JPH0526062B2 - - Google Patents

Description

[Detailed description of the invention] [Industrial application field]

The present invention relates to an improvement in a hydraulic control device for upshifting an automatic transmission.

[Conventional technology]

A gear transmission mechanism and a plurality of frictional engagement devices are provided, and engagement of the frictional engagement devices is selectively switched by operating a hydraulic control device, and any one of the plurality of gears is set. Automatic transmissions for vehicles configured to achieve this are already widely known. The frictional engagement device generally includes 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 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 torque transmission between them. The working oil pressure for the frictional engagement device is generally referred to as line pressure or line oil pressure. Conventionally, in view of the above conditions, this line oil pressure is usually changed according to a value that is considered to represent 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 increases. ing. This line oil pressure control has traditionally been carried out by introducing a throttle oil pressure that changes depending on the throttle opening into a control port of a primary regulator valve for controlling the line oil pressure. This throttle oil pressure is generally generated by a throttle valve whose spool is subjected to a spring force that increases in accordance with the amount of depression of the accelerator pedal. In recent years, electronic automatic transmissions have been developed, and the main part of the control circuit is now composed of electronic circuits. Since information regarding the throttle opening is also handled in the form of electrical signals, devices have been developed in which line oil pressure is controlled based on electrical signals regarding the throttle opening (for example,
56−12555). In a device where the line oil pressure is controlled based on an electric signal, it is theoretically possible to control the line oil pressure to an arbitrarily set pressure regulation target value, and therefore, it is possible to The control oil pressure at the time can also be controlled according to the target value. Furthermore, for example, in JP-A-61-130653 and JP-A-61-188249, a technique was proposed for controlling transient hydraulic pressure (engagement pressure) during gear shifting by controlling accumulator pressure based on an electric signal. Although,
This technique also makes it possible to control the control oil pressure during a shift transition to a target value. However, how to set the pressure regulation target value (command value to the pressure regulation means) at the time when the shift is started is an extremely important problem. On the other hand, Japanese Patent Laid-Open No. 151444/1983 discloses a technique for feedback controlling the control hydraulic pressure in accordance with the changing tendency of the input shaft rotational speed of an automatic transmission.

[Problem to be solved by the invention]

However, in feedback control, if the gain is too small, the followability will be degraded, and if the gain is too large, the system may vibrate, so it is not always easy to perform accurate and stable control. Moreover, in JP-A-60-151444, the pressure regulation target value was determined uniquely from the change trend of the input shaft rotational speed, so even if the feedback control itself was performed well, the target value itself was not necessarily optimal. As a result, there was also a problem in that a sufficient shift shock reduction effect could not be obtained.

[Purpose of the invention]

The present invention has been made in view of such conventional problems, and has a relatively simple configuration, and is capable of properly controlling the control oil pressure in a hydraulic control device to an intended optimum pressure regulation value during a shift transition. An object of the present invention is to provide a hydraulic control device for upshifting an automatic transmission.

[Means to solve the problem]

The present invention provides a hydraulic control device for upshifting an automatic transmission, as shown in FIG. means capable of arbitrarily regulating the control hydraulic pressure in the hydraulic control device; and a means for adjusting the pressure regulation target value of the control hydraulic pressure, which changes from moment to moment in the hydraulic control device during upshifting, to at least the rotation of the engine or automatic transmission. Means for setting in real time depending on the ever-changing rotational speed of any of the members, and a means for changing the degree of dependence of the rotational speed stepwise or continuously during gear shifting, and controlling hydraulic pressure immediately before engagement. The above object has been achieved by providing a means for creating a characteristic of gradually reducing the amount of water.

