JPS6184419A - Hydraulic control device - Google Patents

Abstract

PURPOSE:To optionally adjust the hydraulic pressure of oil fed to a load, by providing a relief valve whose valve opening pressure set by a spring is changed in association with the displacement of a piston, and an electrically driven type flow control valve disposed in a hydraulic pressure supply passage, for controlling the above-mentioned piston. CONSTITUTION:In the case of application of a hydraulic control valve for controlling the hydraulic pressure control of a speed-change clutch in a transmission unit, a variable relief valve 1 and a variable restrictor valve 30 positioned above the former, are disposed in a valve body 2. The relief valve 1 is composed of a spool 3 having piston sections 4, 5 and a piston 8 positioned on the line coaxial with the spool 3, and oil chambers 6-10 are defined in the inside of a cylinder bore. The spring force of spring 19 disposed between the spool 3 and the piston 8 may be adjusted in association with the displacement of the piston 8. The variable restrictor valve 30 is controlled by an electrically driven actuator 35 in accordance with the operating condition of a vehicle, thereby it is possible to control the hydraulic pressure effected upon the piston 8 so that the hydraulic pressure fed to clutch cylinders 22A, 22B is changed.

Description

[Detailed Description of the Invention] [Technical Field of the Invention] The present invention relates to a hydraulic control system for a gear shifting clutch, etc. that operates a transmission shifter to shift gears! This invention relates to a hydraulic control device suitable for employing nK.

[Conventional technology]

As is well known, a power shift transmission using a planetary gear device is configured to shift gears by selectively operating a gear shifting clutch provided at each gear stage. In the transmission, when the hydraulic oil from the hydraulic pump is applied directly to the transmission clutch,
Because the oil pressure in the cylinder of the above clutch rises rapidly,
This causes what is called a gear shift lock. Conventionally, therefore, a so-called modulating pulp, which acts to gradually increase the oil pressure in the cylinder of the clutch, has been used to alleviate the shift stagnation.

[Problems that the invention is trying to solve] However, the above-mentioned conventional modulating pulp can only obtain uniform hydraulic control characteristics determined by the constant of the built-in spring, and for this reason, it is difficult to control the hydraulic pressure due to load fluctuations during gear changes. When the vehicle width is suddenly accelerated or suddenly reduced, it is not possible to sufficiently suppress the shift shift.

[Means for Solving the Problems J] In view of these conventional problems, the present invention provides a method for generating relief pressure that is balanced with the biasing force of the spring when the length of the spring is changed due to the displacement of the piston. I am using a Susu IJ-7 valve. An electric flow rate adjustment means is interposed in the supply path of the hydraulic pressure applied to the piston, and the relief pressure applied to the load is changed by the flow rate adjustment means.

Further, in the present invention, a fixed orifice is added to the above-described structure that functions to reduce the hydraulic pressure applied to the piston.

[Function] According to the present invention, the electric flow rate v! It is possible to arbitrarily change the relief pressure applied to the load by controlling the adjustment means with an electric signal.

〔Example〕

Hereinafter, the present invention will be described in detail with reference to the drawings.The first example is a system in which the oil 8E control #1JK was applied to a speed change clutch of a planetary gear type power shift transmission 1 mounted on a construction vehicle. 1 shows an embodiment of a hydraulic control device according to the present invention. The hydraulic control device according to this embodiment is equipped with a variable relief valve indicated by the reference numeral 1 in the figure. The spool #3 disposed within the pulp body 2 of this relief valve 1 is provided with first and second piston parts 4 and 5 having the same diameter, and these piston parts 4.5 open an oil chamber 6. is defined.

To the left of the oil chamber 6, an oil chamber 7 is formed by the left end surface of the first piston part 4 and the inner surface of the pulp body 2, and to the right of the oil chamber 6, the second piston part 5 is formed. An oil chamber 9 is formed by the outer surface of the pulp body 2, the inner surface of the pulp body 2, and the left end surface of the piston 8, and an oil chamber 10 is formed on the right side of the oil chamber 9 by the right end surface of the piston 8 and the inner surface of the pulp body 2. is formed.

A boat 11 is opened in the oil chamber 6, and a boat 11 is opened in the oil chamber 6.
1 is connected to a switching valve 13 and also to a hydraulic power source 15 via a throttle 14 .

