JPH048241Y2 - - Google Patents

Description

[Detailed Description of the Invention] [Industrial Field of Application] This invention relates to the structure of a valve train for an internal combustion engine, and particularly to one in which the set load of a valve spring is made variable in accordance with the engine speed.

[Conventional technology]

It is known that the seating member of the valve spring is configured as a hydraulic actuator, and the set load of the valve spring is made variable by applying engine oil pressure to the hydraulic actuator. (For example, see Japanese Utility Model Application Publication No. 59-41605.) Since the engine oil pressure increases according to the engine speed, the load on the valve spring becomes small at low speeds, and increases at high speeds. Therefore, a sufficient valve spring load can be obtained at high engine speeds while keeping friction loss small at low engine speeds.

[Problem that the invention attempts to solve]

Conventionally, engine oil pressure has been used to drive the seating member. When the engine rotates at high speeds, the temperature of the lubricating oil increases and the viscosity of the lubricating oil decreases, so in this case, the oil pressure tends to decrease. As a result, the seating member, which is a hydraulic actuator, is unable to perform the lift necessary to obtain the desired valve spring setting load. As a result, there is a risk of valve surging occurring at high speeds.

This invention was made to solve this problem, and the purpose is to ensure the desired valve spring setting load even when the engine oil pressure decreases.

[Means for solving problems]

According to this invention, the seating member of the valve spring is configured as a hydraulic actuator, and the seating member is used as an engine oil pressure source in a valve train of an internal combustion engine driven by oil pressure from a hydraulic passage communicating with the engine oil pressure source. Separately from this, an independent hydraulic pump means is arranged, and the independent hydraulic pump means has its suction side connected to the engine hydraulic pressure source side, its discharge side connected to the seating member, and is capable of supplying hydraulic pressure from the independent hydraulic pump to the seating member. It has a control means for controlling the supply, and a means for detecting a state in which the viscosity of the lubricating oil decreases at high engine speeds, and the control means controls the supply of lubricating oil to the seating member at high engine speeds when the viscosity of the lubricating oil decreases. A valve train structure for an internal combustion engine is provided that is characterized by supplying hydraulic pressure.

When a state in which the viscosity of the lubricating oil decreases during high engine rotation is detected, the control means supplies hydraulic pressure from an independent hydraulic pump means to the seating member.

〔Example〕

In Fig. 2, 1 is the main body of the internal combustion engine;
For example, it has four cylinders and four combustion chambers 2. In this embodiment, this internal combustion engine is configured as a double overhead camshaft (DOHC) type engine, and the intake valve 3
A camshaft 4 for driving the exhaust valve 5 and a camshaft 6 for driving the exhaust valve 5 are provided. The camshafts 4 and 6 are provided with gears 7 and 8 that mesh with each other at one end thereof, and transmit driving force from one to the other. In addition, gears 7 and 8
It may also be driven by being wound around a timing pulley or the like without having a.

FIG. 1 shows in detail a portion of the valve operating mechanism for the intake valve 2. As shown in FIG. The exhaust valve 3 is similarly configured.
The intake valve 2 consists of a valve stem 11 and a cap section 12. The valve stem 11 is guided by a valve guide 13. The valve head portion 12 is provided at the end of the intake port 1B formed in the cylinder head 1A so that the valve seat 14 can be opened and closed.

Reference numeral 16 denotes a valve lifter, which is disposed so as to be slidable in a valve lifter hole 17 formed in the cylinder head 1A of the engine body 1. Valve lifter 16
A shim 18 is disposed on the upper surface of the intake camshaft 4 and faces the cam 20 on the intake camshaft 4. Since this engine has four cylinders, four cams 20 are installed on the camshafts 4 and 6 at angular intervals of 90°. The upper end of the valve stem 11 faces the lower surface of the valve lifter 16, and the retainer 21 is disposed close to this portion. A tapered surface is formed on the inner surface of the retainer 21, and the retainer 22 is seated on this tapered surface, whereby the retainer 21 is wedged to the valve stem 11. Incidentally, a protrusion is formed on the inner circumference of the valve stem 11 and an annular groove 1 is formed on the outer circumference of the valve stem 11 to accommodate the protrusion.
1A, by which the lever 22 is restrained at a predetermined axial position on the valve stem 11.

The upper end of the valve spring 25 is hung by the retainer 21, and the seating member 26 is disposed at the lower end, thereby applying an upward biasing force to the valve 3 in the drawing. The seating member 26 is configured as a hydraulic actuator, and is disposed within the cylinder member 27 so as to be movable up and down. The upward movement of the seat member 26 is limited by a stop 27'. A hydraulic chamber 30 is formed in an annular shape on the lower surface of the seating member 27, and this hydraulic chamber 3
The seating member 26 is lifted in response to the zero pressure, and as a result the set load of the valve spring 25 is changed.