[Effect]

In the present invention, first, open control is adopted instead of feedback control. Therefore,
The control system can be simplified, and problems such as system vibration do not occur. Next, in the present invention, the qualitative tendency of the ideal pressure regulation target value of the control oil pressure in the hydraulic control device during a shift transition, the engine rotation speed during a shift transition,
Or, focusing on the fact that the qualitative trends of the rotational speeds of the rotating members of the automatic transmission are almost the same, the control oil pressure during gear shifting can basically be set to either the engine rotational speed or the rotating member of the automatic transmission. The setting depends on the rotation speed. In other words, the control oil pressure during a shift transition is as follows:
Ideally, it should be lowered gradually. This is because in the latter half of the inertia phase, the energy that the frictional engagement device should absorb has already become quite small, so a high control hydraulic pressure is not necessary, and from the perspective of reducing the shift shock, it is rather low. This is because if the pressure is regulated, the change in the torque of the output shaft at the end of the inertia phase will be smaller, and the shift characteristics will be improved accordingly. Incidentally, the engine rotational speed or the rotational speed of the rotating members of the automatic transmission tends to gradually decrease (some of the rotating members of the automatic transmission increase) with the start of the inertia phase. Therefore, by utilizing this tendency to determine the pressure adjustment target value (command value to the pressure adjustment means) of the control oil pressure during a shift transition, it becomes possible to control the control oil pressure almost ideally. . Note that when a rotating member whose rotational speed increases with the start of the inertia phase is selected as the object of dependence, it is only necessary to make it depend after multiplying it by a negative proportionality constant. Moreover, in the present invention, the degree of dependence of the rotational speed is changed stepwise or continuously during gear shifting. In other words, the engine rotational speed or the rotational speed of the automatic transmission members during gearshifting generally decreases or increases monotonically, but by changing the degree of dependence stepwise or continuously during gearshifting, the rotational speed of the automatic transmission members increases or decreases. It becomes possible to control the control hydraulic pressure according to a more ideal curve in accordance with the temporal process, and even with open control, very good shift characteristics can be obtained. Preferably, the degree of dependence of the rotational speed is
It is to be changed depending on at least the type of transmission, engine load, and vehicle speed. This enables more accurate control depending on the operating state at that time. Note that the control oil pressure may include line oil pressure, oil pressure for line oil pressure control, oil pressure applied to the back pressure chamber of the accumulator, oil pressure in the oil passage immediately before the frictional engagement device, and the like. That is, the present invention does not limit the target control oil pressure.

