JPS62196455A - Hydraulic controller for belt type continuously variable transmission for vehicle - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

Description

DETAILED DESCRIPTION OF THE INVENTION Technical Field The present invention relates to an improvement in a hydraulic control device for a belt-type continuously variable machine for a vehicle.

PRIOR ART A pair of primary side variable pulleys and a primary side variable boolean provided on the primary side rotating shaft and the primary side rotating shaft, respectively, and a transmission heel 1- that is wound around the pair of variable pulleys to transmit power, A belt-type continuously variable transmission for a vehicle is known that includes a pair of primary hydraulic cylinders and a primary hydraulic cylinder that respectively change the effective diameters of the pairs of variable pulleys.

The speed ratio and the tension of the transmission belt of such a continuously variable transmission are such that, as described in Japanese Patent Publication No. 5B-29424, the hydraulic oil from the hydraulic source is supplied to one side of the hydraulic cylinder and at the same time flows out from the other side. The control valve is controlled by a hydraulic control device including a control valve (four-way valve) that changes the speed ratio by the control valve, and an electromagnetic relief valve that regulates the pressure of the hydraulic oil flowing out from the control valve.

A feature of such a hydraulic control device is that relatively high-pressure hydraulic oil supplied from a hydraulic source is supplied to the primary hydraulic cylinder and one of the primary hydraulic cylinders (Jt), and at the same time, the hydraulic oil in the other cylinder is caused to flow out. Accordingly, a speed change control valve is provided that adjusts the speed ratio of the continuously variable transmission by changing the effective diameter of the primary side variable pulley and the primary side variable pulley, and the speed change control valve adjusts the speed ratio of both hydraulic cylinders. The side where the internal oil pressure is high during power transmission (
The relatively high working pressure from the hydraulic source is applied to the hydraulic cylinder located on the drive side (drive side), and the hydraulic pressure regulated by the electromagnetic relief valve is applied to the hydraulic cylinder on the opposite side, so the power transmission direction is opposite. Even in this case, the tension and speed ratio of the transmission belt can be suitably controlled. However, in such a hydraulic control device, at least in the neutral position, the flow cross-sectional area on the outflow side for outflowing the hydraulic oil from one of the primary hydraulic cylinder and the secondary hydraulic cylinder of the speed change control valve is Since the flow cross-sectional area on the inflow side for supplying hydraulic oil to one hydraulic cylinder is equivalent to is the oil pressure between the oil pressure source (first line oil pressure) and the low pressure side oil pressure source (second line oil pressure, etc.). Also, the oil pressure of the output boat to the other hydraulic cylinder is the same. Therefore, the oil pressure on the low pressure side Since the pressure of the hydraulic power source is not dominantly transmitted into the hydraulic cylinder, the pressure of the low pressure side hydraulic power source is adjusted to optimize the clamping force on the transmission belt, so that the pressure of the low pressure side (driven side) hydraulic cylinder is It was difficult to try to control the hydraulic pressure inside.

Problems to be Solved by the Invention In order to solve this problem, it is possible to solve the problem by draining the hydraulic oil in the hydraulic cylinder through a throttle, or by connecting the hydraulic cylinder to a low-pressure side hydraulic source through a throttle. It is conceivable to make the pressure of the output coupler 1 at the neutral position approximately equal to the pressure of the low pressure side hydraulic source. However, when the hydraulic oil in the hydraulic cylinder leaks out through the throttle, the output pressure of the speed change control valve changes due to the leakage of the hydraulic oil.
Since the overall F is low and the rate of change tends to be slow, the amount of operation of the speed change control valve (for example, one shift of the spool valve,
Otherwise, the change in the output oil pressure (increase in pressure curve L) with respect to the control amount for the speed change control valve becomes small, resulting in a large control deviation. In addition, when a hydraulic cylinder is connected to a low-pressure hydraulic source through a throttle, it is constantly affected by the hydraulic pressure of the low-pressure hydraulic source, so the overall output pressure characteristics of the speed change control valve are low and the rate of change is very gradual. As mentioned above, there is a drawback that the control deviation increases.

