JP2008038753A - Variable compression ratio device for internal combustion engine - Google Patents

Abstract

In a variable compression ratio device for an internal combustion engine, an operating hydraulic pressure supplied from a hydraulic source to a hydraulic actuator is changed according to a temperature change of the operating oil to prevent an operation delay of the hydraulic actuator at a low temperature of the operating oil, When hydraulic fluid is hot, excessive operation of the hydraulic source is suppressed to prevent unnecessary energy consumption from increasing.
SOLUTION: The compression ratio of an internal combustion engine E is switched between a high compression ratio and a low compression ratio, and hydraulic actuators 20 ₁ and 20 ₂ that hold the switched compression ratio, and the operation of the hydraulic actuators 20 ₁ and 20 ₂ . In the variable compression ratio device comprising an electromagnetic switching valve 45 for controlling the pressure and an oil pump 46 for supplying the hydraulic pressure to the hydraulic actuators 20 ₁ , 20 ₂ via the switching valve 45, the hydraulic pressure is changed to the hydraulic oil temperature. The electronic control unit 50 is provided to adjust so as to decrease in accordance with the increase of.
[Selection] Figure 1

Description

  The present invention switches the compression ratio of an internal combustion engine between a high compression ratio and a low compression ratio, maintains a hydraulic compression actuator that maintains the switched compression ratio, a switching valve that controls the operation of the hydraulic actuator, and a switching valve. It is related with the improvement of the compression ratio variable apparatus of an internal combustion engine provided with the hydraulic pressure source which supplies working hydraulic pressure to a hydraulic actuator via this.

Such a variable internal combustion engine compression ratio device switches the compression ratio of the internal combustion engine according to its operating conditions, thereby improving combustion efficiency while avoiding abnormal combustion such as knocking, improving output and improving fuel efficiency. Various types have already been known for reducing the above (see Patent Documents 1 to 4).
JP 2005-54619 A JP-A 64-15438 JP-A-9-228858 JP-A-60-142020

  In a conventional variable compression ratio device for an internal combustion engine, the hydraulic pressure supplied from the hydraulic source to the hydraulic actuator is set constant regardless of the temperature change of the hydraulic fluid.

  By the way, since the viscosity of hydraulic oil increases as the temperature decreases, when the hydraulic pressure is low, the hydraulic oil transmission speed decreases due to the high viscosity of the hydraulic oil when the hydraulic oil is at a low temperature. A large operating delay occurs in the actuator. Therefore, if the set value of the hydraulic pressure is increased to eliminate such operation delay, the hydraulic oil transmission speed increases more than necessary due to the low viscosity of the hydraulic oil when the hydraulic oil is at a high temperature. Will be over-actuated, leading to the disadvantage of increasing unnecessary energy consumption.

  The present invention has been made in view of such circumstances. The hydraulic pressure supplied from the hydraulic source to the hydraulic actuator is changed according to the temperature change of the hydraulic oil, and the hydraulic actuator is operated when the hydraulic oil is at a low temperature. It is an object of the present invention to provide a variable compression ratio device for an internal combustion engine that can prevent delays and suppress excessive operation of a hydraulic pressure source when hydraulic oil is hot to prevent unnecessary energy consumption from increasing.

  In order to achieve the above object, the present invention switches the compression ratio of an internal combustion engine between a high compression ratio and a low compression ratio, controls a hydraulic actuator that maintains the switched compression ratio, and controls the operation of the hydraulic actuator. In a variable compression ratio device for an internal combustion engine, comprising a switching valve and a hydraulic pressure source that supplies a hydraulic pressure to a hydraulic actuator via the switching valve, the hydraulic pressure is decreased as the hydraulic oil temperature increases. It is a first feature that a control means for adjusting to the above is provided.

Incidentally, the hydraulic source, corresponds to the oil pump 46 in the embodiment of the present invention to be described later, also the hydraulic actuator first, the second hydraulic actuator 20 _1, 20 _2, switching valve electromagnetic switching valve 45 The control means corresponds to the electronic control unit 50.

  According to the present invention, in addition to the first feature, the control means divides the operating hydraulic pressure into a switching hydraulic pressure for switching the compression ratio and a holding hydraulic pressure for holding the switched compression ratio, The second feature is that the hold oil pressure is adjusted to be smaller than the switching oil pressure in the low temperature region and the high temperature region of the hydraulic oil temperature.

  Further, in addition to the second feature, the present invention has a third feature that the control means adjusts the hold hydraulic pressure to substantially the same pressure as the switching hydraulic pressure in an excessively high temperature range of the hydraulic oil temperature.

  Furthermore, in addition to any of the first to third features, the present invention has a fourth feature that the control means is configured to estimate the hydraulic oil temperature from the engine temperature and the load.

