JP6038055B2 - Exhaust valve driving device and internal combustion engine provided with the same - Google Patents

Description

  The present invention relates to a mechanical exhaust valve driving device driven by a cam and an internal combustion engine including the same.

For example, a marine diesel engine (internal combustion engine), which is a low-speed two-stroke cycle diesel engine, drives an exhaust valve using a hydraulic mechanism. In an electronically controlled engine using an electromagnetic valve for hydraulic control of this hydraulic mechanism, the opening / closing timing of the exhaust valve is controlled to be optimal according to the operating load. On the other hand, since the mechanical engine is a cam hydraulic drive system that operates the exhaust valve actuator in accordance with the change in hydraulic pressure generated by the cam driven plunger, the opening / closing timing of the exhaust valve depends on the cam profile. Therefore, it is difficult to change while driving.
In order to solve this problem, in Patent Document 1, the amount of hydraulic oil guided to the exhaust valve actuator is extracted by extracting the hydraulic oil to the buffer tank from the hydraulic pipe that supplies the hydraulic oil to the exhaust valve actuator that drives the exhaust valve. The structure which reduces is adopted. Thereby, the opening timing of the exhaust valve determined by the cam profile is delayed and the closing timing is advanced.
Patent Document 2 adopts a configuration in which high-pressure hydraulic oil is supplied from a separately provided pressurized hydraulic oil source to a hydraulic oil pipe that supplies hydraulic oil to an exhaust valve actuator that drives an exhaust valve. Specifically, by switching the electronically controlled hydraulic valve, the hydraulic oil from the pressurized hydraulic oil source is additionally supplied to the hydraulic oil pipe, and the exhaust valve is opened earlier than the timing determined by the cam profile. Further, by supplying additional hydraulic oil during the cam operation period, the exhaust valve is closed later than the timing determined by the cam profile.

JP-A-6-288210 JP 2010-106843 A

  As in the above-mentioned Patent Documents 1 and 2, various studies have been made to directly control the hydraulic oil in the hydraulic pipe by extracting or additionally supplying the hydraulic oil to the hydraulic pipe that supplies the hydraulic oil to the exhaust valve actuator. ing. However, the present inventors have focused on examining an alternative method that can change the opening / closing timing of the exhaust valve without directly controlling the hydraulic oil in the hydraulic pipe.

  The present invention has been made in view of such circumstances, and an exhaust valve drive device capable of adjusting the opening / closing timing of an exhaust valve without directly controlling hydraulic oil in a hydraulic pipe and an internal combustion engine provided with the exhaust valve drive device The purpose is to provide an institution.

In order to solve the above-described problems, the exhaust valve driving device of the present invention and the internal combustion engine equipped with the same employ the following means.
That is, an exhaust valve driving device according to a reference example of the present invention includes an actuator that operates an exhaust valve of an internal combustion engine, a hydraulic path that supplies hydraulic oil to the actuator, a plunger that is connected to the hydraulic path, and the plunger , A cam for reciprocating the plunger, and pressing means for pressing the exhaust valve in the closing direction by the supplied fluid, and the actuator is operated by the hydraulic oil pressurized by the plunger The exhaust valve driving device that opens the exhaust valve further includes pressure changing means for changing a supply pressure of the compressive fluid supplied to the pressing means.

It is a mechanical exhaust valve drive device that operates a plunger by a cam. That is, the exhaust valve is opened and closed according to the reciprocation of the plunger driven by the cam operation.
In the present invention, the opening / closing timing of the exhaust valve can be changed by adjusting the pressing force by the pressure changing means for changing the supply pressure of the fluid.
For example, when the pressing force is increased by increasing the fluid supply pressure, the pressing force acts as a reaction force in the stroke in which the plunger pressurizes the hydraulic oil and opens the exhaust valve, and the opening timing of the exhaust valve can be delayed. On the other hand, in the stroke in which the plunger depressurizes the pressurized hydraulic oil and closes the exhaust valve, it is urged by the pressing force, so that the exhaust valve closing timing can be advanced.
In addition, as a fluid, it is preferable to use compressive fluids, such as air and nitrogen, and to use the compression reaction force of a compressive fluid.
The pressing means typically employs a configuration of pressing the valve shaft of the exhaust valve, but may be configured to press an actuator connected to the valve shaft of the exhaust valve.