A spring 16 is built into the oil chamber 7 and serves as a damper when the spool 3 moves to the left.
The oil chamber 9 is formed to have a diameter larger than that of the piston portion 4.5, and is provided in the drain boat 18. A spring 19 is disposed within the oil chamber 9 so as to be interposed between the right end surface of the second piston portion 5 and the left end surface of the piston 8. Note that this oil chamber 9 functions as a relief passage for the pressure oil in the oil chamber 6, as will be described later. Two switching valves are connected to the pressure.

The clutch cylinder 22At22B operates a clutch for changing speeds for each speed stage in an idler gear type power shift transmission (2)/ (not shown), and is selectively driven via the switching valve 13.

A pilot pressure input port of the switching valve 13 is connected to the hydraulic pressure source (hydraulic pump) 15 via a throttle 23 and to a solenoid valve 24 .

On the inner peripheral surface of the valve body 2 where the oil pressure inside the chamber 9 is located, there is a stopper 25 for regulating the leftward movement position of the histones 8.
A relief valve 26 is connected between the throttle 14 and maE #15. Note that this relief valve 26 reduces the output pressure of the hydraulic power source 14 to, for example, 1
It functions to set rKqf/- or 35Kif/-.

A variable throttle 30 is formed above the IJ valve 1 . This variable aperture O is the first. It includes a spool 34 with second and third piston parts 31, 32 and 33, and a direct-acting electric actuator 35 that displaces the spool 34.

Above 1. The second piston part 31.32 has an oil chamber 36,
Also second. The third piston parts 32 and 33 each define an oil chamber 37. In the oil chamber 36, the IJ
A boat 38 communicating with the oil chamber 6 of the 17-f valve 1 is formed, and a boat 39 communicating with the oil chamber 10 of the facing valve 1 is formed in the oil chamber 37.

The second piston portion 32 of the spool 34 has four grooves 40 formed on its circumferential surface extending along the axial direction as shown in FIG. It is formed so that the width and clearance gradually decrease from the left end side to the right end side, and therefore, as the piston portion 32 is displaced, the hydraulic oil between the oil chamber 36 and the oil chamber 37 is The electric actuator 35 has a mechanism for displacing the rod 41 in the left and right direction using a built-in servo motor.
The rod 41 is connected to the spool 34.

As shown in the figure, assuming that hydraulic oil is supplied to the clutch cylinder 22A and the shift clutch for the cylinder 22A is in the engaged state, the hydraulic pressure acting on the clutch cylinder 22AK, that is, the oil pressure in the oil chamber 6 The oil pressure indicates the maximum oil pressure P1 as shown in FIG. The piston 8 is at the maximum oil pressure generation position shown in FIG. and the anti-rupture force of the spring 19 are in balance.

Here, when the electromagnetic 9P24 is switched by a shift command signal from a shift type M helmet path (not shown), the pilot pressure to the switching valve 13 is reduced, so the switching valve 13 is switched. As a result, the pressure oil in the cylinder 22A is drained, and the clutch for the cylinder 22A becomes disengaged.

On the other hand, the switching operation of the switching valve 13 connects the boat 11 of the oil chamber 6 to the clutch cylinder 22B, thereby starting the supply of hydraulic oil to the cylinder 22B. At this time, the oil chamber 6 in the boat 11
As shown in FIG. 3, the oil pressure suddenly drops to a value P2 close to zero, but this is due to the sudden flow of hydraulic oil into the cylinder 22B.

On the other hand, when the oil pressure of the boat 11 suddenly decreases, the switching valve 2
As a result, the pressure oil in the oil chamber IO is quickly drained through the valve 22, and the piston 8 is moved to the minimum oil pressure generation position C (stroke end), for example, as shown in FIG. It is returned to the position where it contacts the right inner iK of the body 2. When the piston 8 is returned to the same position, the spring 19 has its natural length. When the hydraulic oil flows into the cylinder 22B and the cylinder is filled with hydraulic oil, as shown in FIG. When the so-called filling time ends, the oil pressure in the oil chamber 6 starts to rise, thereby returning the switching valve 22 to the state shown in FIG. 1. The oil chamber 10 is connected to the oil chamber 6 via a variable throttle of 300 m40, and therefore, at the same time as the valve 21 is set to its original position, the oil pressure in the core oil chamber 10 also starts to rise, and along with this, The piston 8 moves to the left. On the other hand, since the oil pressure in the oil chamber 6 is equal to the oil pressure in the oil chamber 7, when the oil pressure starts to rise, the oil pressure in the oil chamber 7 causes the spool 3 to resist the droplet 19. and go to the right.