Means for Solving the Problems The present invention was made against the background of the above circumstances.
The gist of this is that the primary rotating shaft and the pair of primary variable pulleys provided on the primary rotating shaft, respectively, and the transmission that is wrapped around these pair of variable pulleys and transmits power. In a belt-type continuously variable transmission for a vehicle, which is equipped with a belt and a pair of primary hydraulic cylinders that change the effective diameters of the pair of variable pulleys, a relatively high pressure is supplied from a hydraulic source. The effective diameter of the primary variable pulley and the primary variable pulley is changed by supplying hydraulic oil to the primary hydraulic cylinder and one of the primary hydraulic cylinders and at the same time flowing out the hydraulic oil in the other cylinder. The hydraulic control device includes a speed change control valve that adjusts the speed ratio of the continuously variable transmission by adjusting the speed ratio of the continuously variable transmission. The rate of change in the flow cross-sectional area on the inlet side for supplying hydraulic oil to one of the hydraulic cylinders is greater than the rate of change in the cross-sectional area on the outlet side for flowing hydraulic oil out of the one hydraulic cylinder. This is because the speed change control valve is constructed so as to be.

Operation and Effects of the Invention With this arrangement, when the speed change control valve is operated from its neutral position, the reservoir for supplying hydraulic oil to one of the downstream hydraulic cylinder and the primary hydraulic cylinder Since the flow cross-sectional area on the inlet side of the oil cylinder is increased or decreased at a larger rate of change than the flow cross-sectional area on the outflow side for draining the hydraulic oil from one of the oil cylinders, the shift control valve from the neutral position Along with the operation, the rising curve of the hydraulic pressure in the one hydraulic cylinder is maintained relatively steeply. Therefore, the control value for the speed change control valve, that is, the amount of increase in the output oil pressure per movement position of the spool valve element of the speed change control valve becomes large, and the control deviation in the speed change control valve becomes small.

Here, the speed change control valve preferably includes a valve body in which a cylinder bore is formed, and a spool valve element having a plurality of lands and slidably fitted into the cylinder bore. A cross-sectional area of a portion of one land formed on the spool valve that is related to the outflow side flow cross-sectional area is smaller than a cross-sectional area of a portion that is related to the inflow side flow cross-sectional area. In this land, the cross-sectional area is increased in the axial direction by providing a large-diameter portion at the end that is related to the flow cross-sectional area on the inflow side, or a groove at the end that is related to the flow cross-sectional area on the outflow side. It changes as you go.

Preferably, in the neutral position, the speed change control valve has an outflow side flow cross-sectional area for outflowing hydraulic fluid from one of the primary side hydraulic cylinder and the primary side hydraulic cylinder. It is configured by being formed larger than the inflow side flow cross-sectional area for supplying hydraulic oil to the hydraulic cylinder. As a result, even in the neutral position, the output oil pressure to the one hydraulic cylinder is controlled by the oil pressure of the low pressure side oil pressure source, so the tension of the transmission belt can be easily controlled by regulating the pressure of the low pressure side oil pressure source. In such a case, it is preferable that the flow cross-sectional area on the outflow side is formed by a negative overlap between the land formed on the spool valve and the valve body, and the flow cross-sectional area on the inflow side is positive. Or formed by a superposition of zeros. Further, the outflow side flow cross-sectional area and the inflow side flow cross-sectional area are formed by the negative overlap between the groove formed on the spool valve and the valve body, and the negative overlap forming the outflow side flow cross-sectional area is It is made larger than the overlap of the monks forming the flow cross-sectional area on the inflow side.

Example Hereinafter, one embodiment of the present invention will be described in detail.

In FIG. 1, the output of an engine IO installed in a vehicle is transmitted to a primary rotating shaft 16 of a belt type continuously variable transmission 14 via a clutch 12.

Bell] - The continuously variable transmission 14 includes a primary rotating shaft I6 and a primary rotating shaft 18, and a primary variable pulley 2o with a variable effective diameter attached to the primary rotating shaft 16 and the primary rotating shaft 18. And change the effective diameter of the primary variable pulley 22, the primary variable boob I720 and the transmission belt 24 that is wrapped around the primary variable pulley 22 to transmit power, and the primary variable pulley 20 and the primary variable pulley 22. The primary side hydraulic cylinder 26 and the primary side hydraulic cylinder 28 are provided. The primary side hydraulic cylinder 26 and the primary side hydraulic cylinder 28 are formed to have the same pressure receiving area, and are provided with the primary side variable boob I described in -I-.
The belt type continuously variable transmission 14 is made smaller because the I type J20 and the primary side variable boob IJ22 are equivalent.