Further, according to the exhaust valve driving apparatus of the reference example of the present invention, the pressure changing means increases the supply pressure of the fluid as the load of the internal combustion engine decreases.

When the supply pressure of the fluid is controlled to increase, the pressing force increases, and the timing for closing the exhaust valve is advanced. The earlier the timing at which the exhaust valve is closed, the more air is sealed in the combustion chamber when the exhaust valve is closed, so that more fresh air is compressed and the compression pressure and combustion pressure of the internal combustion engine increase. . Therefore, by controlling the supply pressure to increase as the load on the internal combustion engine decreases, the combustion of the internal combustion engine is improved even at a low load, and the fuel consumption rate is improved.
In addition, if the control is performed to increase the supply pressure of the fluid and the timing at which the exhaust valve is opened is delayed, the time for performing the gas exchange between the combustion gas and the fresh air in the cylinder may be shortened. In the partial load state where the engine pressure has decreased, the rotational speed of the internal combustion engine is low, so that sufficient time for gas exchange can be taken. In addition, by delaying the opening timing of the exhaust valve, it is possible to maintain the cylinder pressure after combustion for a period of time delayed, so that the cylinder maintained at the cylinder pressure after combustion is maintained. More shaft rotational force can be extracted from the internal gas, and the fuel consumption rate is further improved.

  The exhaust valve driving device of the present invention includes an actuator that operates an exhaust valve of an internal combustion engine, a hydraulic path that supplies hydraulic oil to the actuator, a plunger that is connected to the hydraulic path, and a cylinder that houses the plunger. And a cam for reciprocating the plunger, and pressing means for pressing the exhaust valve in the closing direction by the supplied fluid, and the actuator is operated by the hydraulic oil pressurized by the plunger to exhaust the exhaust In the exhaust valve driving device that opens the valve, the pressing means receives a pressure from the fluid and transmits a pressing force to the exhaust valve, and a pressure receiving area change that enables the pressure receiving area of the pressure receiving member to be changed. And means.

It is a mechanical exhaust valve drive device that operates a plunger by a cam. That is, the exhaust valve is opened and closed according to the reciprocation of the plunger driven by the cam operation.
In the present invention, the opening / closing timing of the exhaust valve can be changed by adjusting the pressing force so that the pressure receiving area of the pressure receiving member of the pressing means can be changed.
For example, if the pressure receiving area is increased to increase the compression reaction force, the pressing force by the fluid acts as a reaction force in the stroke in which the plunger pressurizes the hydraulic oil and opens the exhaust valve, and the opening timing of the exhaust valve may be delayed. it can. On the other hand, in the stroke in which the plunger depressurizes the pressurized hydraulic oil and closes the exhaust valve, it is energized by the pressing force of the fluid, so that the closing timing of the exhaust valve can be advanced.
In addition, as a fluid, it is preferable to use compressive fluids, such as air and nitrogen, and to use the compression reaction force of a compressive fluid.

  Further, according to the exhaust valve driving device of the present invention, the pressure receiving area changing means increases the pressure receiving area as the load of the internal combustion engine decreases.

When the pressure receiving area is controlled to be large, the pressing force by the fluid increases, so that the timing at which the exhaust valve is closed is advanced. The earlier the timing at which the exhaust valve is closed, the more air is sealed in the combustion chamber when the exhaust valve is closed, so that more fresh air is compressed and the compression pressure and combustion pressure of the internal combustion engine increase. . Therefore, by controlling the supply pressure to increase as the load on the internal combustion engine decreases, the combustion of the internal combustion engine is improved even at a low load, and the fuel consumption rate is improved.
Also, if the pressure receiving area is controlled to be large and the timing at which the exhaust valve is opened is delayed, there is a risk that the time for performing the gas exchange between the combustion gas and fresh air in the cylinder may be shortened, but the load will be reduced. In the partial load state, since the rotational speed of the internal combustion engine is low, a sufficient time for gas exchange can be taken. In addition, by delaying the opening timing of the exhaust valve, it is possible to maintain the cylinder pressure after combustion for a period of time delayed, so that the cylinder maintained at the cylinder pressure after combustion is maintained. More shaft rotational force can be extracted from the internal gas, and the fuel consumption rate is further improved.