32 is an oil pan; 33 is an oil pump as an engine oil pressure source driven by the engine;
Reference numeral 34 indicates an oil filter, and 35 indicates a main gear assembly, which constitutes a lubricating oil supply system for the engine and supplies lubricating oil to the crankshaft and valve train. Seating member 26
Hydraulic pressure for driving the engine is branched from this engine lubricating oil supply system. That is, hydraulic pressure is taken out from the main gear rally 35 as shown by arrow X, and the cylinder head 1
It is supplied to a hydraulic passage 36 formed at A. The hydraulic passage 36 communicates with the hydraulic chamber 30 via a passage 36a. Therefore, oil pressure corresponding to the engine speed is applied to the seating member 26. 37 is a pressure regulator, which functions to define the upper limit of the lubricating oil pressure supplied from the pump 33.

According to this invention, an independent pump means is provided for supplying hydraulic pressure to the hydraulic chamber when engine hydraulic pressure becomes insufficient. The pump means are in this example a piston pump and are driven by a camshaft, preferably a driven camshaft, for example an exhaust camshaft 6. The pump means includes a cylindrical piston 41 that is vertically slidably disposed in a cylinder bore 40 formed in the cylinder head 1A, and a pump drive cam 42 that is fixed to the camshaft 6 so as to face the piston 41. As shown in FIG. 2, one or two pump drive cams are positioned on a cam 20 for driving the valve 5. Furthermore, the cam 42 is located at an intermediate angular position of the cam 20 on the intake camshaft 4, which is the drive side shaft. That is, in a four-cylinder engine, the cams 20 are provided at every 90° position, and the cams 42 are provided at 45° positions.

A pump chamber 43 is formed below the piston 41, and a spring 44 for biasing the piston 41 is arranged. The pump chamber 43 has a suction control check valve 45.
It is connected to a main gear gallery 35 of the engine lubricating oil supply system through a discharge control check valve 46 and a hydraulic passage 36 close to the hydraulic chamber 30 through a passage 36b. Furthermore, the pump chamber 43 is connected to the lubricating oil passage 36 via a bypass passage 48, and a control valve 50 is arranged at this connection point. This control valve 5
0 can be configured as a 3-way solenoid valve, and under normal conditions the hydraulic passage 3 is
6. The bypass passage 48 is located so as to communicate with the main gear rally 35, but during operation, the passage 36 and the bypass passage 48 are connected to the main gear rally 35 as shown by the broken line.
and configured to receive increased hydraulic pressure from the pumping means. A relief valve 51 communicates the bypass passage 48 with the oil pan 32.

54 is a control circuit for driving the control valve 50, and is composed of comparators 55, 56, an AND gate 57, and a transistor 58. control valve 50
, the solenoid 50', and the transistor 58 constitute a control means of the present invention, and this control means supplies hydraulic pressure from an independent pump means to the seating member 26 at high engine speeds when the viscosity of the lubricating oil decreases. do the work. Comparator 55 is connected to engine speed sensor 60. The rotation speed sensor 60 generates a signal corresponding to the rotation speed of the crankshaft of the engine, and the comparator 55 outputs "1" when the engine rotation speed N detected by the sensor 60 is equal to _or higher than a predetermined value Nx; When this happens, a logic signal of "0" is generated. Comparator 56 is connected to oil pressure sensor 61. The oil pressure sensor 61 outputs a signal according to the supply pressure of lubricating oil, and the comparator 5
6 generates a logic signal that is "1" when the engine oil pressure P is less than a predetermined value _Px , and "0" when it is greater than the predetermined value Px.
The rotation speed sensor 60, the oil pressure sensor 61, the comparators 55 and 56, and the AND gate 57 constitute means for detecting a state in which the viscosity of the lubricating oil decreases when the engine rotates at high speed. The base of the transistor 58 is connected to the output of the AND gate 57, and the solenoid 50' of the control valve 50 is located in the emitter-collector circuit.

The operation of this invention described above will be explained. When the engine speed N is lower than the predetermined value N _x or when the lubricating oil pressure P is lower than the predetermined value P _x , the comparators 55 and 5
Since a “0” signal is output from either of 6, AND
Gate 57 outputs a signal of "0" and transistor 58 is turned off. Therefore, solenoid 5
0' is demagnetized and the control valve 50 is in the bypass passage 48.
is located so as to communicate with the main gear rally 35. The piston 41 moves up and down by the rotation of the cam 42, but the bypass passage 48 is connected to the main gear rally 35.
Since it is in communication with the pump, it simply rotates idly and no pumping action is performed. On the other hand, the main gear rally 35 communicates with the hydraulic chamber 30 through passages 36 and 36a, and hydraulic pressure acts on the seating member 26 upward in FIG. .