The primary side variable pulley IJ 20 and the primary side variable pulley 22 include fixed rotating bodies 31 and 32 fixed to the primary rotating shaft 16 and the primary rotating shaft 18, respectively, and 1-the primary rotating shaft 16. and movable rotating bodies 34 and 36 which are provided on the primary rotating shaft 18 so as to be non-rotatable but movable in the axial direction and form a V-groove with the fixed rotating bodies 31 and 32, respectively.

The output from the primary rotating shaft 18 of the belt type continuously variable transmission 14 is transmitted to the drive wheels of the vehicle via an auxiliary transmission, a differential gear device, etc. (not shown).

A hydraulic control circuit for operating the vehicle power transmission device configured as described above is configured as described below. That is, the oil tank 38 passes through a return path (not shown).
The hydraulic oil returned to the strainer 40 and the suction oil passage 41
The oil is sucked into the oil pump 42 via the transmission control valve 44.
The oil is fed under pressure to the first line oil passage 50 connected to the zero-input port 46 and the first pressure regulating valve 48 . This oil pump 42
constitutes the hydraulic power source of this embodiment, and is driven by the engine 10 via a drive shaft (not shown).

The first pressure regulating valve 48 causes a part of the hydraulic oil in the first line oil passage 50 to flow out to the second line oil passage 52 in accordance with a first drive signal Vr)1, which will be described later.
f, is controlled. This second line oil passage 52 is connected to a first discharge boat 54 and a second discharge boat 5 of the speed change control valve 44.
6 and a second pressure regulating valve 58, respectively. The second pressure regulating valve 58 is configured to have a relative pressure relative to the first line oil pressure Pet by causing a part of the hydraulic oil in the second line oil passage 52 to flow out to the l/len oil passage 60 in accordance with a second drive signal VD2, which will be described later. The second line oil pressure Pβ2 is controlled to be relatively low. ”
- The first pressure regulating valve 48 and the second pressure regulating valve 58 are constructed from so-called electromagnetic proportional relief valves that are constructed according to the same specifications. Note that in this embodiment, the first line oil pressure pH
corresponds to the hydraulic pressure of the high-pressure side hydraulic source, and the second line hydraulic pressure P12 corresponds to the hydraulic pressure of the low-pressure side hydraulic source.

The speed change control valve 44 is a so-called JL control solenoid valve, and is connected to the input boat 4G, the first discharge boat 54, the second discharge boat 56, the primary hydraulic cylinder 26, and the primary hydraulic cylinder 28. A cylinder bore 66 is formed in the valve body 65 so as to communicate with a pair of first output capotros 2 and second output capotros 4 connected via passages 29 and 30, respectively, and a cylinder bore 66 in the cylinder bore 66 One spool valve 68 is slidably fitted into the spool valve 68, and the spool valve 68 is biased toward the neutral position from both ends of the spool valve 68.
A pair of first springs 70 and a second spring 72 that hold the spool valve 68 in a neutral position, and a pair of first springs 70 and a second spring 72 that hold the spool valve 68 in the neutral position; A first electromagnetic solenoid 74 and a second electromagnetic solenoid 76 are provided to move the electromagnetic solenoid 74 against the urging force of the electromagnetic solenoid 74 . ”-Spool Benko 68
Four lands 78, 80, 82, and 84 are formed sequentially from one end, and a pair of lands 80 and 82 located in the middle are arranged so that the axis of the spool valve 68 is aligned when the spool valve 68 is in the neutral position. It is formed at the same position as the first output capotro 2 and the second output capotro 4 in the direction. Also, on the inner circumferential surface of the cylinder bore 66, when the spool valve element 68 is in the neutral position (171), the position facing the pair of lands 80 and 82, i.e. A land 8 is located at the position where the trolley 4 opens into the inner circumferential surface of the cylinder bore 66.
A pair of first annular grooves 86 and second annular grooves 88 having widths slightly larger than 0 and 82 are formed. The first annular groove 86 and the second annular groove 88 form a constriction whose flow cross-sectional area changes continuously in order to control the flow of hydraulic oil between the lands 80 and 82.