Furthermore, according to the exhaust valve drive device of the present invention, said pressing means has a plurality of said pressure receiving member, the pressure receiving area changing means, changing the number of the pressure-receiving member for transmitting the pressing force to the exhaust valve It is characterized by.

  The pressure receiving area can be changed by changing the number of pressure receiving members that apply the compression reaction force to the exhaust valve. Thereby, the opening / closing timing of the exhaust valve can be arbitrarily changed.

  An internal combustion engine of the present invention includes the exhaust valve driving device according to any one of the above, the exhaust valve driven by the exhaust valve driving device, and a combustion chamber that houses the exhaust valve. It is characterized by.

  Since any one of the exhaust valve driving devices described above is provided, an internal combustion engine capable of adjusting the exhaust valve operation with a simple configuration can be provided.

  By changing the pressing force of the pressing means that presses the exhaust valve in the closing direction by the compression reaction force of the compressive fluid, the opening / closing timing of the exhaust valve can be changed without directly controlling the hydraulic oil in the hydraulic pipe.

It is the schematic block diagram which showed the exhaust valve drive device concerning the reference embodiment of this invention. 3 is a graph showing changes in hydraulic oil pressure and exhaust valve lift when the exhaust valve driving device of FIG. 1 is used. 1 is a schematic configuration diagram illustrating an exhaust valve driving device according to a first embodiment of the present invention. It is the schematic block diagram which showed the state which switched the electronically controlled hydraulic valve of the exhaust valve drive device of FIG. It is the graph which showed the pressure change of the hydraulic fluid at the time of using the exhaust valve drive device of FIG. 3, and the change of an exhaust valve lift.

Embodiments according to the present invention will be described below with reference to the drawings.
[ Reference embodiment]
FIG. 1 shows an exhaust valve driving device 1 according to this embodiment. The exhaust valve drive device 1 is provided in a marine main engine diesel engine (internal combustion engine). A diesel engine for a ship main engine (hereinafter referred to as “diesel engine”) is, for example, a low-speed two-stroke cycle engine, and adopts a uniflow type that is scavenged in one direction so as to supply air from below and exhaust upward. ing. The output from the diesel engine is directly or indirectly connected to the screw propeller via a propeller shaft (not shown).

  As shown in FIG. 1, the exhaust valve drive device 1 includes an exhaust valve 5 that opens and closes an exhaust passage formed in the cylinder cover 3, a piston (actuator) 7 that drives the exhaust valve 5, an air spring. A hydraulic passage 9 for supplying hydraulic oil to the piston 7, a plunger 11 connected to the hydraulic passage 9, and a cam 13 for reciprocating the plunger 11.

  The piston 7 is connected to a shaft portion 5a of the exhaust valve 5 extending in the vertical direction, and reciprocates in the first cylinder 15 in the vertical direction. One end 9 a of the hydraulic path 9 is connected to the hydraulic chamber 17 formed by the first cylinder 15 and the piston 7.

  The air spring portion 6 includes an air cylinder 8 that stores air (compressible fluid) and an air piston 10. An air supply path 12 is connected to the air cylinder 8. A check valve 14 is provided in the air supply path 12, and a buffer tank 16 and an air compressor (pressure changing means) 18 are provided upstream thereof. The air pressurized by the air compressor 18 is accumulated in the buffer tank 16, and the air in the buffer tank 16 is supplied into the air cylinder 8 through the check valve 14. The air pressure in the air cylinder 8 is determined by the pressure in the buffer tank 16, and the pressure in the buffer tank 16 is determined by the air compressor 18 controlled by a control unit (not shown). The air stored in the air cylinder 8 is prevented from flowing back to the buffer tank 16 side by the check valve 14. The check valve 14 forms a space in which the air cylinder 8 is closed, and an air spring (air spring) using air compressibility is formed.

  The air piston 10 is directly or indirectly fixed to the shaft portion 5 a of the exhaust valve 5, and air pressure applied to the air piston 10 acts on the exhaust valve 5. As a result, the exhaust valve 5 is pressed upward in FIG. 1, that is, toward the first cylinder 15.

