JP2812393B2 - Line pressure control device for hydraulically operated transmission - Google Patents

Description

Description: BACKGROUND OF THE INVENTION 1. Field of the Invention The present invention relates to a line pressure control device for a hydraulically operated transmission, and more particularly, to an improvement in pulsation reduction in line pressure.

[Prior Art] For example, as a prior art relating to duty solenoid control in a hydraulic circuit of a continuously variable transmission, Japanese Patent Application Laid-Open No.
There are seven.

In this technique, a duty signal having a fixed cycle and a variable duty ratio is applied to a duty solenoid to control the pilot pressure and line pressure of a control valve.

(Problems to be Solved by the Invention) Since the pulsation of the line pressure in the hydraulic circuit also affects the pilot pressure, its reduction is emphasized. In particular, in the case of a continuously variable transmission, a highly accurate hydraulic pressure is required, and thus a need for a hydraulic circuit with less pulsation is high.

However, the inventors have found that it is difficult to reduce the pulsation when the duty control is performed by changing the duty ratio at a constant period as in the above-described related art. This means that if the period is constant, the pulsation is likely to be at the center of the duty, and the amplitude of the pulsation itself is
Although it is proportional to the discharge amount of the oil pump, it is difficult to eliminate the influence of the discharge amount by the duty control with a constant cycle.

SUMMARY OF THE INVENTION The present invention has been made in view of the problems of the related art, and has as its object to propose a line pressure control device for a hydraulically operated transmission with less pulsation.

[Means and Actions for Achieving the Object] A configuration of a line pressure control device for controlling a line pressure of a hydraulically operated transmission according to claim 1 of the present invention for achieving the above object is shown in FIG. 1A. A duty solenoid driven by a predetermined drive signal, a control valve for controlling a pilot pressure by the duty solenoid to change a discharge amount of an oil pump to regulate a line pressure of a hydraulically operated transmission, It is characterized by comprising means for determining a duty ratio of the duty solenoid according to a state, and means for changing and controlling the frequency of the drive signal based on the determined duty ratio.

According to a second aspect of the present invention, there is provided a line pressure control device for controlling a line pressure of a hydraulically operated transmission, comprising: a duty solenoid driven by a predetermined drive signal; and a pilot pressure by the duty solenoid. A control valve for controlling and changing a discharge amount of an oil pump to regulate a line pressure of a hydraulically operated transmission; a means for determining a duty ratio of the duty solenoid according to an operation state; and a frequency of the drive signal. A means for performing change control based on the determined duty ratio and the discharge amount of the oil pump is provided.

It is to be noted that the line pressure control device according to claim 3 is characterized in that the discharge amount of the oil pump according to claim 2 is determined according to the rotation speed (for example, engine rotation speed) of the input torque of the hydraulically operated transmission. And

(Embodiment) An embodiment in which the present invention is applied to a line pressure control device of a belt-type continuously variable transmission will be described in detail with reference to the accompanying drawings.

<Structure of the continuously variable transmission> First, FIG. 2 is a skeleton diagram showing the entire structure of the continuously variable transmission Z, and FIG. 3 is a diagram of the continuously variable transmission Z shown in FIG. Each of the hydraulic circuits Q is shown. The continuously variable transmission Z and the hydraulic circuit Q are described in Japanese Patent Application No.
Since it is substantially the same as that disclosed in JP-A-63-213722, a detailed description of the transmission Z and the hydraulic circuit Q will be omitted, and portions particularly related to the present invention will be described.

The continuously variable transmission Z is a continuously variable transmission for a front wheel drive vehicle, and includes a torque converter B connected to an output shaft 1 of an engine A, a forward / reverse switching mechanism C, a belt transmission mechanism D, and a reduction mechanism E. A differential mechanism F is basically provided.

The torque converter B is rotatably provided in a converter chamber 7a formed inside the pump cover 7 so as to face the pump impeller 3 and the pump impeller 3 which rotates integrally with the engine output shaft 1. And a stator 5 interposed between the pump impeller 3 and the turbine runner 4 to increase the torque. The turbine runner 4 is a carrier 15 which is an input member of a forward / reverse switching mechanism C to be described later via a turbine shaft 2, and the stator 5 is connected to a mixing case 19 via a one-way clutch 8 and a stator shaft 9. Are linked.

Further, a lock-up piston 6 is arranged between the turbine runner 4 and the pump cover 7. The lock-up piston 6 is slidably mounted on the turbine shaft 2, and is brought into contact with the pump cover 7 by the introduction or discharge of hydraulic pressure into the lock-up chamber 10 so as to be integrated with the pump cover 7. The converter state separated from the pump cover 7 is selectively realized. In the lock-up state, the engine output shaft 1 and the turbine shaft 2 are directly connected without passing through the fluid, and in the converter state, the engine torque is transmitted from the engine output shaft 1 through the fluid.
Each is transmitted to the turbine shaft 2 side.

