2. FIELD OF THE INVENTION AND RELATED ART STATEMENT:
[0001] The present invention relates to an accumulator type valve system for controlling
opening and closing of an exhaust valve or an intake valve for a Diesel engine.
[0002] Figure 8 shows the prior art accumulator type valve system. The system comprises
a five-port directional control valve 01 which controls a hydraulic fluid that is
pressurized by a pump 010 and accumulated to a predetermined pressure in an accumulator
011, a cam 02, a tappet roller and a cam shaft 04 that drive the control valve, an
actuator which moves an intake and exhaust valve 06 provided on a cylinder head 05
by means of the hydraulic fluid, and pipelines 08 and 09. Figure 8 shows the state
in which the tappet roller 03 is on the base circle (namely, the valley) part of the
cam 02, the control valve 01 is introducing the hydraulic fluid to a lower chamber
07a of the actuator 07 through the pipeline 09, and the intake and exhaust valve 06
is closed by being moved upward under the force of the hydraulic fluid. In this situation
the hydraulic fluid in an upper chamber 07b of the actuator 07 is drained into a tank
010a of the pump 010 through the pipeline 08 and the control valve 01. When the tappet
roller 03 is lifted as a result of rotation of the cam 02, the hydraulic fluid in
the lower chamber 07a is drained into the tank 010a through the pipeline 09 and the
control valve 01, and the high pressure hydraulic fluid in the accumulator 011 acts
on the upper chamber 07b via the control valve 01 and the pipeline 08. The force of
the hydraulic fluid moves the intake and exhaust valve 06 downward against the pressure
in the cylinder that acts on the intake and exhaust valve 06, whereby causing the
intake and exhaust valve to open. When the lift of the tappet roller 03 is decreased
and reaches the base circle of the cam by a further rotation of the cam 02, the control
valve 01 takes on the state in the figure and the intake and exhaust valve 06 is closed
by the upward motion as described above.
[0003] As in the above, the prior art apparatus requires five ports for the control valve
01 which makes its structure complex and large-sized, moreover, there are needed two
high pressure tubes 08 and 09 and the actuator 07 is complicated and necessarily large-sized.
In addition, the prior art device makes use of a hydraulic fluid at high pressure
for opening and closing the input and exhaust valve 06, and the high pressure hydraulic
fluid is drained into the tank 010a and discarded after operating the actuator 07,
whereby increasing the consumption of the hydraulic fluid and the energy loss. Moreover,
the prior art device has a problem that it is difficult to control the opening and
closing and the seating of the intake and exhaust valve.
3. OBJECT AND SUMMARY OF THE INVENTION:
[0004] It is the object of the present invention to provide a valve system for an internal
combustion engine which enables to eliminate problems occurring in the prior art system,
to simplify the structures of the control valve and the actuator, to optimally control
the valve closing speed, and to suppress the consumption of the hydraulic fluid to
a necessary minimum and markedly reduce the power loss by performing the valve closing
by means of the spring force.
[0005] In a valve system equipped with an accumulator-incorporated hydraulic power source
which stores the hydraulic fluid by pressuring the fluid, an actuator having a hydraulic
fluid operated piston which drives the intake valve or the exhaust valve, and a control
valve which supplies the hydraulic fluid in the accumulator to the actuator in a controlled
manner, the valve system for the internal combustion engine in accordance with the
present invention is characterized in that; the piston consists of a small-diameter
main piston and a large-diameter sub-piston; and a main communicating port which
supplies the hydraulic fluid to the top surface of the main piston, a sub-communicating
port which supplies the hydraulic fluid to the top surface of the sub-piston, and
a drainage port which drains the hydraulic fluid that acted on the sub-piston are
drilled in a cylinder which guides the piston in an oiltight manner, the main- and
sub-communicating ports and the drainage port are arranged so as to be opened and
closed by the sliding motion of the main piston and the sub-piston, respectively,
and the intake and exhaust valve is equipped with a valve spring which enerizes the
valve in the direction of the valve closing.
[0006] Namely, in the present invention the actuator is given a structure in which the high
pressure hydraulic fluid is used exclusively for the opening of the intake and exhaust
valve, and the energy of the valve spring stored accompanying the lift of the intake
and exhaust valve is used for opening the valve.
[0007] With the above arrangement, the structures of the control valve, the high pressure
pipelines and the actuator are simplified and made small-sized, and it becomes possible
to reduce the consumption of the high pressure hydraulic fluid at the time of valve
opening and to save the driving energy.