As shown in detail in FIG. 2, in the shift control valve 44, when the spool valve element 68 is in the neutral position,
The flow cross-sectional area on the outflow side for causing hydraulic oil to flow out from the primary hydraulic cylinder 26 or 28 is larger than the cross-sectional area of the inflow side flow 1111 for supplying hydraulic oil to the hydraulic cylinders 26 or 28. ing. That is, the gaps 110 and 112 formed in the constricted portions that restrict the flow of hydraulic fluid from each output port 1-62 and 64 to each discharge boat 54 and 56, respectively, are used to restrict the flow of hydraulic fluid from each output port 46 to each output port 2 and 64. The gap is larger than the gaps 114 and 116 formed in the constriction portions that restrict the flow of hydraulic fluid toward the gap. - The gaps 110, 112, 114 and 11.6 are the negative overlap between the land 80.82 and the valve body 65 (
The outer circumferential surface of the land and the inner circumferential surface of the cylinder bore 66 do not face each other and do not overlap each other.
The negative overlap in gaps 110.112 is in gaps 114.
This is greater than the negative overlap at 116.

In addition, in the speed change control valve 44, the lands 80 and 82
Large diameter portions 120 and 122 are formed on the input boat 46 side, respectively. For this reason, the cross-sectional area of the lands 80 and 82 becomes larger as it moves in the axial direction toward the input boat 46, and as a result, the cross-sectional area of the inlet flow 'Jm' increases as the spool valve moves from the neutral position on the 6th day. It is configured such that the rate of change in the area is larger than the rate of change in the flow cross-sectional area on the outflow side. For example, spool valve 68
When moving to the second electromagnetic solenoid 76 side, that is, to the right side in the figure, the inflow side flow cross-sectional area in the gap 116 increases rapidly, whereas the outflow side flow cross-sectional area in the gap 112 gradually decreases. On the contrary, spool bento 68
When moving to the left in the figure, the flow cross-sectional area on the inflow side in the gap 114 increases rapidly, while the flow cross-sectional area on the outflow side in the gap 110 gradually decreases.

As a result, when the spool valve 68 is in the neutral position, the same pressure is output to the first output capotro 2 and the second output capotro 4, but -1=, the gap 11
Since 0.112 is set larger than the gaps 114 and 116, its output pressure is approximately equal to the second line oil pressure P12. In addition, from each output port 2.64, an amount of hydraulic oil is supplied to the primary side hydraulic cylinder 26 to replenish the leakage.
and is supplied to the primary hydraulic cylinder 28.

However, as the spool valve element 68 is moved from the neutral position in its uniaxial direction, for example, in the direction approaching the second electromagnetic solenoid 76 (i.e., rightward in the figure), the first output capotro 2 and the first discharge The cross-sectional area of flow with the boat 54 is continuously increased, and the cross-sectional area of flow between the second output capotro 4 and the second discharge boat 112 is continuously decreased, so that the 1111th capotro 2 becomes the While maintaining the 2-line hydraulic pressure PI2z, the second output capotro 4 and input boat 46
Since the cross-sectional area of flow between the second output capotro 4 and the primary hydraulic cylinder 28 is continuously increased, the working oil pressure output from the second output capotro 4 to the primary hydraulic cylinder 28 is gradually increased and finally reaches the first line oil pressure P.
It is said to be 7!1. For this reason, the balance between the thrust forces of the primary hydraulic cylinder 26 and the primary hydraulic cylinder 28 in the belt-type continuously variable transmission 14 is disrupted, so that while hydraulic oil flows into the primary hydraulic cylinder 28, the primary hydraulic cylinder 2
The hydraulic oil in 6 flows out, and the speed Ite (rotational speed N.ut of the primary rotating shaft 1B/rotational speed Ni of the primary rotating shaft 16) of the belt-type continuously variable transmission [14] decreases.

On the other hand, as the spool valve element 68 is moved from the neutral position toward the first electromagnetic solenoid 74, that is, to the left in the figure, the connection between the second output capo I-64 and the second discharge boat 56 increases. The flow cross-sectional area is continuously increased and the second output capotro 4 becomes the second line hydraulic pressure PI! , 2 is maintained, while the flow cross-sectional area between the first output capotro 2 and the input capo I. Since the area is reduced, the working oil pressure outputted from the first output capotro 2 to the primary side hydraulic cylinder 26 is gradually increased and finally becomes the first line oil pressure PC1. For this reason, the balance of the thrust of the primary hydraulic cylinders 26 and 28 in the belt-type continuously variable transmission 14 is disrupted, so that the hydraulic oil in the primary hydraulic cylinders 28 flows out, while the primary hydraulic Hydraulic oil flows into the cylinder 26, and the speed ratio e of the belt type continuously variable transmission 14 increases. In this way, the variable link control valve 44 has a switching valve function that supplies high-pressure hydraulic oil to one of the hydraulic cylinders 26 and 2B and low-pressure hydraulic oil to the other, and a continuous flow of hydraulic oil. It also has a flow control valve function to adjust the meter.