Here, it should be noted that the torque transmitted from the engine output to the forward / reverse switching mechanism C via the torque converter B depends on the converter transmission torque transmitted via the turbine runner 4 and the lockup piston. 6 is that there is two of the lock-up clutch transmission torque transmitted through. Therefore, as described later in detail, the turbine torque transmitted to the primary shaft 22 must be determined in consideration of these two torques.

The forward / reverse switching mechanism C selectively sets a forward state in which the rotation of the turbine shaft 2 of the torque converter B is transmitted as it is to a belt transmission mechanism D described later and a reverse state in which the rotation is transmitted to the belt transmission mechanism D in a reverse rotation state. In this embodiment, the forward / reverse switching mechanism C is constituted by a double pinion type planetary gear unit, and the forward / reverse switching is executed by the selection operation of the clutch 16 and the brake 17. Things.

The belt transmission mechanism D includes a primary pulley 21 described below coaxially disposed behind the forward / reverse switching mechanism C, and a secondary pulley 31 described below disposed separately in a direction parallel to the primary pulley 21. Between belt 20
It is constructed by stretching. The primary pulley 21 has a fixed conical plate 23 having a predetermined diameter integrally formed with the primary shaft 22 on the primary shaft 22, and a movable conical plate 24 movable in the axial direction with respect to the primary shaft 22, Each is provided and configured. A cylindrical cylinder 25 is fixed to the outer surface 24a of the movable conical plate 24. Further, a piston 26 fixed to the primary shaft 22 side is oil-tightly fitted on the inner peripheral surface side of the cylinder 25, and the piston 26, the above-described cylinder 25 and the movable conical plate 24 are Who
A primary room 27 is configured. Note that line pressure is introduced into the primary chamber 27 from a hydraulic circuit Q described later. And this primary pulley 21
The effective diameter of the belt 20 is adjusted by moving the movable conical plate 24 in the axial direction by the hydraulic pressure introduced to 27 and increasing or decreasing the interval between the movable conical plate 24 and the fixed conical plate 23.

The secondary pulley 31 has a fixed conical plate 33 on the secondary shaft 32,
34 are provided so as to be movable on the secondary shaft 32. Further, a substantially cylindrical cylinder 35 is coaxially fixed to the outer surface 34b side of the movable conical plate 34. Also,
On the inner peripheral surface side of the cylinder 35, a portion near the axis is
A piston 36 fixed to the secondary shaft 32 is inserted in an oil-tight manner. This piston 36, cylinder 35 and movable cone plate
A secondary room 37 is composed of the three members 34. Line pressure is introduced into the secondary chamber 37 from the hydraulic circuit Q, similarly to the primary pulley 21 side. Similarly to the primary pulley 21, the secondary pulley 31 has its movable cone 34 moved toward and away from the fixed cone 33 so that the effective diameter of the belt 20 is adjusted.

The pressure receiving area A _S of the movable conical plate 34 on the secondary pulley side
Is set to be smaller than that of the movable conical plate 24 of the primary pulley 21 (A _P ), and approximately A _P : A _S ≒ 2: 1.
As will be described later, when the primary pulley and the secondary pulley are controlled by a common hydraulic pressure, a difference in the pressing force due to a change in the hydraulic pressure between the two pulleys occurs due to the area difference, and particularly, the centrifugal force correction is performed. In such a case, there arises a problem that the line pressure becomes insufficient, so in this embodiment, as will be described later, the minimum required line pressure is set to solve this problem.

The deceleration mechanism E and the differential mechanism F have conventionally known configurations.

The hydraulic circuit Q shown in FIG. 3 corresponds to the lock-up piston 6 of the torque converter B in the above-described continuously variable transmission Z.
An oil pump 40 provided to control the operation of the clutch 16 and the brake 17 of the forward / reverse switching mechanism C and the primary pulley 21 and the secondary pulley 31 of the belt transmission mechanism D, and driven by the engine A. ing. Here, the oil pump 40 includes a primary chamber 27, a secondary chamber 37, a clutch chamber 61, and a brake chamber 6 as control elements.
2, common (identical) supplied to lock-up device, etc.
It is defined as a pressure source that provides a control source pressure.

Since the oil pressure is commonly supplied from one oil pump 40, pulsation tends to occur in the line pressure. However, in order to suppress the pulsation, the cycle of the solenoid drive signal is made variable. . This will be described later.

The detailed configuration and operation of the hydraulic circuit Q in FIG. 3 will be described in Japanese Patent Application No. 63-213722, and the schematic configuration thereof is as follows.

The basic line pressure is controlled by a line pressure control valve 41 whose pilot pressure is controlled by a solenoid 51. Since this basic line pressure is transmitted directly to the secondary chamber 37, control of the pressure in the secondary chamber 37 requires control of this basic line pressure. The basic line pressure is also transmitted to the primary chamber 27 via the speed ratio control valve 43. The pilot pressure of the control valve 43 is controlled by a switching valve 44 controlled by a solenoid 52.