An orifice path 19 branched from the first branch point 9b is connected to the hydraulic path 9. The orifice path 19 is provided with an orifice 21 which is a fixed throttle.
When the pressure in the hydraulic path 9 exceeds a predetermined value, a predetermined amount of hydraulic oil is discharged from the orifice 21 to the outside of the hydraulic path 9. As a result, a predetermined amount of hydraulic oil is discharged out of the hydraulic path 9 when pressurized by the plunger 11, and the amount of oil remaining in the hydraulic path 9 is reduced when the plunger 11 is depressurized. Is held higher (exhaust valve closing direction) than when pressurized. When the hydraulic oil is sucked down by pushing down the plunger 11, the same amount of oil as in the pressurization is sucked in, so that the piston 7 is surely sucked upward before the decompression by the plunger 11 is completed. The exhaust valve 5 is stably closed.

  A low-pressure hydraulic oil supply path 23 branched from the second branch point 9c is connected to the hydraulic path 9. The low pressure hydraulic oil supply path 23 is supplied with a hydraulic pressure as a base used when opening and closing the exhaust valve 5 from a low pressure hydraulic oil source (not shown). The low pressure hydraulic oil supply path 23 is provided with a check valve 25, and when the hydraulic pressure in the hydraulic pressure path 9 becomes a predetermined value or less, insufficient hydraulic oil is supplied from the low pressure hydraulic oil supply path 23. It has become so. As a result, the base hydraulic pressure, specifically, the base pressure that is the lowest operating hydraulic pressure shown in FIG. 2B is maintained. On the other hand, the check valve 25 is kept closed when the pressure in the hydraulic path 9 is equal to or higher than a predetermined value. That is, the check valve 25 is closed during the pressurization stroke by the plunger 11.

  The plunger 11 reciprocates in the second cylinder 27 in the vertical direction. The other end 9 d of the hydraulic path 9 is connected to a pressurizing chamber (pressurizing space) 29 formed by the second cylinder 27 and the plunger 11.

A connection shaft 35 is attached to the lower portion of the plunger 11, and a cam roller 37 is provided at the lower end of the connection shaft 35. The cam roller 37 rolls on the outer peripheral surface of the lower cam 13, that is, on the profile.
The cam 13 is fixed to the cam shaft 39 and rotates together with the cam shaft 39. The camshaft 39 rotates in synchronization with the crankshaft of the diesel engine.

Next, the operation of the exhaust valve driving device 1 having the above-described configuration will be described with reference to FIG.
First, a case where the air pressure stored in the air cylinder 8 is relatively low will be described, and then a case where the air pressure is relatively high will be described.

<Air pressure; Low>
The case where the air pressure in the air cylinder 8 is relatively low is mainly used when the load of the diesel engine is high. The pressure in the air cylinder 8 is determined by an air compressor 18 controlled by a control unit (not shown).

  2, (a) shows the lift amount of the cam 13, (b) shows the hydraulic pressure in the hydraulic path 9, (c) shows the air spring pressure, which is the pressure in the air cylinder 8, and (d) shows the exhaust valve 5. The lift amount is shown. In the figure, when the air pressure is relatively low, it is indicated by a solid line.

When the cam lift amount increases in accordance with the profile of the cam 13 at the time t0 and the plunger 11 starts to be pushed up, the hydraulic pressure in the pressurizing chamber 29, that is, the hydraulic path 9, starts to rise. At time t1, the cam lift amount reaches the maximum value, the plunger 11 is pushed up to the top dead center, and when the operating oil pressure reaches the maximum value, at time t2, the hydraulic pressure in the hydraulic chamber 17 on the piston 7 side acts, The piston 7 is pushed down by overcoming the pressing force and in-cylinder pressure of the air spring portion 6. Thereby, the exhaust valve lift amount increases and the exhaust valve 5 is opened. At this time, as the piston 7 is pushed down, the hydraulic oil is taken into the hydraulic chamber 17, so that the hydraulic pressure decreases rapidly. Further, when the exhaust valve lift amount increases, the air spring pressure rises as the air piston 10 moves downward to reduce the volume in the air cylinder 8. The exhaust valve lift amount maintains the lift amount as it is for a predetermined period after reaching the maximum value at time t3.
Then, during the period up to time t5 when the plunger 11 is maintained at the top dead center according to the profile of the cam 13, the exhaust valve lift amount is also maintained at the maximum, and the exhaust valve 5 remains open.