The mounting load F ₀ of the valve spring 25 changes according to its lift Δl as follows: ΔF ₀ =Δl×k. Here k is a spring constant. Therefore, in FIG. 3, if point a is the lowest position of the seating member 26 and point b is the highest position of the seating member 26, the installation load when the valve 3 is seated (zero lift) is the minimum F. It will vary from ₀₁ to a maximum of F ₀₂ . At this time, the change in the spring load with respect to the lift changes as shown by a straight line a-b when Δl=0, and as shown by a straight line c-d when Δl=max.
Now, if the force acting on the seating member 26 based on the lubricating oil pressure entering the hydraulic chamber 30 is Foil, this becomes the cross-sectional area of the seating member 26 x hydraulic pressure. Assuming that the foils are in stages from Foil ₁ to Foil ₅ , the relationship between the spring load and the valve lift y in each state will be explained. The valve lift y is 0 when the cam 20 is not lifted (base circle) as shown in FIG.

1 When Foil<F0 ₁ , Foil cannot lift the seating member 26, so
The seating member 26 assumes the lowermost position as shown at A in FIG. Lift begins at point a in Figure 3 and reaches maximum lift at point b. The amount of load reduction is represented by the rectangle ABCD.

2 When Fo ₁ ≦ Foil < F ₀₂ When the base circle is on, the seating member 26 is at position F, which is between A and C where the spring force is balanced with the foil. Since Foil < Fmax1 when lifting, it is pushed by the spring force during lifting and becomes B. The lift starts at point l, and the load characteristic reaches b from f. The amount of load reduction is represented by the figure efbdc.

3 F ₀₂ ≦Foil<Fmax1 When not lifted, F ₀₂ ≦Foil, so the seating member 26 is in position C, but since Foil<Fmax1, it is pushed back to B when lifted. Therefore, the load curve is c
The lift starts at point, passes through gh, and reaches the maximum lift at b. The spring load reduction amount is rectangular ghbd.

4 Fmax1≦Foil<Fmax2 When not lifted, the seating member 26 is in the C position, but since Foil<Fmax2, the seating member 26 is pushed to the E or B position. Therefore, the load curve becomes cjh, and the spring load reduction amount becomes triangle jhd.

5 Fmax2≦Foil The load curve is C-d when not lifted and at D when lifted.

In the above control, 2) When Fo ₁ ≦ Foil < F ₀₂ , hydraulic oil flows out when the seating member 26 is pushed from E to B during lift, but the hydraulic circuit is connected to the main gear rally 35.
Since it communicates with and is open, there is little possibility that it will flow into the outer cylinder. If the engine is a 16-valve engine, even if it flows in, if two valves lift, it will be distributed to the remaining cylinders and the lift will only be ∆l x 2/14 = ∆l/7, so there is no problem.

The above operations are performed on the condition that the necessary engine oil pressure is obtained. However, when the engine rotates at high speeds, the temperature of the lubricating oil increases and its viscosity decreases. As a result, sufficient oil pressure may not be obtained. In this case, there is a risk of surging. FIG. 5 illustrates this relationship. The solid line represents the standard temperature state. Here, Poil ₁ is −Poil ₅ is Foil ₁ in Figure 3 respectively.
-Compatible with Foil ₅ . For this purpose, when the pressure receiving area of the seating member 26 is S>Fmax2/
The point of Poil ₅ that needs to be set to become Poil ₅ is the point at which the relief valve 37 in FIG. 2 starts to open. When the temperature changes from these standard characteristics, the temperature becomes as low as when the temperature is low. Due to the high temperature side characteristics, there is not enough pressure, and there is a risk of surging of the hull at high speeds. Note that when the temperature is low, the maximum discharge pressure simply shifts to the low rotation side, and the problems that occur when the temperature is high do not occur.

According to the embodiment of this invention, the engine speed N
is larger than the predetermined value N _x and the oil pressure P is smaller than the predetermined value P _x , both comparators 55 and 56 are “1”.
give a signal. Therefore, AND gate 57 is “1”
The transistor 58 is turned on, and the solenoid 50' is energized. the result,
Control valve 50 closes both bypass passage 48 and passage 36. As a result, the piston 41 can perform a pumping action. That is, when the piston is ascending, lubricating oil is sucked into the pump chamber 43 through the check valve 45, and when the piston is descending, the check valve 45 is sucked into the pump chamber 43.
is closed, but since the check valve 46 is opened, the lubricating oil in the pump chamber 43 is forced into the hydraulic chamber 30 via the passages 36b and 36a.