The hydraulic oil adjusted by the line pressure adjusting valve 41 is adjusted to a predetermined clutch pressure by the clutch pressure adjusting valve 46 and then introduced into the lock-up control valve 47 via the line 109.
The hydraulic oil introduced into the lock-up control valve 47 is selectively supplied to the lock-up engagement side (LOCK) or the lock-up release side (UNLOCK) by controlling the pilot pressure by the solenoid valve 53.

<Overall Line Pressure Control> FIG. 4 shows signals input to and output from the control unit. Each solenoid valve 51, 52, 53 has
As shown in FIG. 4, a control unit 78 is connected, and each electromagnetic solenoid valve 51, 52, 53 is connected to the control unit 78.
Is driven to control the control element. In FIG. 6, the control unit 78 includes a shift position signal RANGE from a sensor 82 for detecting a shift position (D, 1, 2, R, N, P) operated by the driver, and a primary shaft 22.
Rotational speed N and the primary pulley speed signal N _P from the rotational speed sensor 83 for detecting a _P, secondary pulley speed signal N from speed sensor 84 which detects the rotational speed N _S of the secondary shaft 32 (or vehicle speed) of _S and throttle opening of engine A
Throttle opening signal T from opening sensor 85 that detects TVO
And VO, turbine speed from the turbine speed sensor 87 for detecting a rotational speed signal N _E from the rotational speed sensor 86 for detecting an engine engine speed N _E, the rotational speed N _T of the turbine shaft 2 in the torque converter B The signal _NT and the oil temperature THO from the sensor 88 that detects the oil temperature of the hydraulic circuit Q are input.

1B and 1C are general block diagrams of the line pressure control according to the present embodiment. As shown in these figures,
This control outputs the following three duty signals. These signals include a signal TLUP to the duty solenoid 53 for controlling the hydraulic pressure acting on the lockup piston 6, that is, a pilot pressure of the control valve 41, and a hydraulic pressure acting on the secondary chamber 37. That is, the signal TSEC to the duty solenoid 51 for controlling the basic line pressure and the oil pressure acting on the primary chamber 27 are controlled, that is, the control valve 44
Duty solenoid for controlling pilot pressure in air
Signal TPRM to 52.

Lock-Up Solenoid Control According to FIG. 1B, the generation control of the signal TLUP is as follows. That is, if it is determined by 200 that the current operation state is within the lock range, a differential pressure signal DPLUP is calculated at 201. FIG. 5A shows the map characteristics for determining the lock range. That is, when it is determined by the logic shown in FIG. 5A that it is within the lock-up range, 201 indicates the lock-up clutch differential pressure DPLUP from the lock-up transmission torque initial value TQINT calculated from the engine output torque TQPUMP (described later).
Is calculated. This DPLUP is a target value of the hydraulic pressure difference before and after the piston 6. That is, this hydraulic pressure difference DPLUP
Is the hydraulic pressure difference between LOCK64 and UNLOCK65 (see FIG. 3).

The DPLUP is subjected to a limit correction in 202 based on the characteristics as shown in FIG. 5A. This is because excessive hydraulic pressure is not applied.

And further, the clipped DPLUP, at 203,
It is corrected taking into account the line pressure (this is defined by a signal PSEC defining the hydraulic pressure applied to the secondary chamber 37). That is, according to (c) of FIG. 5B, the maximum value of the clutch pressure is calculated by the line pressure PSEC, and this calculated amount is set as the maximum clutch pressure PLUOFF _MAX . Then, as shown in FIG. 5B (b), DPLUP is converted into a clutch pressure PLUOFF based on the characteristics of FIG. 5B (b), while limiting the DPLUP with this PLUOFF _MAX as the maximum value. Clutch pressure PLUO
The FF is converted into a duty ratio at 204, and power supply voltage correction and the like are performed at 205. Further, at 206, the corrected duty is converted to a period and the solenoid 53
Is output to The details of the duty conversion will be described later.

Control of Primary Chamber Hydraulic With reference to FIG. 1C, an outline of hydraulic control of the primary chamber will be described. First, in 220, the shift signal RANGE, opening TVO of Surotsutoru, operation mode MODE, based on the rotation speed N _S, etc. of the current secondary pulley, reads the target value PREVT the primary rotation speed from Matsupu. FIG. 6 shows the characteristics of this map. Then, the difference DNP between the rotational speed N _P of the target rotational speed PREVT and current primary pulley is calculated by 221, in this DNP, while subjected to a known fed back correction and the ratio fed back control correction, Deyutei controls the solenoid 52 I do. Here, the execution condition of the feedback control is determined by 224, and if the execution condition is satisfied by the selector 225, the value subjected to the feedback correction is determined by:
If it is not satisfied, a non-feedback corrected one is selected and output to the duty conversion stage 226.

The conversion to the duty ratio and the like are the same as in the case of the lock-up.

Control of Secondary Chamber Oil Pressure The hydraulic pressure applied to the secondary chamber 37 (corresponding to the line pressure in this embodiment) is schematically shown in FIG. 1B (b), and is controlled as follows.