When the cam lift amount decreases according to the profile of the cam 13 and the plunger 11 starts to descend at time t5, the hydraulic pressure starts to decrease. When the operating oil pressure falls below a predetermined value, the pushing force and the in-cylinder pressure of the air spring portion 6 overcome and the piston 7 is pushed upward from time t6, whereby the exhaust valve lift amount starts to decrease. When the exhaust valve lift amount starts to decrease, the air spring pressure decreases as the air piston 10 moves upward to increase the volume in the air cylinder 8.
When the cam lift amount reaches the minimum value and the plunger 11 is lowered to the bottom dead center, the exhaust valve 5 is fully closed at time t7.

<Air pressure; high>
Next, when the load of the diesel engine decreases and becomes a low load side, the discharge pressure of the air compressor 18 is increased in accordance with an instruction from a control unit (not shown), and the air pressure in the air cylinder 8 is increased. . As shown by a broken line in FIG. 2C, it can be seen that the air spring pressure is increased.

  When the cam lift amount increases according to the profile of the cam 13 at the time t0 and the plunger 11 starts to be pushed up, the hydraulic pressure in the pressurizing chamber 29, that is, the hydraulic path 9, starts to rise. At time t1, the cam lift amount reaches the maximum value, the plunger 11 is pushed up to the top dead center, and the hydraulic pressure reaches the maximum value. At time t2 'that is later than time t2, the hydraulic pressure in the hydraulic chamber 17 on the piston 7 side acts to overcome the pressing force and in-cylinder pressure of the air spring portion 6 and push down the piston 7. Thereby, the exhaust valve lift amount increases and the exhaust valve 5 is opened. Thus, when the air pressure in the air cylinder 8 is relatively high, the time when the exhaust valve 5 is opened is delayed as compared with the case where the air pressure is relatively low. This is because the compression reaction force by the air spring is increased by increasing the air pressure in the air cylinder 8, and in the stroke in which the plunger 11 pressurizes the hydraulic oil and opens the exhaust valve 5, the pressing force due to the increased compression reaction force is increased. This is because it acts as a larger reaction force. This can also be understood from the fact that the air spring pressure is larger on the broken line (high air pressure) than on the solid line (low air pressure) as shown in FIG. The exhaust valve lift amount maintains the lift amount as it is for a predetermined period after reaching the maximum value at time t3 '.

  Then, when the cam lift amount decreases according to the profile of the cam 13 and the plunger 11 starts to descend at time t5, the hydraulic pressure starts to decrease. At this time, since the air pressure in the air cylinder 8 is relatively high (see the broken line in FIG. 2C), the pressing force due to the compression reaction force of the increased air spring pressure closes the exhaust valve 5. Is energized. As a result, the exhaust valve lift amount starts to decrease earlier than time t6 at time t6 ', and as a result, the exhaust valve 5 is fully closed earlier than time t7 at time t7'.

  In this way, as can be seen by referring to the change in the exhaust valve lift in FIG. 2D, by increasing the air pressure in the air cylinder 8, the opening timing of the exhaust valve 5 is delayed, and the exhaust valve The closing timing of 5 can be advanced. Further, the opening / closing timing of the exhaust valve 5 can be adjusted by appropriately adjusting the amount of increase in air pressure according to a command from a control unit (not shown).

According to the exhaust valve driving device 1 of the present embodiment, the following operational effects can be achieved.
The opening / closing timing of the exhaust valve 5 can be changed by changing the compression reaction force of the air spring and adjusting the pressing force by changing the pressure of the air supplied into the air cylinder 8. Specifically, the compression reaction force is increased by increasing the air spring pressure, and in the stroke in which the plunger 11 pressurizes the hydraulic oil and opens the exhaust valve, the pressing force due to the increased compression reaction force acts as a reaction force, The opening timing of the exhaust valve 5 can be delayed. On the other hand, in the stroke in which the plunger 11 depressurizes the pressurized hydraulic oil and closes the exhaust valve 5, the closing timing of the exhaust valve 5 can be advanced by energizing with the pressing force due to the increased compression reaction force.