  According to the first feature of the present invention, the hydraulic pressure supplied to the hydraulic actuator is decreased in accordance with the increase of the hydraulic oil temperature, so that the hydraulic pressure is increased in the low temperature region of the hydraulic oil temperature. Therefore, despite the high viscosity of the hydraulic oil, it is possible to reduce the delay in the hydraulic actuator operation and ensure a good switching response of the compression ratio, and to reduce the hydraulic pressure in the high hydraulic oil temperature range. As a result, the load on the hydraulic power source is reduced and energy consumption is suppressed as much as possible, while the low viscosity of the hydraulic oil is used to prevent the hydraulic actuator from delaying its operation and to ensure a good switching response of the compression ratio. it can.

  According to the second feature of the present invention, when the compression ratio is maintained in the low temperature region and the high temperature region, the hydraulic pressure of the hydraulic actuator is considered to be relatively small, and the hold hydraulic pressure is adjusted to be smaller than the switching hydraulic pressure. The source load can be further reduced to reduce energy consumption.

  According to the third feature of the present invention, in the excessively high temperature range of the hydraulic oil temperature, the hold hydraulic pressure is substantially the same as the switching hydraulic pressure in anticipation of an increase in the amount of hydraulic oil leaked from the operating actuator due to the lower viscosity of the hydraulic oil. By using the pressure, it is possible to stabilize the holding of the switched compression ratio.

  According to the fourth feature of the present invention, by estimating the hydraulic oil temperature from the engine temperature and the load, the use of a special oil temperature sensor can be omitted and the cost can be reduced.

  Embodiments of the present invention will be described below based on one embodiment of the present invention shown in the accompanying drawings.

  FIG. 1 is a longitudinal sectional front view of an essential part of an internal combustion engine equipped with a variable compression ratio device according to the present invention, and FIG. 2 is an enlarged sectional view taken along line 2-2 of FIG. 3 shows a high compression ratio state, corresponding to FIG. 2, FIG. 4 is a sectional view taken along the line 4-4 in FIG. 2, FIG. 5 is an enlarged sectional view taken along the line 5-5 in FIG. Fig. 7 is a sectional view taken along line -6, Fig. 7 is a sectional view taken along line 7-7 of Fig. 3, Fig. 8 is an explanatory diagram of the switching action from the high compression ratio state to the low compression ratio state, and FIG. 10 is a diagram showing the relationship between hydraulic oil temperature, hydraulic oil viscosity, and hydraulic oil leakage amount. FIG. 11 is a diagram showing the adjustment of the switching hydraulic pressure and hold hydraulic pressure with respect to changes in hydraulic oil temperature. 12 is a diagram showing a compression ratio switching region map, FIG. 13 is a diagram showing a hydraulic oil temperature estimation map, FIG. 14 is a diagram showing a switching oil pressure setting map, and FIG. 15 is a diagram showing a switching oil pressure continuous cycle map, 16 is a diagram showing a decompression coefficient setting map, and FIG. It is over chart.

  First, the basic configuration of the variable compression ratio device for the internal combustion engine E will be described with reference to FIGS.

  1 to 3, in the internal combustion engine E, a crankshaft 9 supported by the crankcase 3 via bearings 8 and 8 ′ is connected to a piston 5 that moves up and down in the cylinder bore 2 a of the cylinder block 2 via a connecting rod 7. Thus, the piston 5 is connected. The piston 5 is slidably fitted to a piston inner 5a connected to the small end portion 7a of the connecting rod 7 via a piston pin 6 and the outer peripheral surface of the piston inner 5a. The piston outer 5b is movable between a predetermined low compression ratio position L (see FIG. 2) and a high compression ratio position H (see FIG. 3), and the piston outer 5b is mounted on the outer periphery thereof. A plurality of piston rings 10 a to 10 c are slidably fitted to the inner peripheral surface of the cylinder bore 2 a and the head portion 5 bh faces the combustion chamber 4 a of the cylinder head 4.

  A plurality of spline teeth 11a and spline grooves 11b that extend in the axial direction of the piston 5 and engage with each other are formed on the sliding fitting surfaces of the piston inner and outer 5a, 5b. The piston inner and outer 5a, 5b , The relative rotation around these axes is not possible. A stop ring 18 that restricts the relative movement of the piston inner 5a and the piston outer 5b in the axial direction is locked to the inner peripheral surface of the piston outer 5b opposite to the head portion 5bh across the piston inner 5a.

Between the piston inner 5a and the head portion 5BH, it is the first cam mechanism 15 ₁ is Kai裝to first control the axial spacing S ₁ between them, also between the piston inner 5a and retaining ring 18, A second cam mechanism 15 ₂ for controlling the second axial interval S ₂ therebetween is interposed. By increasing or decreasing the first and second axial distances S ₁ and S ₂ opposite to each other by the first and second cam mechanisms 15 ₁ and 15 ₂ , the piston outer 5b can move the piston pin with respect to the piston inner 5a. The low compression ratio position L close to and the high compression ratio position H close to the combustion chamber 4a are alternately held.