First, at 210, the engine output torque TQPUMP is calculated.
As calculation of the TQPUMP is shown in the FIG. 7A (a), to the engine output torque TQENG calculated based on Surotsutoru opening TVO and the engine speed N _E, the correction amount by the load correction TQLD and oil pump loss TQOPMP is reduced. As shown in FIG. 7A (b), the engine load includes an air conditioner, a power steering, and the like. In the present embodiment, the correction based on the engine coolant temperature THW is also taken into consideration. The correction of the pump loss, as shown in the FIG. 7A (c), is calculated based on the line pressure PSEC and the engine speed N _E to the secondary pulley.

Based on the engine torque TQPUMP taking these corrections into account, at 211, the torque TQIN input from the forward / reverse switching mechanism C to the primary pulley is calculated. This TQIN
Although the calculation of will be described in detail later, the engine output torque TQPUMP is accurately separated into two, a torque TQLUP transmitted through a lock-up clutch and a torque TQCVD transmitted through a converter, and is divided into a primary pulley. The transmitted torque is calculated accurately. In 212, this TQIN is input, and from the input gear ratio RATIO, the force FSEC required to press the pulley
Is calculated. The calculation of FSEC is described in detail below in connection with FIG. 7C. This pressing force FSEC is converted at 213 into a pressure PSECO after a centrifugal force correction is made at 214.
Therefore, this PSECO is the force FS required to properly transmit the output torque output via the torque converter and the lockup mechanism to the primary pulley according to the speed ratio RATIO.
This is a correction of EC by centrifugal force.

On the other hand, in 215, minimum required pressure PSMIN is calculated to operate the shift operation based on the transmission ratio RATIO and the primary pulley rotational speed N _P, and the like. And in selector 219, PSECO
And the larger of PSMIN is selected. This is a signal for driving the duty solenoid 51 for controlling the line pressure. How the centrifugal pressure is generated and the reason for selecting the larger of PSECO and PSMIN with selector 219 will be explained with reference to FIGS. 7C and 8.

<Details of Primary Pulley Input Torque Calculation> Here, the pulley input torque calculation performed in 211 will be described in detail with reference to FIG. 7B. As previously mentioned,
This control is performed when the engine output is transmitted to the primary pulley of the belt transmission mechanism D via the torque converter B and the lockup clutch, and the torque transmitted via the turbine runner 4 and transmitted via the lockup piston 6. This is performed in consideration of two of the torques to be applied. That is, the engine output torque TQENG is divided into two directions, a torque converter and a lockup mechanism. However, these two divided torques have different torque transmission ratios to the primary shaft 22 because their transmission paths are different. Therefore, in order to accurately grasp the torque transmitted to the shaft 22, it is necessary to accurately grasp how the engine output torque TQENG is divided into two.

In this embodiment, as shown in FIG. 7B, as a method of calculating the bisected torque, first, a torque TQLUP transmitted by a lock-up is calculated, and this torque TQLUP is subtracted from the engine output torque TQPUMP to obtain a torque converter. The torque TQCVD to be transmitted to the motor is calculated. The calculation method may be the reverse of the above, that is, the torque transmitted to only the torque converter may be determined, and then the torque transmitted to the lock-up may be determined.

First, based on the engine output TQPUMP, an initial value DPINT of the differential pressure before and after the lock-up piston is calculated according to the characteristic shown in FIG. 7A. The characteristic (a) is uniquely determined according to the lock-up mechanism. As described above, DPINT is converted into the aforementioned DPLUP by 201 in FIG. 1B, taking into account the lock-up range determination condition (FIG. 5A) and the like. This differential pressure DPLUP indicates the pressure transmitted to the lock-up piston 6. by the way,
The torque actually transmitted to the piston 6 is in the converter chamber.
Because the oil in 7a depends on the viscosity (ie oil temperature)
According to the characteristic as shown in FIG. 7B (b), that is, the oil temperature TH
Depending on the O, so THO is higher the lower the eyes of the torque is transmitted to the piston 6, to set the Rotsukuatsupu transmission ratio H _LU. The transmission ratio H _LU ranges from 0 to 0 as shown in FIG.
1 quantity. Then, from the transmission ratio H _LU and the engine torque TQPUMP, a torque TQLUP actually transmitted to the lock-up piston 6 is obtained from TQLUP = TQPUMP × H _LU as shown in FIG. Therefore, of the engine output, the torque TQCVD input to the torque converter alone is TQCVD = TQPUMP-TQLUP. Therefore, the torque output from the converter alone
TQCNVT is calculated in consideration of the torque ratio TR of (e), and becomes TQCNVT = TQCVD × TR = (TQPUMP−TQLUP) × TR. Therefore, the combined torque TQTRBN from the lockup converter is TQTRBN = TQLUP + TQCNVT. This torque is input to the switching mechanism C and transmitted to the primary pulley. Since the switching mechanism C is equipped with a planetary mechanism, the final torque value is obtained from TQTRBN in consideration of the reduction ratio KSRDCT by this mechanism.
TQIN is calculated.