【Example】

Embodiments of the present invention will be described in detail below based on the drawings. First, FIG. 2 shows an overall outline of a vehicle automatic transmission to which this embodiment is applied. This automatic transmission includes, as its transmission section, a torque converter 20, an overdrive mechanism 40, and an underdrive mechanism 60 with three forward speeds and one reverse speed. The torque converter 20 is of a known type and includes a pump 21, a turbine 22, a stator 23, and a lock-up clutch 24. pump 2
1 is connected to the crankshaft 10 of the engine 1,
The turbine 22 is connected to a carrier 41 of a planetary gear system in an overdrive mechanism 40 via a turbine shaft 22A. In the overdrive mechanism 40, a planetary pinion 42 rotatably supported by the carrier 41 meshes with a sun gear 43 and a ring gear 44. Further, a clutch C ₀ and a one-way clutch F ₀ are provided between the sun gear 43 and the carrier 41, and a brake B ₀ is provided between the sun gear 43 and the housing Hu. The underdrive mechanism 60 is provided with two rows of planetary gears, one on the front side and the other on the gear side. This planetary gear device consists of a common sun gear 61, ring gears 62, 63, planetary pinions 64, 65, and carriers 66, 67, respectively. Ring gear 44 of overdrive mechanism 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 sun 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 an output shaft 70. On the other hand, a brake B ₃ and a one-way clutch F ₂ are provided between the carrier 67 and the housing Hu, and the sun gear 61 and the housing
A brake _B2 is provided between the sun gear 61 and the housing Hu via a one-way clutch _F1 , and a brake _B1 is provided between the sun gear 61 and the housing Hu. This automatic transmission is equipped with a transmission section as described above, and includes a throttle sensor 80 that detects the throttle opening that reflects the load condition of the engine 1, and a throttle sensor 80 that detects the output shaft rotational speed (vehicle speed) of the automatic transmission. A computer (ECU) 84 receives signals from the vehicle speed sensor 82 that detects the rotational speed of the clutch C0 and the _C0 speed sensor 99 that detects the rotational speed of the clutch _C0 .
According to the preset shift map, the electromagnetic solenoid valves S ₁ to S ₂ (for shift valves), SL (for lock-up clutch), and electromagnetic proportional valve S _D (for line hydraulic control) in the hydraulic control circuit 86 operate. The transmission is driven and controlled, and the combination of engagement of each clutch, brake, etc. as shown in FIG. 3 is performed to perform speed change control. In FIG. 3, the ◯ mark indicates the engaged state, and the ◎ mark indicates that the engaged state occurs only during driving. As shown in FIG. 4, the electromagnetic solenoid valve _S1 controls the 2-3 shift valve, and the electromagnetic solenoid valve _S2 controls the 1-2 shift valve and the 3-4 shift valve. And 1-
The shift valves 2 and 2-3 control the underdrive mechanism 60 from the first gear to the third gear, and the shift valve 3-4 controls the gear shift of the overdrive mechanism 40 (third gear). (speed change between the gear position and the fourth gear position) is performed. The electromagnetic solenoid valve SL controls the lock-up clutch 24 in the torque converter 20 via a lock-up relay valve, and the electromagnetic proportional valve _SD controls the hydraulic control device 86 via a primary regulator valve. It is designed to control the line hydraulic pressure within each line (described later). In Fig. 2, reference numeral 90 is a shift position sensor that detects the positions of N (neutral), D (drive), R (reverse), etc. operated by the driver, and 92 is a pattern select switch. , E (economic driving), P (power driving), etc.; 94 is a water temperature sensor that detects the engine cooling water temperature; 96,
Reference numeral 98 indicates a brake switch that detects the operation of the foot brake and hand brake. FIG. 5 shows the main parts of the hydraulic control device 86. In the figure, S _D is the electromagnetic proportional valve, 102 is the pump, 103 is the primary regulator valve, 104 is the 1-2 shift valve, S ₂ is the electromagnetic solenoid valve, and 106 is the electromagnetic solenoid valve operated by the driver. manual valve, 107 is the brake
The accumulators used to control the transient characteristics when hydraulic pressure is supplied to and discharged from _B2 are shown. The electromagnetic proportional valve S _D is itself well known and is composed of spools 109, 110, a coil 108, a spring 113, a plunger 111, and the like. The spool 110 and the plunger 111 are engaged with each other so that they can move integrally in the axial direction. coil 1
08 exerts a force F _C in the downward direction in the figure on the plunger 111 and therefore on the spool 110 in accordance with the load current I _P from the ECU 84 . On the other hand, spring 11
3 exerts a force F _S on the spool 110 in the opposite direction. Further, the discharge pressure of the pump 102 acts on the port 114. The hydraulic pressure acting on ports 115 and 116 is Pθ, land 10 of spool 109
Assuming that the face area of 9A is _A1 , Pθ can be found by equation (1). Pθ=(F _S −F _C )/A ₁ (1) Therefore, by controlling the force F _C generated by the coil 108 in the downward direction in the figure, the port 11