4. BRIEF DESCRIPTION OF THE DRAWINGS:
[0008] Figures 1 through 7 show an embodiment of the valve system in accordance with the
present invention, wherein FIG. 1 is an overall sectional diagram of the system and
FIGS. 2 to 7 are sectional diagrams of the important parts for explaining the action
of the actuator of the system and FIG. 8 is a sectional diagram showing the parts
corresponding to FIG. 1 of the prior art system.
5. DETAILED DESCRIPTION OF A PREFERRED EMBODIMENT:
[0009] Referring to FIGS. 1 through 7, an embodiment of the present invention will now be
described next.
[0010] In FIG. 1 which shows the principal structure of the valve system of the exhaust
valve in the Diesel engine, 1 is a three-port directional control valve, and a piston
1a is driven in the vertical direction via a tappet roller 3 by a cam 2 fixed to a
cam shaft 4. The piston 1a slides freely by keeping oil tightness within a cylinder
1b equipped with a hydraulic fluid inlet port 1c, an outlet and inlet port 1d and
a drainage port 1e. Reference numeral 8 is a high pressure pipe which connects the
control valve 1 to the actuator 7. The actuator 7 consists of a main body 20, a cylinder
21, a piston 22, a cover 23 and a check valve unit 30. In the cylinder 21 there are
formed an upper chamber 7b, an intermediate chamber 7c and a lower chamber 7a, and
on the circumferential wall of the cylinder there are formed a main communicating
port 21a, a sub-communicating port 21b, an intermediate port 21c, a lower port 21d
and an oil passage 21e. The main communicating port 21a connects an inlet and outlet
port 20a of the main body 20 and the upper chamber 7b of the cylinder 21. The sub-communicating
port 21b connects the inlet and outlet port 20a of the main body 20 and the intermediate
chamber 7c. The intermediate port 21c connects a return port 20b of the main body
20 and the intermediate chamber 7c. The lower port 21d connects the return port 20b
of the main body 20 and the lower chamber 7a. The oil passage 21e connects the main
communicating port 21a and the check valve unit 30. The piston 22 has a small-diameter
main piston 22a which partitions the upper chamber 7b and the intermediate chamber
7c, and a large-diameter sub-piston 22b which partitions the intermediate chamber
7c and the lower chamber 7a. The main piston 22a and the sub-piston 22b respectively
slide on the inside of the cylinder 21 in oil-tight manner. The check valve unit 30
has a valve plug 30b which is oil-tightly guided on the inside of a guide 30f and
is energized upward by a stopper 30C and a spring 30a. The valve plug 30b disconnects
the oil passage 21e and the upper chamber 7b with its tip abutted on a seat 30d. The
valve plug 30b blocks the flow of the hydraulic fluid from the upper chamber 7b to
the oil passage 21e, but permits the flow of the hydraulic fluid from the oil passage
21e to the upper chamber 7b. Further, the check valve 30 has an orifice 30e at an
end of the oil passage 21e. The orifice 30e is for discharging air that is mixed in
the hydraulic fluid to the outside of the actuator 7 along with the fluid.
[0011] The exhaust valve 6 slides within a guide 15 fixed to a cylinder head 5 and is energized
in the upward direction by a pneumatic valve spring 40. The upper end of a valve stem
6a of the exhaust valve 6 is abutted on the piston 22. A piston 41 is arranged via
a cone sleeve 43 around the valve stem 6a of the exhaust valve 6, and the piston 41
is pressure-fitted to the cone sleeve 43 by means of a presser 44. The piston 41 is
air-tightly guided in the piston 41. Pressurized air is supplied to the interior of
the cylinder 42 from an air tank 45 through an air hole 42a, whereby forming a pneumatic
valve spring 40 within the cylinder 42. The actuator 7 is fixed to the upper part
of the cylinder 42.
[0012] A hydraulic power source 10 consists of a tank 10a, a filter 10b, a high-pressure
pump 10C, an accumulator 10d, and the like, wherein a high-pressure hydraulic fluid
is sent by the pump 10C to the accumulator 10d and the accumulator stores the high-pressure
hydraulic fluid.
[0013] Next, the valve opening operation of the exhaust valve 6 will be described.