In addition, the set rotation speed when driving the control valve 50
The value of N _x can be around 4000 rpm, for example. This is because, as shown in FIG. 6, the inertia of the valve train increases rapidly around this rotation speed.

On the other hand, the value of the set pressure _P
Set when the pressure at 4000r.pm is below Poil ₆ .

Although the control valve 50 is configured as a solenoid valve in FIG. 2, it can alternatively be configured as a purely mechanical valve responsive to engine oil pressure.

The pump for compensating when the oil pressure decreases does not have to be a piston pump driven by the cam 42 on the camshaft as in the embodiment. When the cam 42 for driving the piston pump 41 is installed on the driven side camshaft 6, the cam 42
The angular position of
is every 90 degrees, so the position is 45 degrees),
Fluctuations in the load applied to the camshaft can be canceled out. That is, the force of the valve spring 25 applied to the drive-side camshaft 4 fluctuates periodically as shown by Fsp in FIG. On the other hand, by providing the piston 41, a force like Fop is applied to the driven side camshaft 6. Since these forces are in opposite phases, they act in a canceling direction, and as a result, rotational fluctuations of the camshaft can be suppressed. Further, in the case where the camshaft is driven by the gears 7 and 8, it is possible to suppress noise generation due to backlash δ of the gears 7 and 8.
Incidentally, in the case where a scissor gear is used for backlash control, there is an advantage that the spring force of the scissor spring can be weakened.

This idea is most suitable when applied to a DOHC type engine as in the embodiment. That is, in the DOHC type, since the inertia force of the valve train is small, the spring load is small, and the hydraulic load for driving the valve spring seating member 26 can be lightened.

[Effect of idea]

According to this invention, by providing an independent pump that is driven at high engine speeds when the viscosity of the lubricating oil decreases, it is possible to supply necessary and sufficient hydraulic pressure to the seating member even when the pressure of the hydraulic pressure source decreases. can.
Therefore, the benefits of low friction and high fuel economy due to the ability to reduce the spring load movement at low rotations are not impaired in the least, and the fear of surging at high rotations can also be eliminated.

[Brief explanation of the drawing]

Fig. 1 is an overall view of the valve train of the present invention, Fig. 2 is a top view of the engine, Fig. 3 is a graph of the valve lift-spring load relationship, and Fig. 4 is the seating member of A when not lifted and B when lifted. Figure 5 showing various positions of
Figure 6 is a graph of the relationship between engine rotation speed and discharge pressure, Figure 6 is a graph of the relationship between engine rotation speed and valve train inertia, Figure 7 is a graph of the relationship between cam rotation angle and force applied to the camshaft, and Figure 8 is a graph of the relationship between cam rotation angle and force applied to the camshaft. is a front view of the gear seen from the direction of FIG. 2. 1...Engine body, 3...Intake valve, 4...Intake camshaft, 5...Exhaust valve, 6...Exhaust camshaft,
7, 8...Gear, 16...Valve lifter, 21...
...Retainer, 25...Valve spring, 26...
... Seating member, 30 ... Hydraulic chamber, 33 ... Oil pump, 41 ... Piston, 42 ... Pump drive cam, 43 ... Pump chamber.

Claims (1)

[Claims for Utility Model Registration] 1. In a valve operating system of an internal combustion engine in which a seating member of a valve spring is configured as a hydraulic actuator, and the seating member is driven by hydraulic pressure from a hydraulic passage communicating with an engine oil pressure source, An independent hydraulic pump means is installed separately from the engine oil pressure source, the suction side of the independent hydraulic pump means is connected to the engine oil pressure source side, the discharge side is connected to the seating member side, and the independent hydraulic pump is connected to the seating member side. and means for detecting a state in which the viscosity of the lubricating oil decreases when the viscosity of the lubricating oil decreases when the engine rotates at high speeds. A valve train structure for an internal combustion engine, characterized by supplying hydraulic pressure to a seating member. 2. The valve train system for an internal combustion engine according to claim 1, wherein the hydraulic pump means is a pump having a reciprocating piston, and the pump is driven by a camshaft of the internal combustion engine. structure. 3. An internal combustion engine has two camshafts, one on the driving side and one on the driven side, and these camshafts are connected in motion by gears, and the cam on the camshaft that drives the piston drives the valve. The valve train structure for an internal combustion engine according to claim 2, wherein the valve train structure is located at an intermediate angular position with respect to the cam. 4. The structure of a valve train for an internal combustion engine according to claim 1, wherein the valve train is of a direct drive type in which a valve lifter is driven by a cam.