5 can be controlled to any value between O and F _S /A ₁ . Conventionally, this oil pressure Pθ corresponds to the throttle pressure generated by a throttle valve whose spool can be mechanically driven via a cam in response to the throttle opening.
Port 1 is used as a hydraulic pressure for controlling the line hydraulic pressure generated by the primary regulator valve 103.
19. 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 in the conventional art. As a result, by controlling the load current I _P to the coil 108 depending on the rotation speed of the clutch C ₀ according to the command from the ECU 84, the rotation speed of the clutch C ₀ , especially qualitative changes in the rotation speed can be controlled. Reflected line oil pressure
This means that PL can be controlled arbitrarily. Note that the pressure regulation relational expression for the primary regulator valve 103 is shown in equation (2). PL={F _S2 + (B ₂ − B ₃ )P _R +B ₂ Pθ}/B ₁ …(2) Here, F _S2 is the acting force of the spring 120, B ₁
~ _B3 is land 121 of spools 123 and 124,
The face area is 122,125. Also, P _R
is the line oil pressure applied to lands 122 and 125 when manual valve 106 is in reverse range. Next, the relationship between the frictional engagement devices will be explained. Here, brake _B2 will be explained as a representative example. The signal pressure of the electromagnetic solenoid valve _S2 acts on the port 126 of the 1-2 shift valve 104. Therefore, the spool 127 of the 1-2 shift valve 104 is in the ON- position of the electromagnetic solenoid valve _S2 .
It slides from right to left in the figure depending on the OFF status. The sliding to the right is due to the force F _S3 of the spring 128. At this time, ports 133 and 129 of the 1-2 shift valve 104 are connected. Line oil pressure from port 130 of manual valve 106 is connected to port 129.
PL is designed to work in the D (drive) range. That is, at the D range selection position of the spool 131 of the manual valve 106, the ports 130,1
29,133 are connected. On the other hand, port 133 is connected to oil passage 135 and check valve 1.
34 to the brake _B2 . Therefore, in the D range, the electromagnetic solenoid valve _S2
Line oil pressure PL to brake B ₂ by ON-OFF
is supplied and discharged. An accumulator 107 is connected to the oil passage 135, and controls the transient oil pressure level when the line oil pressure PL is supplied to and discharged from the brake _B2 . The hydraulic pressure P _B2 during operation of the accumulator 107 is determined depending on the line hydraulic pressure PL applied as back pressure, as shown by the following equation. P _B2 = F _S4 + (C ₁ − C ₂ ) PL/C ₁ …(3) Here, F _S4 is the acting force of the spring 136, C ₁ ,
C ₂ is the face area of the two lands of the accumulator piston 137. From the above equations (1) to (3), the control oil pressure Pθ is determined by the electromagnetic proportional valve.
By controlling the load current to S _D , the hydraulic pressure P _B2 to the brake B ₂ can be arbitrarily controlled while reflecting the rotational speed of the clutch C ₀ . FIG. 6 shows the control flow. First, in step 200, the vehicle speed (rotational speed of the output shaft of the automatic transmission) N ₀ , throttle opening θ, and rotational speed N _C0 of the clutch C ₀ are read. Step 202
F is a flag for program control. Initially, it is set to zero, so the step
Proceed to 204. In step 204, a gear change judgment is made according to the vehicle speed N ₀ and the throttle opening θ. When it is determined that there is no gear shift, the process proceeds to step 206 to determine the load current I _P4 to the electromagnetic proportional valve S _D determined from the current gear position, throttle opening θ, and vehicle speed.
After that, output this _IP4 and reset it. On the other hand, when it is determined in step 204 that a gear change is to be determined, the process proceeds to step 208 to output the gear change, and in step 210, the load current I _P1 determined from the type of gear change and the throttle opening θ is determined. Thereafter, in step 212, the clutch is
Rotational speed N _C0 of C ₀ becomes vehicle speed N ₀ Low gear stage gear ratio i _L
The start of the inertia phase is determined based on whether the value becomes smaller than the value multiplied by . If NO, step
Step after setting flag F to 1 in 213
Proceed to step 222 to output the load current I _P1 , and step 223
Return to step 210 via . On the other hand, when a YES determination is made in step 212, that is, when the inertia phase has started, the process proceeds to step ₂₁₄ , where the load current _is Confirm I _P2 . Thereafter, the process proceeds to step 216 to detect a predetermined time during the inertia phase. Here, the predetermined time is defined as the point in time when the rotation speed N _C0 of the clutch C ₀ becomes smaller than the product of the output shaft rotation speed N ₀ multiplied by the high gear gear ratio i _H plus the constant N ₁ . . Here, the constant
_N1 is a constant determined by the type of speed change and the throttle opening θ. Until this relationship is established, the flag F is set to 2 in step 225, and the process proceeds to step 222 to output the load current I _P2 .