When the spool 3 moves a certain distance to the right, the pressure oil in the oil chamber 6 is relieved into the oil chamber 9 as shown in FIG. In other words, the spool 3 operates so that its rightward force and the reverse force of the spring 220 are balanced. As a result, the oil pressure in the oil chamber 6, that is, the variable IJ IJ
- The relief pressure of ffl is increased, for example, in the manner shown in characteristic A in FIG.
In order to change the increase in IJ-F pressure in the oil chamber 6, which indicates the clutch pressure, it is sufficient to change the piston position 80 and reduce the pressing force of the spring 20 on the three sides 101 on the spool 3. , it has oil flowing into it [10! I
All you have to do is change the oil flow rate.

Since the maximum possible aperture 30 shown in FIG.
By displacing the spool 34, the flow rate of the hydraulic oil flowing into the oil chamber 10&C from the oil amount 6 can be changed. Therefore, by operating the 'dE dynamic actuator 35, the relief pressure (clutch pressure) can be increased. OT function is to arbitrarily change the rise time of .

Incidentally, the degree of shift shift (change in transmitted torque) is mainly determined by the parameters listed below.

a. Gear shift stage, running load (steady load) C0 vehicle body slack (inertia load) d. Throttle amount e, rate of increase in clutch oil pressure (rate of increase in transmitted torque) Now, if the heat capacity of the clutch plate is extremely large, it will be counter-productive. In this case, K, whose gear shift 71 always remains below a certain level, can set the clutch oil pressure increase rate shown in e by taking into account the above parameters a-%-d.In other words, for ease of explanation, for example, Considering the case where only the vehicle weight changes, as shown in Figure 3 (when the vehicle weight is large, the increase rate of the clutch pressure with respect to time is set to be large, and when the vehicle weight is small, the increase rate is set to be small). Therefore, it is possible to reduce the shifting effort.

When creating an actual control algorithm based on the above concept, there are the following five problems.

a. There is no proportional relationship between clutch transmission torque and clutch oil pressure.

b. It is difficult to measure the running load value on the actual machine.

Problem a arises because the coefficient of friction of the clutch plate changes depending on the sliding speed. For example, in the wet type clutches used in construction machinery, even if the oil pressure is increased, the transmitted torque remains the same as shown in Figure 6. It changes like this.

As shown in the figure, the rate of change of the transmitted torque has a peak at a certain point, and this peak provides a fairly large shift shift. Since the peak of the transmitted torque occurs just before the clutch is completely engaged, this peak can be suppressed by reducing the rate of increase in clutch pressure just before the clutch is completely engaged.

The above process can be implemented as follows.

In other words, as the clutch pressure increases, the rotational speed of the input shaft and the rotational speed of the output shaft of the transmission become close (as shown in Fig. 7), and when the clutch is fully engaged, both rotational speeds match. (Actually, the ratio of both rotational speeds, that is, the gear ratio, becomes a predetermined value.) Therefore, at the time when the clutch is fully engaged and the time immediately before the counterfeit time, the rotational speed of the input and output shafts is You can find out by observing.

Therefore, the rate of increase in clutch pressure is reduced (for example, the clutch pressure is kept constant) just before the clutch is fully engaged based on the above input/output ++ib rotation speed, and the clutch is re-clutched when the clutch is fully engaged. By increasing the pressure, the peak of the transmitted torque can be suppressed and the shift shift can be further reduced.Next, the above-mentioned problem will be considered. In the clutch pressure increase characteristic illustrated in FIG. 7, a question arises as to whether the initial clutch pressure increase rate cannot be determined if the running load is unknown. However, the above-mentioned initial clutch pressure increase rate can be almost determined by the gear position, vehicle weight, and throttle amount. This is because the running load is involved in the point at which the peak of the transmitted torque occurs, as shown in Figure 8. Based on the above considerations, this embodiment causes the controller 42 shown in FIG. 1 to execute the process shown in FIG. 9. ing. First, based on the outputs of the gear position sensor 45 (which detects the position of the gear lever, for example) and the throttle amount sensor 46 and the single fiber sensor 47, the gear position and the throttle amount (
Throttle opening) and vehicle weight are detected, and (Step 1
00) Next, it is determined whether a shift command signal is output from a shift determination circuit (not shown) (step 101). Then, the judgment result in step 101 is YES.
Then, an optimal latch pressure opening command is searched for depending on the axle condition, that is, the gear position, throttle amount, and vehicle weight obtained by each of the above-mentioned sensors (step 102).