As shown in FIGS. 2, 3 and 6, the first cam mechanism 15 ₁ includes a first fixed cam 16 ₁ of the upper to be formed in the head portion 5bh inner wall of the piston outer 5b, integrally on the upper surface of the piston inner 5a consists projecting from the pivot portion 12 One fitted rotatablyゝfirst rotating cam plate lower part is supported on the upper surface of the piston inner 5a 17 ₁ Tokyo on.

First rotating cam plate 17 _1, the first and second rotational position A is set around its axis, as it is capable of rotating between B, and cooperates ₁ and the first fixed cam 16 by its reciprocating rotational, said first axial spacing S ₁ may increase or decrease. Specifically, the first fixed cam 16 ₁ is composed of a plurality of cam peaks 16 ₁ , 16 ₁ ... Arranged in the circumferential direction, and the first rotating cam plate 17 ₁ has a plurality of cam peaks also arranged in the circumferential direction. 17 ₁ , 17 ₁ ... Are integrally formed.

Thus, when the first rotary cam plate 17 ₁ is in the first rotational position A, the upper first fixed cam is located in the valley between the adjacent cam peaks 17 ₁ , 17 _{1 of} the first rotary cam plate 17 _1. 16 ₁ of the cam nose 16 ₁ are possible out (see (a), (b) in FIG. 8), as a result, transition to a low compression ratio position L or high compression ratio position H of the piston outer 5b is allowed The And if the upper and lower cam lobes 16 ₁ a, 17 ₁ a is Kamiae, the first cam mechanism 15 ₁ will reduce the axial contraction state is in the first axial spacing S _1.

Also when the first rotating cam plate 17 ₁ is in the second rotational position B, the first rotating cam plate 17 ₁ and the first fixed cam 16 ₁ cam lobes 16 ₁ a, 17 ₁ a to each other the flat top surface to abut (see FIG. 8 (a)) that is, the first cam mechanism 15 ₁ is a axial expanded state, the increased first axial spacing S _1, a high compression ratio position of the piston outer 5b Will be held at H.

The piston inner 5a, and the first between rotating cam plate 17 _1, a first actuator 20 ₁ is rotated alternately first rotating cam plate 17 ₁ to the first rotational position A and the second rotational position B is provided.

First actuator 20 ₁ is configured as shown in FIGS. 2 and 4. That is, the piston inner 5a, a piston pin 6 interposed therebetween a pair of bottom extending parallel therewith the cylinder bore 21 _1, 21 _1, the long hole 29 ₁ extending through the upper wall of the intermediate portion of each cylinder bore 21 ₁ 29 ₁ and a pair of pressure receiving pins 28 ₁ , 28 _{1 that} are integrally projected on the lower surface of the first rotating cam plate 17 ₁ and are arranged on the diameter line thereof through the long holes 29 ₁ , 29 ₁ . It faces the cylinder holes 21 ₁ and 21 ₁ . The long holes 29 ₁ and 29 ₁ do not prevent the pressure receiving pins 28 ₁ and 28 ₁ from moving between the first rotation position A and the second rotation position B together with the first rotation cam plate 17 ₁ .

An operating plunger 23 ₁ and a bottomed cylindrical return plunger 24 ₁ are slidably fitted in each cylinder hole 21 ₁ with a corresponding pressure receiving pin 28 ₁ interposed therebetween. At this time, the operating plungers 23 ₁ , 23 ₁ and the return plungers 24 ₁ , 24 ₁ are arranged symmetrically with respect to the axis of the piston 5.

In each cylinder hole 21 ₁ , a first hydraulic chamber 25 ₁ is formed in which the inner end of the actuating plunger 23 ₁ faces. When hydraulic pressure is supplied to the chamber 25 ₁ , the actuating plunger 23 ₁ receives pressure and receives the pressure. The first rotary cam plate 17 ₁ is rotated to the second rotational position B via the pin 28 ₁ .

Also the open end of each cylinder bore 21 _1, cylindrical spring retaining tube 35 ₁ is engaged through the retaining ring 36 _1, between the plunger 24 ₁ returning said this spring holding cylinder 35 ₁ spring 27 _first return to the back biasing the plunger 24 ₁ on the pressure receiving pin 28 ₁ side is provided in a compressed state.

And Thus, a first rotation position A of the first rotating cam plate 17 ₁ is defined by the tip pressure pin 28 ₁ of the actuating plunger 23 ₁ abutting the bottom surface of each cylinder bore 21 ₁ is in contact, the first second rotational position B of the rotation cam plate 17 ₁ is defined by abutting the tip of the plunger 24 ₁ spring holding cylinder 35 ₁ back pressed to the pressure receiving pin 28 _1.

Also as shown in FIGS. 2, 3 and 6, the second cam mechanism 15 ₂ includes a second fixed cam 16 _{and second} upper portion formed on the lower end wall of the piston inner 5a, the piston outer on the stop ring 18 5b has a second rotating cam plate 17 ₂ which the lower portion rotatably fitted to the inner peripheral surface of the inner periphery of the piston outer 5b, abutting annular upper surface of the second rotating cam plate 17 ₂ and the shoulder portion 19 is formed, the shoulder portion 19 and the stop ring 18 and the second rotating cam plate 17 ₂ is rotatably clamped, the axial movement is prevented with respect to the piston outer 5b.