The process returns to step 214 via steps 223 and 227. When the predetermined time has come, _the process proceeds to step 218, where the load current I _P3 is determined from the type of gear change, the throttle opening θ, and the rotational speed N _C0 of the clutch C0. Thereafter, in step 220, the end of the synergistic phase is detected based on whether the rotational speed _NCO of the clutch _C0 becomes smaller than the output shaft rotational speed _N0 multiplied by the high gear ratio _iH . Until the end is detected, the flag F is set to 3 in step 229, and then the load current I _P3 is outputted in step 222. When the end is detected, flag F is set in step 230.
is set to zero, and from the next flow onwards, in step 222, the load current I _P4 in step 206 is output. In addition, in this flow, shifting from 1st speed to 2nd speed and shifting from 2nd speed to 3rd speed are shown as representative examples, but in the case of a shift from 3rd speed to 4th speed, clutch at the end of the inertia phase.
The rotational speed of C ₀ N _C0 is zero, so step 216
And the gear ratio i _H of 220 is zero only in this case. An example of a map for determining the load current I _P4 in step 206 is shown in FIG. Here, an example of the first gear is shown as a representative. As is clear from the figure, as the output shaft rotational speed N ₀ increases, the load current I _P4 is set higher,
Also, the smaller the throttle opening, the lower the load current I _P4
is set high. In the electromagnetic proportional valve S _D , the higher the load current I _P4 , the lower the line oil pressure is regulated, so the higher the output shaft rotational speed N ₀ is, the smaller the throttle opening θ is. The oil pressure will be set lower. FIG. 8 shows an example of a map used to determine the load current I _P1 in step 210. As is clear from the figure, the larger the throttle opening, the more
Also, the lower the speed change process, the lower the load current.
I _P1 is set small and therefore the line oil pressure is set high. On the other hand, the load current I _P2 up to a predetermined time after the start of the inertia phase may be determined by the load current I _P1 +α(N _Cp1 /N _C0 −1)·I _P1 until the start of the inertia phase. Also, the load current I _P3 from the specified time of the inertia phase to the end of the inertia phase is I _P2 + α (N _C1 /N _C0
−N _C01 /N _C02 )・_IP2 . Here, N _C01 is N _C0 at the start of the inertia phase,
N _C02 is N _C0 at a predetermined time. Therefore,
N _C01 /N _C0 represents the magnitude of N _C0 at the present time compared to N _C01 at the start of the inertia phase. Note that this N _C01 /N _C0 becomes N _C0 /{ρ ₀ / when shifting from 3rd speed to 4th speed.
(1+ρ ₀ )N _CO1 }. Here, ρ ₀ is the gear ratio. Further, the above α is an N _C0 correction strength coefficient, and is determined depending on the type of shift, the throttle opening θ, and the time during the inertia phase. An example of this is the 9th
As shown in the figure. FIG. 10 shows the speed change transient characteristics in the above embodiment. When the shift output is output at time a, the line oil pressure P _L is changed by the load current I _P1 . The inertia phase starts from time b, and when the start of the inertia phase is detected at time c, the load current I _P2 causes the clutch C ₀
Change the line oil pressure P _L depending on the rotational speed N _C0 . When N _C0 ≦N ₀ ×i _H +N ₁ is satisfied at time d, the line oil pressure P _L is further changed depending on the rotational speed N _C0 of the clutch C ₀ in relation to the load current I _P3 . Time e
When N _C0 ≦N ₀ ×i _H is established and this is recognized at time f, the line oil pressure P _L is changed in relation to the load current i _P4 . Since the line oil pressure P _L is the back pressure of the accumulator, the clutch oil pressure changes as shown in the figure, and good shifting characteristics can be obtained. That is, clutch
Since the oil pressure decreases depending on the rotational speed N _C0 of C ₀ , the variation in the torque T ₀ from the line connecting the output shaft torques before and after the shift (dotted chain line) becomes small. In addition, in the latter half of the inertia phase, the friction coefficient of the friction engagement device generally increases as it changes from the dynamic friction coefficient to the static friction coefficient, resulting in an overshoot of the output shaft torque _T0 , but here, the rate of decrease in oil pressure Since ΔT is increased, ΔT in the figure decreases, and very good characteristics can be obtained. As shown here, it is desirable that the clutch oil pressure decreases toward the end of the shift.
This is because the output shaft torque in the inertia phase is determined by the oil pressure, so by lowering the oil pressure toward the end of the shift and gradually reducing the output shaft torque in the inertia phase, the output shaft torque at the end of the inertia phase is reduced. This is because the sudden change in the output shaft torque is reduced and good characteristics as shown in the figure can be obtained. In addition, as is clear from the figure, in this embodiment, the line oil pressure is relatively low when not shifting to reduce wasteful loss of power, and when shifting, the line oil pressure is reduced only when necessary. By increasing the line oil pressure, the above-mentioned good characteristics are obtained.