The optimal clutch pressure increase rate according to the condition of the axle can be known in advance from the actual clutch pressure, etc., and this increase rate is realized by controlling the rolling actuator 35 as described above.

A memory (not shown) included in the controller 42 stores in advance a table of latch pressure commands for the actuator 35 for applying a clutch pressure increase rate adapted to various conditions of the axle. The FJq search in step 102 means reading out the control command from the memory based on the output of each sensor.@The clutch pressure control command retrieved in step 102 is 35 to the electric actuator 35 (step 103), thereby displacing the spool 40 of the variable diaphragm 30 to a position corresponding to the command. In addition, this spool 34
The displacement is carried out during the filling time shown in FIG. Further, the displacement position of the spool 34 is detected by position detection means such as a potentiometer built into the t-actuator 35.

When the filling time ends, the clutch pressure increases according to a predetermined rate of increase, for example in the manner shown in FIG. 7. Therefore, the controller 42 determines whether the clutch slippage has become below a set value based on the outputs of the rotation sensors 43 and 44, which detect the rotational speed of the input shaft and output shaft of the transmixer 1, respectively (step 104). When the slippage of the clutch becomes less than the set value, that is, when it is determined that the clutch is about to engage, a command is given to the actuator 35 to keep the clutch pressure constant (step 105). is moved to the left, the oil chamber 6 and the oil chamber 10 are cut off, and the piston 8
displacement is stopped.

As a result, the clutch pressure (hydraulic pressure in the oil chamber 6) is maintained constant as shown in FIG.
The process continues until it is determined that the clutch is fully engaged, that is, until the determination result at the next step 106 becomes YFiS.

When the determination result in step 106 is YFjS, a command is issued to increase the clutch pressure again, as shown in FIG. is applied to the actuator 35 (step 108), which causes the piston 8 to operate again and the clutch pressure to rise. When the piston 8 moves to the left until it contacts the stopper 25, the clutch pressure shown in the figure becomes constant.

Incidentally, in the above embodiment, it is not possible to reduce the clutch pressure over time, but if the oil chamber 10 is connected to the drain tank via the fixed throttle 48, as shown by the dotted line in FIG. It is also possible to reduce the pressure.

That is, the flow rate q1 of the hydraulic oil now flowing into the oil chamber 10 from the oil chamber 6 via the variable throttle 30 is the fixed throttle 48
The flow of hydraulic oil drained from the oil chamber 10 through ff1Q
2, the clutch pressure changes in the following manner depending on the magnitude relationship of these flow rates, and FIG. 10 shows this in a graph.

Q, 1>Q2 Clutch pressure monotonous increase Q1=Q2
Clutch pressure is constant Ql<Q2 Clutch pressure monotonically decreases, and in this case, in step 105 of the same figure, the clutch pressure is changed to decrease, that is, the increasing gradient Δp of clutch pressure
It is also possible to make /Δt negative.

The fixed running throttle 48 is formed so that the flow rate of the hydraulic oil drained therethrough is smaller than the maximum flow rate of the hydraulic oil flowing into the oil chamber 10 via the variable throttle 30.

FIG. 11 shows an uninvented embodiment in which a solenoid valve 49 is used in place of the front 6-pin valve 30. This implementation? lI
In this case, the solenoid valve 49 is interposed between one port 20 of the oil chamber 10 and the boat 38 of the oil chamber 6 via a throttle 50, and the switching switch yP22 is connected to the other port 20 of the oil chamber 10. ing.