When the air spring pressure is controlled to be high, the pressing force due to the compression reaction force is increased, so that the timing at which the exhaust valve 5 is closed is advanced. The earlier the timing at which the exhaust valve 5 is closed, the more air is sealed in the combustion chamber when the exhaust valve is closed, so the amount of fresh air that is compressed increases and the compression pressure and combustion pressure of the diesel engine increase. Become. Therefore, by controlling so that the air spring pressure increases as the load on the diesel engine decreases, the combustion of the diesel engine can be improved and the fuel consumption rate can be improved even at a low load.
Further, if the air spring pressure is controlled so as to increase and the timing at which the exhaust valve 5 is opened is delayed, there is a possibility that the time for performing the gas exchange between the combustion gas and the fresh air in the cylinder may be shortened. In the partial load state in which the engine pressure has decreased, the rotational speed of the diesel engine is low, so that sufficient time can be taken for gas exchange. Further, by delaying the opening timing of the exhaust valve 5, the in-cylinder pressure after combustion can be maintained without being reduced by the amount of time that the opening timing is delayed, so that the in-cylinder pressure after combustion is maintained. More shaft rotational force can be extracted from the in-cylinder gas, and the fuel consumption rate can be further improved.

First Embodiment
Next, a first embodiment of the present invention will be described with reference to FIGS.
This embodiment differs from the reference embodiment in that the air spring pressure is changed, whereas the pressure receiving area of the air piston on which the air spring pressure acts is changed. Other configurations that are common to the other components are assigned the same reference numerals, and the description thereof is omitted.

  As shown in FIG. 3, the air spring portion 6 'includes a main air cylinder 8a and a sub air cylinder 8b. A main air piston (pressure receiving member) 10a is fitted into the main air cylinder 8a, and a sub air piston (pressure receiving member) 10b is fitted into the sub air cylinder 8b. The main air piston 10a and the sub air piston 10b are fixed to each other by the connecting member 40, and not only the force applied to the main air piston 10a but also the force applied to the sub air piston 10b is transmitted to the shaft portion 5a of the exhaust valve 5. It has become so. That is, there are two air springs, a main air spring composed of the main air cylinder 8a and the main air piston 10a, and a sub air spring composed of the sub air cylinder 8b and the sub air piston 10b. The air spring also acts on the exhaust valve 5.

An air supply path 12 is connected to the main air cylinder 8a via a check valve 14 as in the reference embodiment. The main air cylinder 8a is connected to a first communication path 44 connected to an electromagnetic valve (pressure receiving area changing means) 42. A second communication path 46 is connected between the electromagnetic valve 42 and the sub air cylinder 8b. Furthermore, a third communication path 48 connected to the air supply path 12 on the upstream side (buffer tank 16 side) of the check valve 14 is connected to the electromagnetic valve 42.
The electromagnetic valve 42 is a switching valve for switching the flow path, and is controlled by a control unit (not shown). Specifically, in the state shown in FIG. 3, the electromagnetic valve 42 connects the main air cylinder 8a and the sub air cylinder 8b by connecting the first communication path 44 and the second communication path 46, and The second communication path 46 and the third communication path 48 are blocked, and the sub air cylinder 8b and the air supply path 12 are disconnected. Thereby, the main air cylinder 8a and the sub air cylinder 8b become a space closed by the check valve 14, and an air spring is formed in which the pressure receiving surfaces of the main piston 10a and the sub piston 10b are the pressure receiving areas, and the large pressure receiving area state is obtained. It is formed.
On the other hand, in the state shown in FIG. 4, the electromagnetic valve 42 disconnects the first communication path 44 and the second communication path 46 to disconnect the main air cylinder 8a and the sub air cylinder 8b, and By connecting the second communication path 46 and the third communication path 48, the sub air cylinder 8 b and the air supply path 12 are connected. As a result, the space closed by the check valve 14 is only the main air cylinder 8a, and an air spring in which only the pressure receiving surface of the main piston 10a has a pressure receiving area is formed, and a small pressure receiving area state is formed. The auxiliary air cylinder 8b communicates with the air supply path 12 on the upstream side of the check valve 14, and does not form a closed space that functions as an air spring, so that it does not act as an air spring.
In this way, by switching the electromagnetic valve 42 by a control unit (not shown), an air spring that can obtain a large compression reaction force as the large pressure receiving area state shown in FIG. 3 is configured, and the small pressure receiving area shown in FIG. An air spring that can obtain a small compression reaction force as a state is configured.