FIG. 3 shows the output hydraulic pressure characteristics of the speed change control valve 44 which operates as described above. In the figure, the horizontal axis represents the amount of displacement of the spool valve 68 from the neutral position, that is, the shift control valve 44
Control meter for ■. Not shown.

Belt-type continuously variable transmission of the vehicle [14 includes a primary rotating shaft 16
the first rotation sensor 90 for detecting the rotation speed N i n of the primary rotation shaft 18; and the rotation speed N o u of the primary rotation shaft 18
A second rotation sensor 92 for detecting t is provided, and the first rotation sensor 90 and the second rotation sensor 9
From 2 onwards, the rotational speed N1. , represents the rotation signal SRI and the rotation speed N. A rotation signal SR2 is output to the controller 94 without indicating U. Further, the engine 10 is provided with a throttle sensor 96 provided in the intake pipe for detecting the throttle valve opening θ, and an engine rotation sensor 98 for detecting the engine rotation speed N8. The throttle sensor 96 and the engine rotation sensor 98 output a throttle signal Sθ representing the throttle valve opening θth and a rotation signal SE representing the engine rotational speed N8 to the controller 94.

The controller 94 has a CP[J102, ROM10
4. A so-called microcomputer including a RAM 106 and the like, which constitutes the control means of this embodiment. -1-C
PU] 02 processes input signals according to a program stored in advance in the ROM 104 while utilizing the storage function of the RAMI 06, and operates the first pressure regulating valve 48 and the second pressure regulating valve 48 to control the first line oil pressure and the second line oil pressure. 2 Pressure regulating valve 5
8 to the first drive signal MDI and the second drive signal Vl'')2
The first electromagnetic solenoid 74 and the second electromagnetic solenoid 7 simultaneously supply the speed ratio e.
6 are supplied with speed ratio signals RAI and RA2 to drive them.

The speed ratio control of the belt-type continuously variable linkage machine 14 is performed, for example, by repeatedly executing a speed ratio control routine shown in FIG. 4. That is, first, by executing step S1, the rotational speed N of the primary rotating wheel 16,
fi, rotational speed N of the primary rotating shaft 18; uL, throttle valve opening θ1, and engine rotation speed N are the rotation signal SRI
and is read into RAM 106 based on SR2, throttle signal Sθ, and rotation signal SE. Next, in step S2, according to the following equation (1) stored in advance in the ROM 104, the speed ratio e is equal to the above rotational speed N8ゎ and NO
+,.

Calculated from.

e-Nout/N, a... (1) Also, in step S3, the target rotational speed N i is determined based on the throttle valve opening θ from the relationship stored in the RAM 106.
fi' is determined, and from the above fil2 its target rotational speed N i n '' and from the actual rotational speed N.ut are ``1''.
Calculate the target speed ratio e''.'' - Note 11 Standard rotational speed N,
The relationship for determining ,,II is, for example, shown in FIG.
This is determined in advance so that the engine 10 operates exclusively at . In the following step S4, the speed ratio control value V is calculated according to the following equation (2) stored in advance in the ROM]04. is calculated. In step S5, this speed ratio control value V. If is positive, the spool valve element 68 is moved to the left and the rotational speed N of the primary rotation shaft I8 is increased. ut
The speed ratio signal RA2 is outputted so that the speed ratio signal RA2 increases, and when the speed ratio signal RA2 is negative, the spool valve 68 is moved to the right and the speed ratio signal RA2 is outputted so that the rotational speed Nrn of the primary side rotating shaft 16 increases. A ratio signal RAI is output. Also, the speed ratio control value ■. The magnitude corresponds to the magnitude of the speed ratio signal RA1 or the speed ratio signal RA2, that is, the amount of movement of the spool valve 68. Therefore, ζ, as is clear from the following equation (2), "varnish speed change control".

is determined so that the actual speed ratio e matches the target speed ratio e*. Note that Equation (2) is a control constant.

Vo −k (e” e) /e H4121 Then, in step S5, the speed ratio control value V determined in the previous step is output, and the speed ratio e is controlled.