In this embodiment, if the RA electromagnetic valve 9 is driven by PWM (Pulse Width Modulation) No. 19, a desired clutch pressure increase rate can be obtained by the twelfth movement. That is, by making the duty ratio large, medium, or small, the latch pressure increasing characteristic can be improved as illustrated in FIG. 13. Of course, in this example, it is also possible to connect the clutch to the drain tank via the fixed throttle by turning off the supply of oil to the oil chamber 10 by the solenoid valve 49. Note that this embodiment can also be controlled according to the example shown in FIG. 8, and in this case, the controller 42 has a built-in PWM drive circuit.

In the embodiment shown in FIG. 1 and FIG. 1/FIG. 1, hydraulic oil is supplied from the oil chamber 6 to the oil chamber 10 via the variable throttle 30 and the solenoid valve 49, respectively. 5K may be used to supply hydraulic oil from the hydraulic pressure rM15 into the oil chamber IO. However, in this case, the linearity of the clutch pressure rise characteristic is slightly lost compared to the former case.

Furthermore, in the embodiment shown in FIG. 1, the electric actuator 3
Although a servo motor is used as the power source for the drive unit 5, it is of course possible to use other drive sources such as a step motor or the like instead.

〔Effect of the invention〕

According to the present invention, the hydraulic pressure applied to the load can be arbitrarily controlled by an electric signal. Therefore, for example, by applying the present invention to the clutch pressure control of a power shift transmission as shown in the first embodiment, it is possible to control the hydraulic pressure so as not to cause a malfunction.

[Brief explanation of drawings]

1st I8! ! , I is a block diagram schematically showing an embodiment of the hydraulic control device according to the present invention, FIG. 2 is a perspective view illustrating the shape of the groove formed in the spool of the variable throttle, and FIG. FIG. 1 is a graph illustrating the hydraulic pressure increase characteristic obtained by the embodiment shown in FIG.
Figure 6 is a graph illustrating the relationship between clutch pressure and transmitted torque, Figure 7 is a graph illustrating a bad method of hydraulic control when suppressing the peak of transmitted torque, and Figure 8 is a graph illustrating the relationship between clutch pressure and transmitted torque. Graph showing changes, No. 9
The figure is a flowchart illustrating the quick filling means of the controller when performing the control shown in Figure 7, Figure 10 is a graph illustrating how the clutch pressure increases and decreases, and Figure 11 is
12 is a waveform diagram illustrating the PWM signal waveform applied to the solenoid valve shown in FIG. 11; Fig. 13 is a graph illustrating the clutch pressure increase characteristic corresponding to the duty ratio value of PWM number 1-I.
J leaf valve, 2...pulp body, 3-spool,
4.-first piston part, 5-second piston part, 6.
7.9.10--oil chamber, 8--piston, 11,17
゜18, 20.20.38-boat, 13.21--switching valve, 14, 23.48.50--throttle, 15...hydraulic source, 16.19-spring% 22A, 22B, -clutch cylinder , 24° 49-Solenoid valve, 25... Stopper, 30- Variable throttle, 34... Spool, 35- Electric actuator, 40... Groove 1.42- Controller,
43-human power shaft rotation sensor, 44-output shaft rotation sensor, 4
5-gear stage sensor, 46-throttle amount sensor, 47.
-Vehicle weight sensor. Figure 7-B Time Figure 8 □ Time Figure 9 Figure 10 □ Time $11 Figure 1-0 Figure 12 Figure 13 Figure 1 hour

Claims (4)

[Claims] (1) Change the length of the spring by the displacement of the piston,
A relief valve that generates a relief pressure balanced with the biasing force of the spring, and an electric flow rate adjustment means interposed in a hydraulic pressure supply path to act on the piston, and apply the relief pressure to the load. Hydraulic control device. (2) The hydraulic control device according to claim (1), wherein the flow rate adjusting means is a variable throttle operated by an electric actuator. (3) The hydraulic control device according to claim (1), wherein the flow rate adjusting means is a solenoid valve. (4) Changing the length of the spring by the displacement of the piston,
a relief valve that generates a relief pressure that is balanced with the biasing force of the spring; an electric flow rate adjustment means that is interposed in a hydraulic pressure supply path that acts on the piston; and a function that reduces the hydraulic pressure that acts on the piston. A hydraulic control device is provided with a fixed orifice for supplying the relief pressure to a load.