TQIN = TQTRBN × reduction ratio The value of this reduction ratio KSRDCT is shown in (f) of FIG. 7B.

Now, here, using (d) and (f) of FIG. 7B,
The calculation of the torque ratio TR will be described. The torque ratio is a ratio of the torque transmitted by the converter to the input torque, and is obtained from the speed ratio E. This speed ratio E
It is one that is determined based on the ratio between the rotation speed N _P and the engine speed N _E of the primary pulley (N _P / N _E) and the shift range (RANGE), as the FIG. 7B (e) Is determined. In addition, at the time of reversing (R), assuming that KLRRT is a deceleration constant at the time of reversing of the forward / reverse switching mechanism, the speed ratio E _R is N _P / N _E / KL
RRT. When RANGE is at P and N, E is of course “0”
It is. The torque ratio is determined according to a characteristic such as (d). This characteristic is set to the maximum value “2” when the speed ratio E is zero, that is, when the vehicle is stopped, and is gradually reduced from “2” as the speed ratio E gradually increases from zero. When the reduction ratio E reaches about 0.8, the torque ratio TR
Is set to saturate to “1” and thereafter to be maintained at this “1”.

As described above, the pulley input torque calculation performed in 211, that is, the torque TQIN transmitted to the primary pulley shaft 22 was accurately obtained.

<Details of Secondary Pressure PSEC Calculation> Details of the secondary pressure PSEC calculation will be described by describing the controls performed in 212 to 216 in detail.

First, the calculation of the pressing force performed at 212 in FIG. 1B will be described with reference to FIG. 7C. This pressing force is a force to be applied to the piston 36 so that the effective diameters of the primary pulley and the secondary pulley are maintained at required amounts against the tension of the belt. Since the effective diameter changes according to the speed ratio, this pressing force must change according to the speed ratio RATIO.

The gear ratio RATIO is calculated as follows according to FIG. 7C.

First, as shown in the first Figure 7C (b), comparing the current primary pulley rotational speed N _P, the magnitude of the calculated target pulley speed PREVT at 220. This is because, in order to obtain the required pressing force, it is necessary to generate a larger pressing force between the pressing force required to maintain the current gear ratio and the pressing force required after shifting. N _P and PREV
Of T and the larger value selected by selector 218
If N _P , RATIO is 218, Becomes Then, the pressing force FSEC is calculated from the speed ratio RATIO and the input torque TQIN according to the characteristic (a) of FIG. 7C. In this case, the larger the speed ratio, that is, the overdrive, the larger the pressing force is required.

The details of the centrifugal oil pressure correction will be described in detail. The details of this correction control are shown in parts (c) to (g) of FIG. 7C. First, the reason why centrifugal oil pressure correction is necessary, and
The problems involved in performing this correction will be described with reference to FIG.

As shown in Figure 8A, when the secondary pulley of the primary pulley and A _S of the effective cross-sectional area A _P is rotated by the belt, the line pressure P _S (line pressure according to the common to these pulleys secondary side ), The pressing force F _P , F _S against each pulley is given by: F _P = A _P × P _S + K _P × N _P ² ... (1) F _S = A _S × P _S + K _S × N _S ² + F _SP (2) In each equation, each K is a predetermined constant,
The second term is the force due to the centrifugal force, and the third term in the second equation is the force due to the spring (38 in FIG. 2). Then, the ratio H _F of the pressing force between the pulleys (shown in the following equation)
Is a prerequisite for proper gear shifting. If this value is not set properly, there is a possibility that a situation may occur where the effective diameter of the secondary pulley does not change even though the effective diameter of the primary pulley is reduced.

Figure 8B is obtained by the F _P, the variation characteristics to the primary pulley rotational speed N _P in the graph. Similarly, Figure 8C shows
It shows that of F _S. As shown in these figures, the pressing force due to the centrifugal force increases in a square manner as the rotation speed N _P (N _S ) increases. Since this centrifugal force acts in the direction of pressing the pulley, subtracting this amount from the line pressure reduces the load on the hydraulic pump and contributes to a reduction in fuel consumption. The dashed line in FIG. 8C shows the change in pressure A _S × P _S when the line pressure P _S is reduced by the contribution of the pressure due to the centrifugal force. By the way, since the area ratio between the pulleys is A _P : A _S ≒ 2: 1 as described above, when the line pressure P _S is reduced by the centrifugal correction amount, the F _S becomes substantially constant. Even if you can keep it,
A _P × P _S decrease amount becomes large, the pressing force ratio H _F is no longer kept at a proper value. This is because the contribution by the centrifugal force is effective by the square of the rotation speed, but the absolute value itself is small. Therefore, when the centrifugal correction is large, the line pressure is considerably reduced. In such a case, the line pressure is reduced to less than the necessary minimum amount. The rotation speed NS of the pulley increases, so that the rotation speed N _S increases. That is, if the line pressure P _S is reduced by ΔP due to the centrifugal force correction, the change in the pressing force due to this is A _P × ΔP.
F _P is from being considerably reduced. Therefore, in this embodiment, monitors the decrease in P _S, if falls below a minimum line pressure is preferable to carry the Rimitsuto correction.
The lower limit value PSMIN at this time is defined by the following equation obtained by solving equation (3) for PS.