Next, the operation of the exhaust valve driving device 1 ′ having the above configuration will be described with reference to FIG. 5.
First, the case where the pressure receiving area of the air spring portion 6 ′ is relatively small, that is, the case where the electromagnetic valve 42 is in the state shown in FIG. 4, will be described. Next, the case where the pressure receiving area is relatively large, that is, the electromagnetic valve 42 is illustrated. A case where the state 3 is set will be described.

<Pressure receiving area; Small>
When the pressure receiving area of the air spring portion 6 ′ is relatively small, it is mainly used when the load of the diesel engine is high. In this case, the solenoid valve 42 is in the state shown in FIG. 4, the main air cylinder 8a and the sub air cylinder 8b are not connected, and only the pressure receiving surface of the main air piston 10a is the pressure receiving area of the air spring. As a result, the force acting on the exhaust valve 5 is relatively small.

  In FIG. 5, as in FIG. 2, (a) is the lift amount of the cam 13, (b) is the hydraulic pressure in the hydraulic path 9, (c) is the air spring pressure that is the pressure in the air cylinder 8, (d ) Shows the lift amount of the exhaust valve 5. In the figure, when the pressure receiving area is relatively small, it is indicated by a solid line.

When the cam lift amount increases in accordance with the profile of the cam 13 at the time t0 and the plunger 11 starts to be pushed up, the hydraulic pressure in the pressurizing chamber 29, that is, the hydraulic path 9, starts to rise. At time t1, the cam lift amount reaches the maximum value, the plunger 11 is pushed up to the top dead center, and when the operating oil pressure reaches the maximum value, at time t2, the hydraulic pressure in the hydraulic chamber 17 on the piston 7 side acts, The piston 7 is pushed down by overcoming the pressing force and in-cylinder pressure of the air spring portion 6 '. Thereby, the exhaust valve lift amount increases and the exhaust valve 5 is opened. At this time, as the piston 7 is pushed down, the hydraulic oil is taken into the hydraulic chamber 17, so that the hydraulic pressure decreases rapidly. Further, when the exhaust valve lift amount increases, the air spring pressure increases as the main air piston 10a moves downward to decrease the volume in the air cylinder 8a. On the other hand, with respect to the secondary air cylinder 8b, as shown by a one-dot chain line in FIG. 5C, the secondary air cylinder 8b communicates with the air supply path 12, so that the secondary air piston 10b moves downward. Even if the volume in the sub air cylinder 8b is decreased, the air spring pressure does not increase. The exhaust valve lift amount maintains the lift amount as it is for a predetermined period after reaching the maximum value at time t3.
Then, during the period up to time t5 when the plunger 11 is maintained at the top dead center according to the profile of the cam 13, the exhaust valve lift amount is also maintained at the maximum, and the exhaust valve 5 remains open.

When the cam lift amount decreases according to the profile of the cam 13 and the plunger 11 starts to descend at time t5, the hydraulic pressure starts to decrease. When the operating oil pressure falls below a predetermined value, the pushing force and the in-cylinder pressure of the air spring portion 6 'overcome and the piston 7 is pushed upward from time t6, whereby the exhaust valve lift amount starts to decrease. When the exhaust valve lift amount starts to decrease, the air spring pressure decreases as the main air piston 10a moves upward to increase the volume in the main air cylinder 8a. In this case, since the sub air piston 10b is connected to the air supply path 12 even if the sub air piston 10b moves upward and the volume in the sub air cylinder 8b increases, The spring pressure remains constant.
When the cam lift amount reaches the minimum value and the plunger 11 is lowered to the bottom dead center, the exhaust valve 5 is fully closed at time t7.