【Effect of the invention】

As explained above, according to the present invention, the control hydraulic pressure during gear shifting can be controlled in a preferable direction with a simple configuration, and the durability of the friction engagement device can be ensured.
An excellent effect can be obtained in that it becomes possible to simultaneously reduce the shift shock.

[Brief explanation of the drawing]

FIG. 1 is a block diagram showing the gist of the present invention, and FIG. 2 is an overall skeleton diagram of an automatic transmission for a vehicle to which an embodiment of the hydraulic control device during upshifting of an automatic transmission according to the present invention is applied. FIG. 3 is a diagram showing the operating state of the frictional engagement device in the automatic transmission;
FIG. 4 is a diagram showing the input/output relationship of the control system, FIG. 5 is a hydraulic circuit diagram of the main parts of the hydraulic control system, FIG. 6 is a flowchart showing the control routine, and FIGS. Fig. 8 is a diagram showing an example of the map of I _P4 and I _P1 , respectively, Fig. 9 is a diagram showing an example of the map of N _C0 correction intensity coefficient α, and Fig. 10 is a diagram showing the effect of the above embodiment. FIG. 2 is a qualitative diagram of speed change transient characteristics. 1...Engine, 20...Torque converter, 40
...overdrive mechanism, 60...underdrive mechanism, 84...ECU, 99...rotational speed sensor of clutch _C0 , 86...hydraulic control circuit, S _D ...electromagnetic proportional valve, 103...primary regulator valve, 1
07...Accumulator, 140...Duty control valve.

Claims (1)

[Scope of Claims] 1. A means for detecting the rotational speed of either the engine or a rotating member of the automatic transmission; a means for arbitrarily regulating the control hydraulic pressure in the hydraulic control device of the automatic transmission; and during upshifting. setting in real time a pressure regulation target value of the control oil pressure that changes from moment to moment in the hydraulic control device in dependence on a rotation speed that changes from moment to moment of at least one of the rotating members of the engine or the automatic transmission; An automatic transmission characterized by comprising: means for changing the degree of dependence of the rotational speed stepwise or continuously during shifting, and creating a characteristic of gradually decreasing the control hydraulic pressure immediately before engagement. Hydraulic control device during upshift. 2. The hydraulic control device for upshifting an automatic transmission according to claim 1, wherein the degree of dependence of the rotational speed is changed depending on at least one of the type of shift, engine load, and vehicle speed. 3. A hydraulic control device for upshifting an automatic transmission according to claim 1 or 2, wherein the control hydraulic pressure is line pressure. 4. A hydraulic control device for upshifting an automatic transmission according to claim 1 or 2, wherein the control hydraulic pressure is a hydraulic pressure for line hydraulic control. 5. A hydraulic control device for upshifting an automatic transmission according to claim 1 or 2, wherein the control hydraulic pressure is a hydraulic pressure applied to a back pressure chamber of an accumulator. 6. Claim 1 or 2, wherein the control hydraulic pressure is the hydraulic pressure in the oil passage immediately before the frictional engagement device.
Hydraulic control device during upshift of automatic transmission described in 2.