Returning to FIG. 7C, the description will be continued. The centrifugal force correction and the spring force correction are performed for the required pressing force FSEC according to (c) and (d) of FIG. The pressure exerted on the piston 36 of the secondary pulley by the piston area A _S
Calculate PSECO. That is, It is. Then, the selector 219 selects the smaller of PSECO and PSMIN. Then, in the third Figure 7C (h), a material obtained by multiplication of the safety factor K _SF for the sake of certainty of operation with respect to who this selected and PSEC.

Thus, the gear ratio, i.e., be filed in the pulley rotation speed N _S is anything, the line pressure PSEC in order to ensure the shift operation is secured.

<Solenoid Control> Next, control performed on the solenoids 51, 52, and 53 will be described. In this solenoid control, in order to reduce hydraulic pressure pulsation and ensure the durability of the solenoid, the drive frequency of the duty solenoid is adjusted based on the discharge amount of the hydraulic pump (that is, the engine speed N _E ) and the duty ratio. Control.

FIG. 9 shows the whole of this control. Since the control for the three solenoids is the same,
The duty conversion for the solenoid 52 as a representative of the solenoid will be described. First, in FIG. 9A, the signal CSTRK is converted into a duty. At this time,
When the oil temperature THO is high, the output DUTY is made large.
In (b), the battery power supply voltage is corrected for this duty. Then, upper and lower clip processing is performed in FIG. 9 (c), and fail safe processing is further performed in (d). Here, XSFT1F is a flag that stores the occurrence of a fail state.

On the other hand, it is set based on the DUTY determined the mode of driving cycle of the solenoid in the engine rotational speed N _E (a) in (e). In this embodiment, there are three types of periodic modes, LOW, MIDDLE, and HIGH, according to the zone determination. Then, in (f), the driving cycle is set to 10.5 ms (LO
W), 21.0 ms (MIDDLE), or 31.5 ms (HIGH). Thus, Deyutei control of this embodiment is to perform the driving frequency controlled by the engine rotational speed N _E and the DUTY value. The reason for the frequency control is that the pulsation is likely to be generated in the pump output when the duty ratio is variable while the frequency is constant. The drive frequency is set higher as the engine speed _NE is higher, in other words, as the discharge amount from the hydraulic pump 40 is larger. This is because the amplitude of the hydraulic pulsation tends to increase as the discharge amount increases, so that the driving cycle is shortened and the ripple is reduced.

<Control Procedure> The line pressure control and the duty control of the duty solenoid used in the hydraulic circuit in the hydraulic control device for the continuously variable transmission provided with the torque converter having the lock-up have been described above. The above control is performed by the control unit.
78 can be realized by both digital and analog computers. Therefore, a control procedure when such control is realized by a digital computer will be described next with reference to FIG. In step S2, the shift position signal RANGE is read from the sensor 82. When the shift position is at P or N, the transmission does not operate, so the turbine torque TQTRBN is set to “0” in step S23, and the
Proceed to 24.

Shift position is D, 1, 2, when in the R, the process proceeds to step S6, respectively from the sensors 86 and 85, the engine speed is read N _E and Surotsutoru opening TVO. Then, in step S7,
Accordance connexion to the FIG. 7A on the basis of the N _E and TVO, etc. (a),
Calculate engine torque TQENG. In step S8, the load is corrected in accordance with (b) of the figure, and in step S0, the torque loss is corrected by the oil pump to obtain the engine torque output TQPUMP.

Next, in step S10, the oil temperature THO is read from the sensor 88, and the initial value DPINT of the differential pressure of the lockup clutch is calculated based on TQPUMP according to (a) of FIG. 7B. As described above, the differential pressure DPLUP is calculated by the control for calculating the duty on the primary side based on this DPINT. Therefore, b look up class Tutsi transmission ratio H _LU from the DPLUP is accordance connexion calculated in FIG. 7B (b), at step S12, based on these DPLUP and H _LU, Russia look up class Tutsi transmission torque TQLUP is calculated . That is, TQLUP = DPLUP × H _LU . In step S14, the turbine speed _NT of the torque converter is read from the sensor 87. Then, in step S16, the speed ratio E becomes More required. Here, _NT is equivalent to N _P (or N _P × KLRRT) in (e) of FIG. 7B. Then, in step S18, the torque ratio TR is calculated. These have been described in connection with (d) and (e) of FIG. 7B. Next, in step S20, converter transmission torque TQCNVT is calculated. That is, TQCNVT = (TGPUM−PTQLU) × TR. Next, in step S22, a combined torque TQTRBN of the transmission torque TQLUP via the lock-up clutch and the transmission torque TQCNVT via the converter is calculated. That is, TQTRBN = TQLUP + TQCNVT. By correcting the planetary reduction ratio,
TQIN.

