₃ ) P _R + D ₂ Pθ} / D ₁ (2) where F _S2 is the acting force of the spring 120, and D _{1 to} D ₃ are the spools 123 and 124. The face area of lands 121, 122, and 125. Also, P _R is the line pressure Maniyuaru valve 106 is applied to the lands 122 and 125 when in the reverse range. On the other hand, 1-2 port 126 of shift valve 104, the signal pressure of the shift solenoid valve S ₂ is applied. Therefore,
Spool 127 of the 1-2 shift valve 104, in response to ON-OFF of the shift solenoid valve S ₂ right figure - to slide to the left. Sliding to the right is due to the force F _{S3 of the} spring 128. At this time, the ports 133 and 129 of the 1-2 shift valve 104 are connected. Port 129 has port 13 of manual valve 106
The line hydraulic pressure PL starting from 0 operates in the D (drive) range. That is, the ports 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, port 133 is
It is connected to the friction engagement device via 35. Oil passage 13
An accumulator is connected to 5, and controls a transient hydraulic pressure level when supplying and discharging the line hydraulic pressure PL to and from the friction engagement device. This will be described with reference to FIG. In FIG. 5, reference symbol S _LN denotes a linear solenoid for back pressure control, 208 denotes an accumulator control valve, 210
Is a modulator valve, and 212 is an accumulator. In this figure, the brake B ₂
Are representatively shown. As is clear from FIG.
Brake B ₂ is a friction engagement device is engaged when achieving a shift to the second speed stage from the first speed stage, not take particular centrifugal hydraulic pressure. First, the oil circuit arrangement as described above, the line pressure PL is created in consideration to centrifugal oil pressure (considered as a centrifugal hydraulic In brake B ₂ is zero). This line pressure PL is a branch point
It is applied to port 210A of modulator valve 210 via oil line 139 from 136 (FIG. 4). Modulator valve 2
10 receives a predetermined modulator pressure Pm in response to the line pressure PL.
Occurs at port 210B in a well-known manner. Linear solenoid S _LN generates a solenoid pressure PS ₁ corresponding receiving this Mojiyureta pressure Pm to the difference between the turbine speed NT and the turbine target rotation speed NT ₀ in a known manner. That is, the rotation speed NT of the turbine 22 is input to the computer 84 as described above. The turbine speed NT is compared to a turbine target rotation speed NT ₀ set in advance according to the type of engine torque and speed. For example, in the case of a 1 → 2 shift, the execution of the 1 → 2 shift lowers the turbine rotational speed NT. If turbine speed NT
Falls earlier than the target rotation speed NT ₀ (NT−NT ₀ <
0), the shift is going too fast,
In order to reduce the engagement transition oil pressure of the brake B _2, the NT-
A load current command based on the duty ratio corresponding to NT ₀ is applied to the linear solenoid S _LN , and the linear solenoid S
_LN is to generate the solenoid pressure PS ₁ in accordance with the load current in a known manner. In this embodiment, when the duty ratio increases (10
It approaches 0%, the), and summer as solenoid pressure PS ₁ is increased to be generated. The solenoid pressure PS ₁ is input to the port 208A of Aki Yumu regulator control valve 208. The accumulator control valve 208 controls the line pressure PL ₁ (the line pressure of the oil passage 135 = the same oil pressure as the line pressure PL considering the centrifugal oil pressure generated in the oil passage 139 of FIG. 4) and the linear solenoid S _LN. as input signals the solenoid pressure PS ₁ from the port 208B of the line pressure PL ₂ (oil passage 135 line pressure = 4 view of the line pressure P of considering centrifugal hydraulic pressure generated in the branch point 136
Pressure at the same level as L) is adjusted to the accumulator back pressure Pac. That is, the accumulator back pressure Pac is, in other words, basically the line pressure PL ₂ (= PL) adjusted by the line pressure PL ₁ (= PL) and the urging force of the spring 108C, and the linear solenoid _SD those that had it occurred correction to the solenoid pressure PS _1. It should be noted that the accumulator back pressure Pac has a characteristic of decreasing as the duty ratio increases. When the shift determination (in this case, the shift from the first speed to the second speed) is performed by the computer 84, the solenoid valve S ₁
The shift valve 114 is switched in a known manner via
Line pressure PL (P _B0 ) generated by the hydraulic control circuit in FIG.
There starts to be aerodrome supplied to the brake B ₂ via the oil passage 135. In response to this supply, the piston 212A of the accumulator 212
Begins to rise. While the piston 212A is increasing, the hydraulic pressure supplied to the brake B ₂ (P _B0) is maintained substantially constant hydraulic downward force and balancing ivy acting downward biasing force and the piston 212A of the spring 212B Will be done. The force that pushes the piston 212A downward is equal to the accumulator back pressure P acting on the back pressure chamber 212C of the accumulator 212.
Generated by ac. Therefore, the accumulator back pressure