The second rotating cam plate 17 _2, as it can rotate between the third rotation position C and the fourth rotational position D is set around its axis, in cooperation with the ₂ second fixed cam 16 by its reciprocating rotational , The second axial distance S ₂ is increased or decreased. Specifically, the second fixed cam 16 ₂ is composed of a plurality of cam peaks 16 ₂ a, 16 ₂ a,... Arranged in the circumferential direction, and the second rotating cam plate 17 ₂ has a plurality of the same in the circumferential direction. Cam ridges 17 ₂ a, 17 ₂ a... Are integrally formed.

The rotation angle between the third and fourth rotation positions C and D of the second rotation cam plate 17 ₂ is the same as the rotation angle between the first and second rotation positions A and B of the first rotation cam plate 17 ₁ . Is set. The least effective height of the second fixed cam 16 ₂ and the second rotating cam plate 17 _{and second} cam lobes 16 ₂ a, 17 ₂ a, the first fixed cam 16 ₁ and the first rotating cam plate 17 _first cam nose It is set to be the same as that of 16 ₁ a and 17 ₂ a.

Thus, when the second rotary cam plate 17 ₂ is at the third rotational position C, the cam peaks 16 ₂ a and 17 ₂ a of the second rotary cam plate 17 ₂ and the second fixed cam 16 ₂ are flat. thereby abutting the top surface (see FIG. 8 (d)) that is, the second cam mechanism 15 ₂ is a axial expanded state, increasing the second axial spacing S _2, the piston outer 5b low Hold at compression ratio position L.

The ₂ second rotating cam plate 17 when in the fourth rotational position D, and the valleys between the cam lobes 17 ₂ a, 17 ₂ a adjacent the second rotating cam plate 17 of the _second fixed cam 16 ₂ The cam crest 16 ₂ a can enter and exit (see (a) and (c) of FIG. 8), and as a result, the piston outer 5b is allowed to shift to the low compression ratio position L or the high compression ratio position H. When the upper and lower cam peaks 16 ₂ a and 17 ₂ a are engaged with each other, the second cam mechanism 15 ₂ is in the axially contracted state, and the second axial interval S ₂ is reduced.

Between the piston inner 5a and the second rotating cam plate 17 _2, the second actuator 20 ₂ for rotating alternately the second rotating cam plate 17 ₂ to the third rotation position C and the fourth rotational position D is provided.

Second actuator 20 ₂ has a structure as shown in FIGS. 2 and 6. That is, the piston inner 5a, the piston pin 6 interposed therebetween pair of bottomed cylinder bore extending parallel therewith 21 _2, 21 ₂ and a long hole 29 ₂ extending through the upper wall of the intermediate portion of each cylinder bore 21 ₂ , 29 ₂ and is provided, on the lower surface of the second rotating cam plate 17 ₂ is integrally projected a pair of pressure receiving pins 28 arranged on a diameter line _2, 28 ₂ through the slots 29 _2, 29 ₂ It faces the cylinder holes 21 ₂ and 21 ₂ . Each long hole 29 ₂ does not prevent the pressure receiving pin 28 ₂ from moving between the third rotation position C and the fourth rotation position D together with the second rotation cam plate 17 ₂ .

An operating plunger 23 ₂ and a bottomed cylindrical return plunger 24 ₂ are slidably fitted in each cylinder hole 21 ₂ with a corresponding pressure receiving pin 28 ₂ interposed therebetween. At this time, the operating plungers 23 ₂ , 23 ₂ and the return plungers 24 ₂ , 24 ₂ are arranged symmetrically with respect to the axis of the piston 5.

In each cylinder hole 21 ₂ , a second hydraulic chamber 25 ₂ is formed in which the inner end of the actuating plunger 23 ₂ faces. When hydraulic pressure is supplied to the chamber 25 ₂ , the actuating plunger 23 ₂ receives the hydraulic pressure and receives the pressure. The second rotary cam plate 17 ₂ is rotated to the fourth rotational position D via the pin 28 ₂ .

A cylindrical spring holding cylinder 35 ₂ is locked to the open end of each cylinder hole 21 ₂ via a retaining ring 36 ₂ , and between the spring holding cylinder 35 ₂ and the return plunger 24 _2. spring 27 ₂ return to its back biases the plunger 24 ₂ to the pressure receiving pin 28 ₂ side is provided in a compressed state. Thus, the second actuator 20 ₂ is configured symmetrically with the first actuator 20 ₁ .

Thus, the third rotational position C of the second rotary cam plate 17 ₂ is such that the pressure receiving pins 28 ₂ , 28 ₂ are provided at the tips of the operating plungers 23 ₂ , 23 _{2 that} are in contact with the bottom surfaces of the cylinder holes 21 ₂ , 21 _2. The fourth rotation position D of the second rotating cam plate 17 ₂ is defined by the return plunger 24 ₂ pushed by the pressure receiving pin 28 ₂ contacting the tip of the spring holding cylinder 35 _2. .