In such speed ratio control, in this embodiment, the land 8 formed on the spool valve element 68 of the speed change control valve 44
0.82 are formed with large diameter portions 120° and 122, respectively, so that the rate of change in the flow cross-sectional area on the inflow side as the spool valve 68 moves from the neutral position is greater than the rate of change in the flow cross-sectional area on the outflow side. As shown in FIG. 3, the higher pressure side of the output oil pressure of the first output capo 1-62 or the second output capo 4, in other words, the oil pressure P1□ or P. The rise on the high pressure side of ut is steep. Moreover, in this embodiment, at the neutral position, the flow 1ffl cross-sectional area on the outflow side is made larger than the flow cross-sectional area on the inflow side, so the oil pressure of the output port 2.64 is controlled by the second line oil pressure Pe2. The low pressure side is maintained at the second line oil pressure l]t7!2. As a result, a large change (gain) in the output oil pressure based on the amount of movement of the spool valve element 68 from the neutral position, in other words, the control value for the speed change control valve 44 is obtained, resulting in a control deviation (steady deviation) in speed ratio control. ΔV
This means that o becomes smaller.

Also, the lower pressure side of the output oil pressure of the I-starting 1 output capotro 2 or the 2nd output capotro 4 is the 2nd line oil pressure P7!2
As a result, the second line hydraulic pressure P12, which is a lower pressure side hydraulic pressure source, is regulated in the second pressure regulating valve 58, thereby increasing the hydraulic pressure P8 in the driven side hydraulic cylinder. ．． Or P. ,,
t, that is, the clamping force on the transmission belt can be easily controlled.

Incidentally, Fig. 7 shows the output hydraulic pressure characteristics of a conventional speed change control valve without any modifications to the speed change control valve and hydraulic cylinder, and Fig. 8 shows the output hydraulic characteristics of a conventional speed change control valve in which no modifications are made to the speed change control valve and the hydraulic cylinder.
2 shows the output oil pressure characteristics of the speed change control valve when a throttle communicating with the speed change control valve is provided. As is clear from the figure, in Fig. 7, the hydraulic pressure of the output boat on the low pressure side is
This does not match the line oil pressure P7!2, making it difficult to control the tension of the transmission belt 24 by the second pressure regulating valve 58, that is, to control the clamping force. In addition, in FIG. 8, the control deviation ΔVo becomes large because the high-pressure side output boat rises slowly over the entire area due to oil leakage from the handle 11 through the throttle.

Further, FIG. 9 shows the output hydraulic pressure characteristics when the large diameter portions 120 and 122 are not formed in the speed change control valve 44 of the above-described embodiment. The output oil pressure characteristic shown in FIG. 3 has a steep rise compared to the output oil pressure characteristic shown in FIG. 9. In addition, FIGS. 8 and 9 show the same differential pressure of re-output captro 2.64 as in FIG. 3, that is, the thrust is balanced in order to maintain the speed ratio of the belt type continuously variable transmission 14 at a certain value. P (=Pin-1) to the difference in pressure values in the state where the A control value (control deviation ΔVO) corresponding to ut ) is shown.

Such output hydraulic characteristics are such that the rate of change in the flow cross-sectional area of the transmission 1b control valve 44 is different on the inflow side of the passage where hydraulic oil flows into the hydraulic cylinder and on the outflow side of the passage where hydraulic oil flows out from the same hydraulic cylinder. Due to different things. In other words, the cross-sectional area of the flow 1ffl in the case shown in Fig. 9 changes at the same rate as shown in Fig. 10, but the
In the case of this embodiment shown in the figure, the rate of increase in the flow cross-sectional area is different between the inflow side and the outflow side, as shown in FIG. 11. Here, in the figure, A indicates the minimum flow i ([1 cross-sectional area (inflow side)) of the flow path between the input boat 46 and the first output capotro 2, B indicates the input boat 46 and the second output capotro - the minimum flow cross-sectional area (inflow side) of the flow path between I. Area (outflow side), D is the second
This is a line indicating the minimum flow cross-sectional area (outflow side) of the flow path between the discharge boat 56 and the second outlet port 4. 1st
As is clear from FIG. 1, the rate of increase in the flow cross-sectional area on the inflow side (the slope of line A or line B) in the speed change control valve 44 is compared to the rate of increase in the cross-sectional area on the outflow side (slope of line C or line D). It is set to be large. Note that the flow 1ffl cross-sectional area S between the discharge boat 54.56 and the output port 2.64 when the spool valve element 68 is in the neutral position is equal to the hydraulic pressure in the low-pressure side hydraulic cylinder equal to the second line hydraulic pressure P12. It is a value. Therefore, this value is determined by the specifications of the continuously variable transmission a14, and is equivalent to 31 (1) in FIG.