Mogi Yu regulator valve 210 as described above to pac, to arbitrarily control the transient pressure P _B0 upon engagement of the linear solenoid S _LN and Aki Yumu regulator control Yotsute brake B ₂ to be controlled via a valve 208 It becomes possible. Specifically, when the duty ratio is further increased (closed to 100%), the engagement pressure is further reduced. Linear solenoid S _LN, as described above, to be controlled in dependence on the difference between the turbine rotational speed NT and the turbine target rotation speed NT _0, after all, by such hydraulic system, the turbine rotational speed NT turbine target the rotational speed NT ₀ can be fed back controlled so that along connexion changes. Moreover, in this embodiment, since the line pressure itself is already set in consideration of the centrifugal oil pressure, the control of the accumulator back pressure Pac is performed without considering the centrifugal oil pressure at all. depends only on the deviation between the speed NT ₀ now suffices work simply fed back control, the solenoid valve of large is not required and it is possible to perform a fine and delicate control easily. That is, in controlling the back pressure Pac of the accumulator, the logic of the feedback control can be largely shared between the shift in which the centrifugal hydraulic pressure is not generated and the shift which is generated as in this embodiment. FIG. 6 shows an example of this control logic. First, in steps 301 and 302, the value of the flag F is determined. This flag F is for controlling the control flow, and is initially set to zero, so that the process proceeds to step 303. In step 303, the type of shift (and the presence or absence of centrifugal oil pressure in connection therewith) is determined. Specifically, for example, it is determined whether the shift is from the first gear to the second gear or from the third gear to the fourth gear, and when the determination of YES is made, the centrifugation is particularly performed. Since the shift is not related to the hydraulic pressure, the process proceeds to step 304 to output a shift output to that effect. From step 305 onwards, known feedback control is performed. In brief, first, at step 305, it is determined whether or not the inertia phase has been entered (a period during which the rotating member of the automatic transmission is undergoing a change in the rotational speed for shifting), and it is determined that the inertia phase has been entered. that the target value NT ₀ of turbine speed NT at step 306 (depending on the type of transmission and Surotsutoru opening degree) is set. In step 307, it is determined whether or not the feedback control is "present". If the feedback control is "present", the process proceeds to step 308, and a duty output to that effect is output, and the feedback control is executed. At step 309, it is determined whether or not the control for temporarily reducing the engagement pressure at the end of the shift is "present". When it is determined that the control at the end of the shift is "present", the routine proceeds to step 310, where a duty output for reducing the engagement pressure is issued. At step 311, it is determined whether or not the shift is completed. Unless the shift is completed, the flag F is set to 1 at step 315 via step 314, and step 305 and subsequent steps are repeated via step 301. Eventually, when it is determined that the shift has been completed, the routine proceeds to step 312, where the duty ratio is reset to zero, and at step 313, the flag F is reset to zero. On the other hand, if the type of shift is a shift from the second speed to the third speed, the process proceeds to step 317 to output a shift output. At step 318, it is determined whether or not the inertia phase has been reached. If it is determined that the inertia phase has been reached, step 319 is performed.
, A target value NT ₀ of the turbine speed NT is set.
In step 320, the rotational speed N _C2 of the clutch C ₂ from the second speed stage to be engaged during shifting to the third speed stage is read, the control oil pressure for controlling the line pressure PL at step 321 ( corresponding to the conventional Surotsutoru pressure) Pshita is changed to _{Pθ = Pθ 0 -α × Nc C2} 2 ...... (3). Here, alpha is a constant depending on the distance from the rotation center of the clutch C _2, the influence degree is Matsupu of for different for the line pressure for each Surotsutoru opening. Thereafter, the flow proceeds to step 307, and the above-described flow is similarly executed. FIG. 7 shows shift characteristics at the time of shifting from the second speed to the third speed. In the figure, (A) shows the shift characteristic of the conventional technology, and the hatched portion corresponds to the centrifugal oil pressure. As is clear from the duty ratio line, the effect of the centrifugal oil pressure is too strong, and the duty ratio has reached 100%, indicating that the duty ratio has not been completely offset. In this case, it is obvious that if an attempt is made to prepare an electromagnetic valve that can completely offset the centrifugal hydraulic pressure, the electromagnetic valve will increase in size or cost accordingly. However, when the centrifugal oil pressure component is offset by lowering the line pressure as shown in FIG. 7B, the duty ratio only needs to handle the feedback control. It can be seen that sufficiently small and delicate feedback control can be realized even with a small solenoid valve.