In the above, the first rotating cam plate 17 ₁ and the first actuator 20 ₁ , and the second rotating cam plate 17 ₂ and the first actuator 20 ₂ are different from each other in the inertia force difference between the piston inner 5a and the piston outer 5b, and the piston outer 5b. Friction force received from the inner surface of the cylinder bore 2a, negative pressure, positive pressure, etc. received by the piston outer 5b from the combustion chamber 4a side, etc. The piston outer 5b is allowed to move between the low compression ratio position L and the high compression ratio position H.

1 and 2 again, a cylindrical oil chamber 41 is defined between the piston pin 6 and the sleeve 40 press-fitted into the hollow portion thereof. The oil chamber 41 is defined as the first and second actuators 20. _First and second distribution oil passages 42 ₁ and 42 ₂ connected to both hydraulic chambers 25 ₁ and 25 ₂ of ₁ and 20 ₂ are provided across the piston pin 6 and the piston inner 5a. The oil chamber 41 is connected to an oil passage 44 provided across the piston pin 6, the connecting rod 7 and the crankshaft 9. The oil passage 44 is connected to an oil pump 46 serving as a hydraulic pressure source via an electromagnetic switching valve 45 and an oil sump. 47 is switchably connected. The oil pump 46 is driven by an electric motor 48.
[Switching from high compression ratio position to low compression ratio position]
Now, it is assumed that the piston outer 5b is held at the high compression ratio position H as shown in FIG. Therefore, in the first cam mechanism 15 ₁ , the upper and lower cam peaks 16 ₁ a and 17 ₁ a are in the axially expanded state with the top surfaces facing each other, and in the second cam mechanism 15 ₂ , the upper and lower cam peaks 16 ₂ a, 17 ₂ a are in an axially contracted state in which the _two a mesh with each other.

In this state, when the electromagnetic switching valve 45 is in a non-energized state, that is, in an off state as shown in FIG. 1, and the oil passage 44 is opened to the oil sump 47, the hydraulic pressures of the first and second actuators 20 ₁ and 20 ₂ chamber 25 _1, 25 _2, since both are open to the oil reservoir 47 through the oil chamber 41 and the oil passage 44, first the actuator 20 _1, returns the plunger 24 ₁ back pressure pin 28 by the biasing force of the spring 27 _{1 1} is pressed, and the first rotating cam plate 17 ₁ attempts to rotate in the first rotational position a, the second actuator 20 _1, the pressure receiving pin 28 ₂ by the biasing force of the plunger 24 ₂ return spring 27 ₂ return pressing and, to rotate the second rotating cam plate 17 ₂ to the third rotational position C.

Therefore, the piston 5 moves to the intake stroke, the piston inner 5a, since the downward inertial force prior to the piston outer 5b acts, the first cam mechanism 15 _1, between the piston inner 5a and the piston outer 5b Freed from thrust loads. Therefore, first, the first rotating cam plate 17 ₁ is quickly rotated into the first rotational position A via the pressure pin 28 ₁ by the urging force of the spring 27 _first return of the first actuator 20 _1. As a result, as shown in FIG. 8 (b), the first cam mechanism 15 ₁ of the upper and lower cam lobes 16 ₁ a, 17 ₁ a is allowed arrangement meshing is shifted by a half pitch from each other.

Next, when the piston 5 comes in the latter half of the compression stroke, an upward inertial force acts on the piston inner 5a prior to the piston outer 5b, so that the piston outer 5b is as shown in FIG. while meshing the first cam mechanism 15 ₁ of the upper and lower cam lobes 16 ₁ a, 17 ₁ a to each other, i.e., while the first cam mechanism 15 ₁ is contracted in the axial direction, and lowered relative piston inner 5a, It occupies the low compression ratio position L.

With such a piston outer 5b is lowered relative piston inner 5a, will be in the second cam mechanism 15 _2, the second rotating cam plate 17 ₂ to the second fixed cam 16 ₂ descends, and since accompanied no upper and lower cam lobes 16 ₂ a, 17 ₂ a is released from the engagement state, the second rotating cam plate 17 ₂ to the pressure receiving pin 28 ₂ by the urging force of the spring 27 ₂ return of the second actuator 20 ₂ Through the third rotation position C. As a result, as shown in FIG. 8D, the upper and lower cam ridges 16 ₂ a and 17 ₂ a of the second cam mechanism 15 ₂ abut the flat top surfaces against each other. By such an axial expansion action of the second cam mechanism 15 ₂ , the second axial interval S ₂ is increased and the low compression ratio position L of the piston outer 5b is maintained, and the internal combustion engine E is subjected to low compression. It becomes a specific state.
[Switching from low compression ratio position to high compression ratio position]
Next, when the internal combustion engine E is operated at a high speed, the electromagnetic switching valve 45 is energized, that is, turned on, and the oil passage 44 is connected to the oil pump 46. As a result, the discharge hydraulic pressure of the oil pump 46 is changed to the oil passage 44 and oil chamber 41. since applied to full hydraulic chamber 25 _1, 25 ₂ through, the first actuator 20 _1, the first rotating cam plate via the pressure pin 28 ₁ actuating plunger 23 ₁ receives the first oil pressure chamber 25 ₁ of the hydraulic 17 ₁ attempts to rotate toward the second rotational B, the second actuator 20 _2, actuating plunger 23 ₂ via the pressure receiving pin 28 ₂ receives the second hydraulic chamber 25 _{and second} hydraulic second rotating cam plate 17 ₂ is to be rotated toward the fourth rotational position D.