Next, another embodiment of the present invention will be described. In addition, in the following embodiments, parts with similar functions are given the same reference numerals and explanations will be omitted.

In the above-mentioned embodiment, the output hydraulic characteristics shown in FIG. 3 are obtained by forming the large diameter portions 120 and the lengths 122 in the lands 80 and 82 of the spool valve 68, respectively. As shown in FIGS. 12, 13, and 14, notches 124 are formed on the first discharge boat 54 side and the second discharge boat 56 side of the lands 80 and 82, respectively, to reduce the rate of change in the flow cross-sectional area. It is okay to do so. Since this notch 124 reduces the rate of change in the cross-sectional area of flow, the rate of change in the cross-sectional area of flow on the side where the notch 124 is not formed is relatively 10W.

Also, for example, as shown in FIGS. 15 and 16, a first annular groove 8 formed on the inner circumferential surface of the cylinder bore 66 may be used.
6 and the discharge boat 54, 56 side of the second annular groove 88,
It is also possible to form cutouts 126 that reduce the rate of change in the flow cross-sectional area on the outflow side and relatively increase the rate of change in the flow cross-sectional area on the inflow side.

In addition, in the speed change control valve 44 of the embodiment described above, the second line oil passage 52 is connected to the input boat 4G, and the second line oil passage 52 is connected to the first discharge boat 1-5.
4 and the second discharge boat 56 may be connected to the first line oil passage 50. In this case, the spool valve 68 is provided with three lands, as shown in FIG. 17 or 18, for example. In FIG. 17, notches 130 are formed on both sides of the central land, and the
In the figure, the land on both sides has a large diameter, so that the rate of change in the flow cross-sectional area on the inlet side is larger than the rate of change in the flow cross-sectional area on the outflow side.

Although the embodiment of the present invention has been described above based on the drawings, the present invention can also be applied to other aspects.

For example, in the speed change control valve 44 shown in FIG.
20 and 122, respectively, so that the rate of change in the flow cross-sectional area on the inlet side is larger than the rate of change in the flow cross-sectional area on the outflow side in the lands 80 and 82 of the spool valve 68, but if one land, e.g. Even if the large diameter portion 122 is formed only in the land 82 and the rate of change on the inflow side is larger than the rate of change on the outflow side for the inflow side flow cross-sectional area and the outflow side flow cross-sectional area in which the land 82 is involved, This results in the following effects.

In addition, in the speed change control valve 44 of FIG. 2 of the above-described embodiment, the flow cross-sectional area of the gaps 110, 112 and 114, 116 was set depending on the magnitude of the negative overlap state, but the gaps 114, 116 were The overlap may be positive or zero, and the gaps 110, 112 may be negative overlaps.

Further, although the speed change control valve 44 in the above-described embodiment was equipped with a single spool valve element 68, it may be a control valve having a plurality of spool valve elements.

Further, the spool valve 6 of the speed change control valve 44 of the above-mentioned embodiment
8 is a type of valve that is directly driven by an electromagnetic solenoid 74, 76, but it may also be a type of valve that is driven by hydraulic pressure on a pylon controlled by an electromagnetic valve or the like. . In this case, the solenoid valve can be subjected to so-called duty control.

Furthermore, although the aforementioned speed change control valve 44 was provided with electromagnetic actuators 74 and 76 for driving its spool valve element 68, other electromagnetic actuators such as a pulse motor could also be provided. good.

Note that the above-mentioned embodiment is merely one embodiment of the present invention, and various modifications may be made to the present invention without departing from the spirit thereof.

BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a diagram showing the configuration of an embodiment of the present invention. Second
This figure is a diagram illustrating in detail the main parts of the speed change control valve shown in FIG. 1. FIG. 3 is a diagram showing the output oil pressure characteristics of the speed change control valve of FIG. 2. FIG. 4 is a flowchart illustrating the speed change control operation in the apparatus shown in FIG. FIG. 5 is a diagram showing a relationship used to explain FIG. 4, and a relationship for determining the target rotational speed. FIG. 6 is a diagram showing the minimum fuel consumption rate curve of the engine of FIG. 1. FIG. 7, FIG. 8, and FIG. 9 are diagrams showing output hydraulic characteristics of conventional speed change control valves, respectively. FIG. 10 is a diagram showing the change characteristics of the flow cross-sectional area within the speed change control valve having the output hydraulic characteristics shown in FIG. 9. 11th
The figure is a diagram showing the change characteristics of the flow cross-sectional area within the speed change control valve having the output hydraulic characteristics shown in FIG. 3. Figure 12, 1st
3 and 14 are views showing essential parts of other embodiments of the present invention, and are a side view, a sectional view taken on a plane including the axis, and a sectional view taken on a plane perpendicular to the axis. . 15th
16 and 16 are views showing essential parts of another embodiment of the present invention, and are a sectional view taken along a plane parallel to the axis and a sectional view taken along a plane perpendicular to this plane. Figures 17 and 18
Each figure corresponds to FIG. 2 in another embodiment of the present invention. 14: Belt type continuously variable transmission 16: Primary side rotating shaft 182 Primary side rotating shaft 20 Knee variable pulley 22: Primary side variable pulley 24: Transmission belt 26: Primary side hydraulic cylinder 2B 2nd primary side hydraulic cylinder 44: Shift control valve 65: Valve body 66: Cylinder bore 68 Nispool valve so, a2: Land applicant Tokota Jidosha Co., Ltd. Nt
. (x+o3 cow rpm) Figure 5 Figure 6 (-)Oli! 11 Quiet value 00 Figure 9￥i'l 'fRmV'. Figure 11 Figure 17 Go Elephant Ne City -, tT: Three = = <1:1 shot> Showa 6
1f1st March 29th

Claims (5)

[Claims] (1) A pair of primary variable pulleys and a secondary variable pulley provided on the primary rotating shaft and the secondary rotating shaft, respectively, and a transmission belt that is wound around the pair of variable pulleys to transmit power; In a vehicle belt-type continuously variable transmission comprising a pair of primary and secondary hydraulic cylinders that change the effective diameters of the pair of variable pulleys, relatively high-pressure hydraulic oil is supplied from a hydraulic source. The effective diameters of the primary variable pulley and the secondary variable pulley are changed by supplying hydraulic oil to one of the primary hydraulic cylinder and the secondary hydraulic cylinder and simultaneously causing hydraulic oil in the other to flow out. A hydraulic control device equipped with a speed change control valve that adjusts a speed ratio of the continuously variable transmission, wherein when the speed change control valve is operated from a neutral state, the primary hydraulic cylinder and the secondary hydraulic cylinder are activated. so that the rate of change in the flow cross-sectional area on the inlet side for supplying hydraulic oil to one of the hydraulic cylinders is greater than the rate of change in the flow cross-sectional area on the outflow side for draining the hydraulic oil from the one hydraulic cylinder,
A hydraulic control device for a belt-type continuously variable transmission for a vehicle, characterized in that the speed change control valve is configured. (2) The speed change control valve includes a valve body in which a cylinder bore is formed, and a spool valve element having a plurality of lands and slidably fitted into the cylinder bore. The vehicle according to claim 1, wherein a cross-sectional area of a portion of one land formed on the child that is related to the outflow-side flow cross-sectional area is smaller than the inflow-side flow cross-sectional area. Hydraulic control device for belt type continuously variable transmission. (3) In its neutral position, the speed change control valve has an outflow side flow cross-sectional area for causing hydraulic oil to flow out from one of the primary side hydraulic cylinder and the secondary side hydraulic cylinder.
The belt-type continuously variable transmission for a vehicle according to claim 1 or 2, wherein the belt-type continuously variable transmission for a vehicle is formed to be larger than the flow cross-sectional area on the inflow side for supplying hydraulic oil to the one hydraulic cylinder. Hydraulic control device. (4) The speed change control valve includes a valve body in which a cylinder bore is formed, and a spool valve element having a plurality of lands and slidably fitted into the cylinder bore. At the neutral position, the outflow side flow cross-sectional area is formed by a negative overlap between the land formed on the spool valve and the valve body, and the inflow side flow cross-sectional area is formed by a positive or zero overlap. A hydraulic control device for a belt-type continuously variable transmission for a vehicle according to claim 3. (5) The speed change control valve includes a valve body in which a cylinder bore is formed, and a spool valve element having a plurality of lands and slidably fitted into the cylinder bore. At the neutral position, the outflow side flow cross-sectional area and the inflow side flow cross-sectional area are formed by a negative overlapping state of the land formed on the spool valve and the valve body, and the outflow side flow cross-sectional area is formed. 4. The hydraulic control device for a belt-type continuously variable transmission machine for a vehicle according to claim 3, wherein the negative overlapping state forming the inflow side flow cross-sectional area is larger than the negative overlapping state forming the inflow side flow cross-sectional area.