【The invention's effect】

As described above, according to the present invention, the influence of the centrifugal oil pressure is not considered by the solenoid valve for controlling the accumulator back pressure, but is considered by the solenoid valve for controlling the throttle pressure or the line pressure. Therefore, the above consideration can be made extremely easily, so that it is not necessary to use a particularly large solenoid valve for controlling the accumulator back pressure, and fine and delicate adjustment can be performed at low cost. Is obtained. Also, regarding the back pressure control of the accumulator, it is not necessary to separately control the shift in which the centrifugal hydraulic pressure is generated and the shift in which the centrifugal hydraulic pressure is not generated, and these control logics can be shared. Furthermore, whether or not a centrifugal oil pressure is generated is determined based on the type of shift, and the shift operation in which no centrifugal oil pressure is generated does not execute the correction operation itself, so that erroneous correction can be prevented accordingly.

[Brief description of the drawings]

FIG. 1 is a block diagram showing the gist of the present invention, FIG. 2 is a skeleton diagram showing an overall configuration of an automatic transmission for a vehicle according to an embodiment of the present invention, and FIG. 3 is a friction diagram of the automatic transmission. 4 is a diagram showing the vicinity of an electromagnetic proportional valve for controlling line pressure, and FIG. 5 is a diagram showing the vicinity of an electromagnetic valve for controlling back pressure of an accumulator. FIG. 6 is a flowchart showing a control flow executed in the apparatus of the above embodiment, and FIG. 7 is a shift transient characteristic diagram showing the effect of the above embodiment in comparison with the conventional example. It is. S _LT …… Proportional solenoid valve (solenoid valve for line pressure control), S _LN …… Linear solenoid (electromagnetic valve for back pressure control of accumulator), B ₂ …… Brake (friction system that does not generate centrifugal hydraulic pressure during shifting) Joint device), C ₂ … clutch (frictional engagement device that generates centrifugal oil pressure during gear shifting), Pθ… control oil pressure (oil pressure equivalent to conventional throttle pressure), PL… line pressure, Pac… accumulator back Pressure.

──────────────────────────────────────────────────続 き Continuation of front page (56) References JP-A-2-46364 (JP, A) JP-A-2-180361 (JP, A) JP-A-57-204350 (JP, A) JP-A-59-204 34049 (JP, A) (58) Fields investigated (Int. Cl. ⁶ , DB name) F16H 59/00-61/12 F16H 61/16-61/24 F16H 63/40-63/48 B60K 41/00 -41/28

Claims (5)

(57) [Claims] When the frictional engagement device is engaged by supplying a line pressure to a hydraulic servo of the frictional engagement device, the transient engagement pressure is supplied to a back pressure chamber of the accumulator by a throttle pressure. Alternatively, the line pressure is controlled by inputting a hydraulic pressure adjusted by a back pressure adjusting solenoid valve, and the centrifugal hydraulic pressure generated by the rotation of the hydraulic servo of the friction engagement device during the control. Means for detecting the type of shift, means for determining whether there is an effect of the centrifugal oil pressure depending on the type of conversion, and centrifugal oil pressure. As long as the shift is determined to have an effect, the effect of the centrifugal oil pressure is corrected using at least one of the throttle pressure solenoid valve or the line pressure solenoid valve that regulates the throttle pressure or the line pressure itself. Features Hydraulic control device for automatic transmission. 2. The hydraulic control of an automatic transmission according to claim 1, wherein the correction relating to the effect of the centrifugal oil pressure is performed depending on the square of the rotation speed of a hydraulic servo of the friction engagement device. apparatus. 3. The automatic transmission according to claim 1, wherein the correction relating to the effect of the centrifugal oil pressure is performed depending on a distance from a rotation center of a hydraulic servo of the friction engagement device. Hydraulic control device. 4. The hydraulic control device for an automatic transmission according to claim 1, wherein the correction relating to the effect of the centrifugal hydraulic pressure is performed using a line pressure solenoid valve. 5. The hydraulic control device for an automatic transmission according to claim 1, wherein the correction relating to the effect of the centrifugal hydraulic pressure is started after the inertia of the shift is started.