[0014] Referring to FIG. 2, the tappet roller 3 is lifted with the rotation of the cam shaft
4, the piston 1a is raised to disconnect the drainage port 1e from the outlet and
inlet port 1d, and the outlet and inlet port 1d is communicated with the inlet port
1c. As the piston 1a is further raised later, the hydraulic fluid in the accumulator
10d is supplied to the inlet and outlet port 20a of the actuator 7 via the control
valve 1, whereby a portion of the fluid is supplied from the inlet and outlet port
20a to the upper chamber 7b via the main communicating port 21a, the oil passage 21e
and the check valve unit 30, while the remaining portion is supplied to the intermediate
chamber 7c from the inlet and outlet port 20a via the sub-communicating port 21b.
Under the action of these portions of the high-pressure hydraulic fluid the piston
22 is moved downward in the direction of the arrow → to be brought to the condition
as illustrated in FIG. 3. Figure 3 depicts the state of the system at the time the
main communicating port 21a is just about to be communicated with the upper chamber
7b. The valve plug 30b of the check valve unit 30 which has been separated from the
seat 30d up to this point is thereafter brought into contact with the seat 30d under
the direct inflow of the hydraulic fluid through the main communicating port 21a into
the upper chamber 7b, and the communication of the oil passage 21e with the upper
chamber 7b is interrupted. The piston 22 is further moved downward and the state shown
in FIG. 4 is reached. The state in FIG. 4 depicts the situation which the communication
of the sub-communicating port 21b with the intermediate chamber 7c is about to be
interrupted. With a further downward motion, from this condition, of the piston in
the direction of the arrow →, the hydraulic fluid acts only on the top surface of
the main piston 22a without acting on the top surface of the sub-piston 22b. Then,
the piston 22 is lowered and the state shown in FIG. 5 is reached. The state in FIG.
5 depicts the condition at the time when the intermediate chamber 7c is just about
to be communicated with the intermediate port 21c. It should be noted that the system
is designed such that the fluid in the intermediate chamber 7c is allowed to expand
during the displacement δ₁ (see FIG. 4) of the piston 22 from the position in FIG.
4 to that in FIG. 5, but following the condition in FIG. 5 the fluid in the intermediate
chamber 7c is discharged to the intermediate port 21c by the main piston 22a under
the condition of positive pressure. The drain oil in the intermediate port 21c is
drained to the tank 10a of the hydraulic power source 10 via the return port 20b.
The hydraulic fluid that enters the upper chamber 7b from the inlet and outlet port
20a via the main communicating port 21a acts on the top surface of the main piston
22a, so that the piston 22 is moved downward in the direction of the arrow → and reaches
the condition shown in FIG. 6. The condition in FIG. 6 illustrates the timing at which
the lower chamber 7a and the lower port 21d are just about to be disconnected. With
a further downward motion of the piston 22, the pressure of the fluid within the lower
chamber 7a is raised because of the sealing, and the fall of the piston 22 is brought
to a gentle stop by the pressure of the hydraulic fluid. Namely, the lower chamber
7a forms a cushion chamber, and the valve-opening motion is completed by the achievement
of a maximum lift position (ℓ
max in FIG. 7) by the exhaust valve 6. It is to be noted that although the hydraulic
fluid is further kept acting on the top surface of the main piston 22a via the inlet
and outlet port 20a, the main- and sub-communicating ports 21a and 21b and the upper
chamber 7b, the exhaust valve 6 is not moved. However, the pressure within the cylinder
42 of the pneumatic valve spring has been increased during the above period.
[0015] Next, the valve-closing operation of the exhaust valve 6 will be described.
[0016] During the operation the piston 22 moves in the direction of the arrow --→ in FIG.