Therefore, the piston 5 moves to the exhaust stroke, for receiving the upward inertial force piston inner 5a is in advance of the piston outer 5b, the second cam mechanism 15 ₂ is interposed between the piston inner 5a and retaining ring 18 Freed from thrust loads. Therefore, first, the second rotating cam plate 17 ₂ is quickly rotated to the fourth rotational position D via the pressure receiving pin 28 ₂ by the pressing force of the hydraulic plunger 23 ₂ of the second actuator 20 ₂ . As a result, as shown in FIG. 9B, the upper and lower cam peaks 16 ₂ a and 17 ₂ a of the second cam mechanism 15 ₂ are arranged so as to be able to engage with each other with a half-pitch shift.

Next, when the piston 5 comes in the latter half of the intake stroke, a downward inertial force acts on the piston inner 5a in advance of the piston outer 5b, so that the piston outer 5b is as shown in FIG. while meshing the second cam mechanism 15 and _second upper and lower cam lobes 16 ₂ a, 17 ₂ a from each other, i.e., while the second cam mechanism 15 ₁ is contracted in the axial direction, relative to rise relative to the piston inner 5a, The high compression ratio position H will be occupied.

With such a piston outer 5b is relatively raised with respect to the piston inner 5a, will be the first cam mechanism 15 _1, the second fixed cam 16 ₁ with respect to the first rotating cam plate 17 ₁ rises, and since accompanied no upper and lower cam lobes 16 ₁ a, 17 ₁ a is released from the engagement state, the first rotating cam plate 17 ₁ is receiving pin 28 by the pressing force of the first actuator 20 ₁ of the working plunger 23 _first hydraulic ₂ to quickly rotate to the second rotational position B (see FIG. 5). As a result, as shown in FIG. 8 (d), the first cam lobes of the upper and lower cam mechanism _{15 1 16 1 a, 17 2} a is brought into contact opposite to each other flat top surface. The axial expansion effect of such a first cam mechanism 15 _1, a first axial spacing S ₁ is increased, it will retain a high compression ratio position H of the piston outer 5b. Thus, the internal combustion engine E is in a high compression ratio state. The configuration of the variable compression ratio device follows that disclosed in Patent Document 1.

  As shown in FIG. 1, an electronic control unit 50 that controls the operation of the electromagnetic switching valve 45 is connected.

By the way, the temperature of the hydraulic fluid (hereinafter referred to as hydraulic fluid temperature) supplied from the electromagnetic switching valve 45 to the first and second actuators 20 ₁ and 20 ₂ , the viscosity of the hydraulic fluid, and the first and second actuators 20 _1. , 20 ₂ has a relationship as shown in FIG. That is, in the low temperature region A of the hydraulic oil, the hydraulic oil has a high viscosity, so the hydraulic pressure transmission speed is low and the hydraulic oil leakage amount is small. Further, in the high temperature region B of the hydraulic oil, the hydraulic pressure transmission speed and the hydraulic oil leakage amount increase due to a decrease in the viscosity of the hydraulic oil. Further, in the excessively high temperature region C of the hydraulic oil, the hydraulic pressure transmission speed and the hydraulic oil leakage amount further increase due to further decrease in the viscosity of the hydraulic oil.

In consideration of the characteristics of the hydraulic pressure transmission speed and hydraulic oil leakage amount with respect to the hydraulic oil temperature, the electronic control unit 50 operates hydraulic pressures of the first and second actuators 20 ₁ and 20 ₂ as shown in FIG. Is divided into a switching hydraulic pressure for switching to a high compression ratio and a holding hydraulic pressure for maintaining a high compression ratio, and the operation of the oil pump 46 is performed to adjust each hydraulic pressure according to changes in the operating oil temperature. Is to control.

That is, the switching oil pressure supplied to the first and second actuators 20 ₁ and 20 ₂ when the electromagnetic switching valve 45 is on changes when the hydraulic oil temperature changes to the low temperature region A, the high temperature region B, and the excessively high temperature region C. Taking into account the increase in hydraulic fluid transmission speed, adjustment is made sequentially to large, medium, and small. In addition, when the hydraulic oil temperature changes to the low temperature region A and the high temperature region B, the hold hydraulic pressure is adjusted to a hydraulic pressure that is lower than the switching hydraulic pressure in the corresponding temperature region in consideration of an increase in the hydraulic fluid transmission speed. Further, when the hydraulic oil temperature reaches the excessively high temperature region C, the hold hydraulic pressure is adjusted to substantially the same pressure as the switching hydraulic pressure in the excessively high temperature region C in consideration of an increase in the amount of hydraulic oil leakage.