In step S24, the sensor 84 and 83, reads a secondary pulley rotation speed N _S and the primary pulley rotational speed N _P,
In step S26, the gear ratio RATIO (= N _P / N _S ) is calculated. Then, in step S28, the force FSEC required to press the pulley is calculated from TQIN and RATIO. This detailed description has been described with reference to FIG. 7C (a). Then, in steps S30 and S32, centrifugal force correction and spring correction are performed in accordance with (c) and (d) of FIG.
Get O. This PSECO becomes the target line pressure, that is, the piston pressure on the secondary pulley side.

This PSECO may drop below the minimum line pressure due to centrifugal force correction. Therefore, the minimum line pressure PSMIN according to the speed ratio and the like at that time is calculated by the following procedure. That is, in step S36, follow connexion to the aforementioned (3), calculates the pulley pressing force ratio H _F. Then, in step S38,
According to the equation (4), the minimum line pressure PSMIN required for shifting is calculated. Next, in step S40, the larger one of PSECO and PSMIN is selected. In step S42, performs correction by slave connexion safety factor K _SF in the first Figure 7C (h). Further, in step S44, the PSEC is converted into a duty ratio.
In step S46, the power supply voltage is corrected, and in step S48
Then, the duty DUTY is converted into the drive frequency. Then, it outputs to the solenoid 51 in step S50.

<Modifications> The present invention can be variously modified without departing from the gist thereof.

For example, in the above-described embodiment, when obtaining the turbine torque TQTRBN, first, the lock-up clutch transmission torque is calculated, and a value obtained by subtracting this from the engine output torque TQPUMP and adding the torque ratio TR to the converter transmission torque TQTRBN is calculated.
CNVT. Therefore, the following modified example is proposed. That is, first, the converter transmission torque is calculated, and a value obtained by subtracting this from the engine output torque TQPUMP and taking into account the lock-up clutch transmission ratio H _LU is used as the lock-up clutch transmission torque TQLUP.

Further, the following modification is proposed. In the above embodiment, the turbine torque was calculated to determine the line pressure. Therefore, the present invention is intended to be applied to detection of turbine torque itself. This is because TQTRBN accurately reflects the output from the torque converter.

Further, the following modification is proposed. In the above embodiment, DPINT was read from the map, and TQLUP was calculated from this DPINT. Alternatively, the pressure difference before and after the lock-up piston 6 may be directly measured by a sensor, and the TQLUP may be calculated from this.

Further, the following modification is proposed. In the above embodiment, the duty frequency control of the duty solenoid
The frequency is determined in consideration of two factors, namely, the pump discharge amount. However, the same effect can be obtained even if the frequency is determined in consideration of only the duty or the discharge amount.

Further, the drive frequency control of the duty solenoid shown in the above embodiment is not limited to the continuously variable transmission shown in FIG. 2 and FIG. Is also applicable to the transmission.

Further, the technique of adjusting the line pressure after detecting and separating the line pressure into the torque through the lockup clutch and the torque through the converter as shown in the above embodiment is shown in FIGS. 2 and 3. However, the present invention is not limited to the continuously variable transmission described above, and any hydraulic transmission can be applied to a general transmission and a turbine torque detecting device.

[Effects of the Invention] As described above, the line pressure control device for a hydraulically operated transmission according to claim 1 of the present invention has a duty cycle that determines a cycle of a drive signal for driving a duty solenoid according to an operation state. Determined according to the value of the rate. Therefore, the pulsation of the line pressure that fluctuates according to the operation state is efficiently driven because the solenoid itself is driven by the drive signal having the frequency determined according to the value of the duty ratio determined according to the operation state. Is suppressed.

Further, the line pressure control device according to claim 2 of the present invention,
Furthermore, since the cycle of the drive signal is determined in consideration of the pump discharge amount, the pulsation that tends to occur at the center of the duty is reduced well.

[Brief description of the drawings]