2 to FIG. 6. When the cam 2 starts rotating from the state in FIG. 6 and the tappet
roller 3 reaches the base circle and the piston 1a of the control valve 1 achieve
the state in FIG. 1, the high-pressure hydraulic fluid in the upper chamber 7b is
drained into the tank 10a of the hydraulic power source 10 via the main- and sub-communicating
ports 21a and 21b, the inlet and outlet port 20a and the high-pressure tube 8 through
the control valve 1. The exhaust valve 6 is energized upward via the cone sleeve 43
and the presser 44 by the pneumatic pressure of the pneumatic valve spring 40 that
acts on a piston 41. The valve closing motion is achieved by the pneumatic valve spring
40. When the state in FIG. 6 goes to that of FIG. 5, the intermediate chamber 7c is
closed by the sub-piston 22b. Thereafter, the hydraulic pressure the intermediate
chamber 7c goes up accompanying the rise of the piston 22, performs a primary cushioning
action by the fluid pressure during the displacement δ₁ shown in FIG. 4, and gently
decelerates the speed of upward motion of the piston 22. When the piston 22 reaches
the condition in FIG. 3 by a further motion of the piston 22 in the direction of the
arrow --→, the communication of the upper chamber 7b with the main communicating
port 21a is interrupted. Since, however, the seat 30d of the check valve unit 30 is
closed, the pressure of the fluid in the upper chamber 7b is raised thereafter and
the upper chamber 7b forms a fluid pressure cushioning chamber. The speed of the upward
motion of the piston 22 is further decreased accompanying the secondary cushioning
action of the upper chamber 7b shown in FIG. 2. During the displacement δ₂ until the
seating of the exhaust valve 6, the speed of the exhaust valve 6 is so controlled
as to be optimum for the valve seating to terminate the upward movement, completing
the valve-closing operation. The motion described in the above is represented by diagrams
in which FIG. 7(a) shows the relationship between the lift ℓ of the piston 22 and
the crank angle ϑ
k, and FIG. 7(b) shows the relationship between the exhaust valve resistance F
v, forces F₁ and F₂ acting on the piston 22 and the crank angle ϑ
k. In FIG. 7(a), ℓ₁ indicates the distance during which the hydraulic fluid acts on
the main- and sub-pistons as shown in FIG. 2. The pressure P of the hydraulic fluid,
diameter d
s of the sub-piston and the lift ℓ₁ are related to the force F₁ pressing the exhaust
valve downward and the consumption of the hydraulic fluid by the following relations:
F₁ =
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0001)
d
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0002)
P ,
Q₁ =
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0003)
d
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0004)
ℓ₁,
where the force F₁ and the exhaust valve resistance F
v satisfies the inequality F₁>F
v.
[0017] The interval of the lift from ℓ₁ to ℓ
max corresponds to the period during which the hydraulic fluid acts only on the main
piston 22a. The diameter d
m of the main piston is related to the force F₂ pressing the exhaust valve downward
and the consumption Q₂ of the hydraulic fluid by the following relations:
F₂ =
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0005)
d
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0006)
P ,
Q₂ =
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0007)
d
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0008)
(ℓ
max - ℓ₁) ,
where the force F₂ and the exhaust valve resistance F
v satisfies the inequality F₂>F
v.
[0018] Since the consumption Q of the hydraulic fluid in the case of the prior art system
which is not two-stage type is given by
Q =
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0009)
d
![](https://data.epo.org/publication-server/image?imagePath=1990/41/DOC/EPNWA1/EP90250079NWA1/imgb0010)
ℓ
max ,
it follows that
Q₁ + Q₂ < Q
and it can be seen that it is possible to reduce the consumption of the hydraulic
fluid to a large extent.
[0019] Moreover, during the valve-closing movement over a distance ℓ₂, a primary cushioning
is operative for the period corresponding to the displacement δ₁ which decelerates
the upward motion of the piston 22, and a secondary cushioning is operative for the
period corresponding to the displacement δ₂, so that it is possible to optimally
control the seating speed of the exhaust valve, whereby prolonging the service life
of the exhaust valve.
[0020] It should be noted that in the above description the control valve 1 is operated
by mechanically driving the piston 1a with the cam 2. However, it is also possible
to drive the piston 1a electrically or to use an electromagnetic valve as the control
valve.
[0021] With the aforementioned construction of the present invention, the structure of the
control valve is simplified to involve only three ports, the number of the high-pressure
pipes is reduced to one, and the construction of the actuator is correspondingly simplified.
As a result, there are created cushioning actions due to the hydraulic pressure which
eliminate the mechanically colliding parts, and it becomes possible to optimally control
the seating of the exhaust valve by causing the valve to terminate its valve-opening
cycles in a gentle manner.
[0022] Moreover, the hydraulic fluid is arranged to act on the piston in two stages in response
to the resistance of the exhaust valve, the consumption of the hydraulic fluid is
suppressed to a necessary minimum, and the driving energy of the hydraulic fluid is
reduced, whereby contributing markedly to the reduction of the power loss.
[0023] Furthermore, although the aforementioned embodiments are described in conjunction
with the valve system of an exhaust valve, the valve system of the present invention
can naturally be applied also to an intake valve.