As described above, the switching hydraulic pressure supplied to the first and second actuators 20 ₁ and 20 ₂ is large, medium, small depending on the change of the hydraulic oil temperature to the low temperature region A, high temperature region B, and excessively high temperature region C. In the low temperature region A, by increasing the switching hydraulic pressure, the operation delay of the first and second hydraulic actuators 20 ₁ and 20 ₂ is reduced despite the high viscosity of the hydraulic oil, Switching response with a good compression ratio can be ensured.

In the high temperature region B and the excessively high temperature region C of the hydraulic oil temperature, the switching hydraulic pressure is decreased sequentially. Therefore, the low viscosity of the hydraulic oil is utilized while reducing the load on the oil pump 46 as a hydraulic pressure source and minimizing energy consumption. Therefore, it is possible to prevent the operation delay of the first and second hydraulic actuators 20 ₁ and 20 ₂ and still ensure a good switching response of the compression ratio.

Also when a high compression ratio maintained at a low temperature area A and the high temperature region B, first, second hydraulic actuator 20 _1, 20 ₂ of the load in view of the relatively small, the hold pressure, switching of the corresponding temperature area Since the pressure is lower than the hydraulic pressure, the load on the oil pump 46 can be further reduced and the energy consumption can be suppressed.

In addition, when the high compression ratio is maintained in the excessively high temperature region C, the holding oil pressure is increased by expecting an increase in the amount of hydraulic fluid leakage from the _first and second hydraulic actuators 20 ₁ and 20 ₂ due to the low viscosity of the hydraulic fluid. Since the switching hydraulic pressure in the high temperature region C is adjusted to substantially the same pressure, the high compression ratio can be maintained regardless of the increase in the amount of hydraulic fluid leakage.

  In order to adjust the hydraulic pressure, the electronic control unit 50 includes a rotational speed sensor 51 that detects the engine rotational speed Ne, a throttle sensor 52 that detects the opening θ of the throttle valve of the engine E, and the temperature of the engine E ( For example, an output signal from a temperature sensor 53 or the like that detects a coolant temperature (Tw) is input.

  The electronic control unit 50 includes a compression ratio switching region map M1 (FIG. 12), a hydraulic oil temperature estimation map M2 (FIG. 13), a switching oil pressure setting map M3 (FIG. 14), and a switching oil pressure continuation cycle map M4 (FIG. 15). And a decompression coefficient setting map M5 (FIG. 16).

  Next, the control programming execution procedure of the electronic control unit 50 will be described with reference to the flowchart of FIG.

  First, in step S1, the engine speed Ne is read from the output signal of the speed sensor 51, the engine temperature Tw is read from the output signal of the temperature sensor 53, and the engine E is output from the engine speed Ne and the output signal θ of the throttle sensor 52. Load (torque) Tq is calculated and read.

  In step S2, the compression ratio switching region is determined by the compression ratio switching region map M1 (FIG. 12) based on the engine speed Ne and the load Tq. In step S3, is the determination region in step S2 a low compression ratio region? If YES (that is, if it is determined that the compression ratio range is low), the process proceeds to step 4 and the electromagnetic switching valve 45 is turned off.

  If the determination in step S3 is NO (that is, if it is determined as a high compression ratio region), the process proceeds to step 5, where it is determined whether or not the determination region in step S2 has changed from the previous time. If YES, the hydraulic oil temperature is retrieved from the hydraulic oil temperature estimation map M2 (FIG. 13) based on the engine speed Ne and load Tq in step 6, and then switched on the basis of the hydraulic oil temperature and engine speed Ne in step 7. The switching oil pressure is retrieved from the oil pressure setting map M3 (FIG. 14), and then the operation of the oil pump 46 is controlled to match the discharge pressure of the oil pump 46 with the switching oil pressure retrieval value in step 8, and then the electromagnetic switching is performed in step 9. The valve 45 is turned on.

Thus, the high compression ratio switching hydraulic pressure supplied from the oil pump 46 to the first and second actuators 20 ₁ and 20 ₂ when the electromagnetic switching valve 45 is turned on has the operating oil temperature in the low temperature region A and the high temperature as described above. When the region B and the overheated region C are changed, the control is sequentially performed in the order of large, medium, and small.

  Next, when the process proceeds from step 9 to step 10, the switching oil pressure continuation cycle is retrieved from the switching oil pressure continuation cycle map M4 (FIG. 15) based on the hydraulic oil temperature and the engine speed Ne, and is input to the oil pressure continuation counter M5. In step 11, the hydraulic pressure continuation counter 54 starts counting and returns to step 1.