FIG. 1A is a diagram for explaining the configuration of the present invention, FIGS. 1B and 1C are diagrams showing the entire control of one embodiment according to the present invention, and FIG. 2 is a belt-type continuously variable motor according to the present invention. FIG. 3 is a skeleton diagram schematically showing a configuration of a continuously variable transmission to which one embodiment of a transmission line pressure control device is applied; FIG. 3 is a configuration of a hydraulic circuit connected to the continuously variable transmission shown in FIG. 2; FIG. 4 is a wiring diagram illustrating a configuration of a control circuit of a hydraulic circuit, FIG. 5A is a diagram illustrating characteristics for determining a control range of a lock-up operation, and FIG. 5B is a diagram of a solenoid for a lock-up clutch. FIG. 6 is a diagram illustrating control for calculating a duty, FIG. 6 is a diagram illustrating control when calculating a target rotation speed of a primary pulley, and FIG. 7A is a diagram illustrating control when calculating an engine torque TQPUMP. Fig. 7B shows the control when calculating the turbine torque TQTRBN. 7C is a diagram illustrating control when performing centrifugal force correction and the like, FIGS. 8A to 8D are diagrams illustrating how inconvenience occurs when performing centrifugal force correction, FIG. 9 is a diagram for explaining a method for controlling the drive cycle of the duty solenoid, and FIG. 10 is a diagram showing a control procedure when control according to the embodiment is performed by a digital computer. …… Engine, B …… Torque converter, C…
... forward / backward switching mechanism, D ... belt transmission mechanism, E ... reduction mechanism, F ... differential mechanism, RATIO ... speed ratio, E ... speed ratio, TR ... torque ratio, _NE ... engine rotation The number, N _P ...... primary shaft rotational speed, N _S ...... secondary shaft rotational speed, N _T ......
Turbine speed, P ... Line pressure, Q ... Hydraulic circuit, R
... Electrical control circuit, Z ... Continuously variable transmission, 1 ... Output shaft,
2 ... turbine shaft, 3 ... pump impeller, 4 ... turbine runner, 5 ... stator, 6 ... rock-up piston, 7 ... pump cover, 7a ... converter room, 8 ...
... One-way clutch, 9 ... Stator shaft, 10 ... Lock-up chamber, 11 ... Ring gear, 12 ... Sun gear, 13 ...
... 1st pinion gear, 14 ... 2nd pinion gear, 15 ...
Carrier, 16 ... clutch, 17 ... brake, 18 ... accumulator, 19 ... mission case, 20 ... belt, 21 ... primary pulley, 21a ... belt receiving groove, 22
…… Primary shaft, 23 …… Fixed conical plate, 24 …… Movable conical plate, 24a …… Outside surface, 25 …… Cylinder, 26 …… Piston, 27 …… Primary chamber, 31 …… Secondary pulley, 31
a ... belt receiving groove, 32 ... secondary shaft, 33 ... fixed conical plate, 34 ... movable conical plate, 34a ... conical friction surface, 34b ...
... Outside surface, 35 ... Cylinder, 36 ... Piston, 37 ... Secondary chamber, 38 ... Pressing spring, 40 ... Oil pump, 41 ... Line pressure regulating valve, 41a ... Pilot chamber, 41b ... Spool , 41c… Spring, 41d… Pressure regulating port, 41e… Drain port, 42… Redusing valve, 43… Speed ratio control valve, 43a… Spool, 43b…
… Spring, 43c …… Line pressure port, 43d …… Drain port, 43e …… Pilot port, 43f …… Pilot chamber, 43g …… Reverse port, 44 …… Switching valve, 44a ……
Spool, 44b Spring, 44c Pilot port, 44d First pilot pressure introduction port, 44e Second
Pilot pressure introduction port, 44f Pilot pressure supply port, 45 Shift valve, 46 Clutch pressure adjustment valve, 47
… Lock-up control valve, 48 …… Relief valve, 51
... First electromagnetic solenoid valve, 52... Second electromagnetic solenoid valve, 53... Third electromagnetic solenoid valve, 61.
62 …… Brake room, 70 …… Valve housing room, 71 …… Flow path, 72 …… Branch line, 73 …… Reservoir tank, 74…
… Return line, 75 …… Plunger, 76 …… Coil spring, 77 …… Electromagnetic coil, 78 …… Control unit, 82 ……
Shift position sensor, 83 …… Primary rotation speed sensor, 84
…… Secondary rotation speed sensor, 85 …… Throttle opening degree sensor, 86 …… Rotation speed sensor, 87 …… Turbine rotation speed sensor, 101, 102, 103, 104, 105, 106, 107, 108, 109…
... a line.

Continuation of the front page (58) Field surveyed (Int.Cl. ⁶ , DB name) F16H 59/00-61/12 F16H 61/16-61/24 F16H 63/40-63/48

Claims (3)

(57) [Claims] A line pressure control device for controlling a line pressure of a hydraulically operated transmission, comprising: a duty solenoid driven by a predetermined drive signal; and a pilot pressure controlled by the duty solenoid to reduce a discharge amount of an oil pump. A control valve for changing and adjusting the line pressure of the hydraulically operated transmission; a means for determining a duty ratio of the duty solenoid according to an operation state; and a frequency of the drive signal being changed based on the determined duty ratio. And a control unit for controlling the line pressure of the hydraulically operated transmission. 2. A line pressure control device for controlling a line pressure of a hydraulically operated transmission, comprising: a duty solenoid driven by a predetermined drive signal; and a pilot pressure controlled by the duty solenoid to reduce a discharge amount of the oil pump. A control valve for changing and adjusting the line pressure of the hydraulically operated transmission; a means for determining a duty ratio of the duty solenoid according to an operation state; a duty ratio for which the frequency of the drive signal is determined; and the oil pump. A line pressure control device for a hydraulically operated transmission, comprising: means for performing a change control based on a discharge amount of a hydraulically operated transmission. 3. The line pressure of a hydraulically operated transmission according to claim 2, wherein the discharge amount of the oil pump is determined according to the number of revolutions of the input torque of the hydraulically operated transmission. Control device