If the output of the hydraulic pressure continuation counter 54 becomes zero after step 1 to step 5 again after this, the routine proceeds to step 13 where the pressure change coefficient setting map M6 (FIG. 16) is changed to the previous switching hydraulic pressure. The hold hydraulic pressure to be supplied to the first and second actuators 20 ₁ and 20 ₂ is set by multiplying the pressure reduction coefficient R retrieved from (1), and the operation of the oil pump 46 is controlled accordingly.

  In the decompression coefficient setting map M6 (FIG. 16), the decompression coefficient R increases from a predetermined value (0.5 in the illustrated example) as the hydraulic oil temperature shifts to the low temperature region A and the high temperature region B. When entering the overheated region C, it is 1. Thus, the hold hydraulic pressure after switching the high compression ratio is adjusted to a hydraulic pressure that is lower than the switching hydraulic pressure in the corresponding temperature region when the hydraulic oil temperature is in the low temperature region A and the high temperature region B as described above. When the oil temperature enters the excessively high temperature region C, it is controlled to the same pressure as the switching oil pressure in the excessively high temperature region C.

The present invention is not limited to the above embodiment, and various design changes can be made without departing from the scope of the invention. For example, when the first and second actuators 20 ₁ and 20 ₂ are supplied with hydraulic pressure from the oil pump 46, the variable compression ratio device is configured to drive the piston inner 5a from the high compression ratio position to the low compression ratio position. You can also The adjustment of the hydraulic pressure can be performed by a pressure regulating valve provided in the downstream oil passage of the oil pump 46, instead of the method based on the operation control of the oil pump 46. The hydraulic oil temperature used for hydraulic pressure adjustment by the electronic control unit 50 can also be detected by a temperature sensor provided in the hydraulic oil path. The present invention is applicable not only to those described in Patent Document 1, but also to those described in Patent Documents 2 to 4.

The principal part longitudinal cross-sectional front view of an internal combustion engine provided with the compression ratio variable apparatus of this invention. A low compression ratio state is shown in the enlarged sectional view taken along line 2-2 in FIG. FIG. 3 is a diagram corresponding to FIG. 4-4 sectional drawing of FIG. FIG. 5 is an enlarged sectional view taken along line 5-5 in FIG. FIG. 6 is a sectional view taken along line 6-6 of FIG. FIG. 7 is a cross-sectional view taken along line 7-7 in FIG. 3. Switching operation explanatory diagram from a high compression ratio state to a low compression ratio state. FIG. 4 is a switching operation diagram from a low compression ratio state to a high compression ratio state. The relationship line figure of hydraulic oil temperature, hydraulic oil viscosity, and hydraulic oil leak amount. The diagram which shows the adjustment aspect of the switching hydraulic pressure and hold | maintenance hydraulic pressure with respect to the change of hydraulic fluid temperature. The diagram which shows a compression ratio switching area | region map. The diagram which shows a hydraulic-oil temperature estimation map. The diagram which shows a switching hydraulic pressure setting map. The diagram which shows a switching hydraulic pressure continuation cycle map. The diagram which shows a decompression coefficient setting map. The flowchart which shows the execution procedure of the control programming of an electronic control unit.

Explanation of symbols

E · · · · engine 20 _1, 20 ₂ ... hydraulic actuators (first, second hydraulic actuator)
45 ... Switching valve (electromagnetic switching valve)
46 ... Hydraulic power source (oil pump)
50 ... Control means (electromagnetic control unit)

Claims (4)

It switches the compression ratio of the internal combustion engine (E) in a high compression ratio and a low compression ratio, a hydraulic actuator (20 _1, 20 ₂₎ for holding the switching compression ratio, the hydraulic actuator (20 _1, 20 ₂₎ a control valve for controlling the operation (45), the compression ratio of the switching valve (45) hydraulic actuator through a (20 _1, 20 ₂₎ hydraulic power source for supplying hydraulic pressure to and a (46), an internal combustion engine In the variable device,
A variable compression ratio device for an internal combustion engine, comprising control means (50) for adjusting the hydraulic pressure so as to decrease in accordance with an increase in the hydraulic oil temperature. In the internal combustion engine variable compression ratio device according to claim 1,
The control means (50) divides the hydraulic pressure into a switching hydraulic pressure for switching the compression ratio and a holding hydraulic pressure for holding the switched compression ratio, and a low temperature region (A) of the hydraulic oil temperature and A variable compression ratio device for an internal combustion engine, characterized in that the hold hydraulic pressure is adjusted to be smaller than the switching hydraulic pressure in the high temperature region (B). In the internal combustion engine variable compression ratio device according to claim 2,
The compression ratio variable device for an internal combustion engine, characterized in that the control means (50) adjusts the hold hydraulic pressure to substantially the same pressure as the switching hydraulic pressure in an excessively high operating oil temperature range (C). The compression ratio variable device for an internal combustion engine according to any one of claims 1 to 3,
The compression ratio variable device for an internal combustion engine, characterized in that the control means (50) is configured to estimate the hydraulic oil temperature from the engine temperature and the load.

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