<Pressure receiving area; Large>
Next, when the load of the diesel engine decreases and becomes a low load side, the position of the electromagnetic valve 42 is changed to the state shown in FIG. 3 in accordance with an instruction from a control unit (not shown), and the air spring portion 6 ′. Increase the pressure receiving area. Specifically, the main air cylinder 8a and the sub air cylinder 8b are connected by the electromagnetic valve 42, and the total pressure receiving surface of the main air piston 10a and the sub air piston 10b is the pressure receiving area of the air spring. As a result, the force acting on the exhaust valve 5 is relatively large. The air spring pressure is determined by the air compressor 18 and is not different from the above-described <pressure receiving area; small>. However, since the pressure receiving area is large, the force acting on the exhaust valve 5 is relatively large. .

  When the cam lift amount increases according to the profile of the cam 13 at the time t0 and the plunger 11 starts to be pushed up, the hydraulic pressure in the pressurizing chamber 29, that is, the hydraulic path 9, starts to rise. At time t1, the cam lift amount reaches the maximum value, the plunger 11 is pushed up to the top dead center, and the hydraulic pressure reaches the maximum value. Then, at time t2 'delayed from time t2, the hydraulic pressure in the hydraulic chamber 17 on the piston 7 side acts to overcome the pressing force and in-cylinder pressure of the air spring 6' and push down the piston 7. Thereby, the exhaust valve lift amount increases and the exhaust valve 5 is opened. As described above, when the pressure receiving area of the air spring portion 6 ′ is relatively large, the time when the exhaust valve 5 is opened is later than when the pressure receiving area is relatively small. This increases the pressure receiving area of the air spring portion 6 ′ to increase the compression reaction force by the air spring. In the stroke in which the plunger 11 pressurizes the hydraulic oil and opens the exhaust valve 5, the pressing by the increased compression reaction force is performed. This is because the force acts as a larger reaction force. The exhaust valve lift amount maintains the lift amount as it is for a predetermined period after reaching the maximum value at time t3 '.

  Then, when the cam lift amount decreases according to the profile of the cam 13 and the plunger 11 starts to descend at time t5, the hydraulic pressure starts to decrease. At this time, since the pressure receiving area of the air spring portion 6 ′ is relatively large, the pressing force due to the increased compression reaction force is urged in the direction of closing the exhaust valve 5. As a result, the exhaust valve lift amount starts to decrease earlier than time t6 at time t6 ', and as a result, the exhaust valve 5 is fully closed earlier than time t7 at time t7'.

  Thus, as can be seen by referring to the change in the exhaust valve lift in FIG. 5 (d), by increasing the pressure receiving area of the air spring portion 6 ′, the opening timing of the exhaust valve 5 is delayed, and the exhaust gas is exhausted. The closing timing of the valve 5 can be advanced.

According to the exhaust valve driving device 1 ′ of the present embodiment, the following operational effects can be achieved.
When the pressure receiving area of the air spring portion 6 ′ is increased to increase the compression reaction force, the pressing force by the compression reaction force acts as a reaction force in the stroke in which the plunger 11 pressurizes the hydraulic oil and opens the exhaust valve 5, and the exhaust gas is exhausted. The opening timing of the valve 5 can be delayed. On the other hand, in the stroke in which the plunger 11 depressurizes the pressurized hydraulic oil and closes the exhaust valve 5, the closing timing of the exhaust valve 5 can be advanced because it is urged by the pressing force due to the compression reaction force.

As described above, when the timing at which the exhaust valve 5 is closed becomes earlier, the period for taking new air into the combustion space of the diesel engine becomes longer, so the amount of fresh air to be compressed increases, and the compression pressure and combustion pressure of the diesel engine increase. . Therefore, by controlling so that the pressure receiving area increases as the load of the diesel engine decreases, combustion improvement of the diesel engine is performed even when the load is low, and the fuel consumption rate is improved.
Also, if the pressure receiving area is controlled to be large and the timing at which the exhaust valve 5 is opened is delayed, there is a possibility that the time for performing the gas exchange between the combustion gas and fresh air in the cylinder may be shortened. Since the rotational speed of the diesel engine is low in the lowered partial load state, a sufficient time for gas exchange can be taken. Further, by delaying the opening timing of the exhaust valve 5, the in-cylinder pressure after combustion can be maintained without being reduced by the amount of time that the opening timing is delayed. More shaft rotational force can be extracted from the in-cylinder gas, and the fuel consumption rate is further improved.

  Further, since the pressure receiving area is changed by changing the number of air cylinders and air pistons that apply the compression reaction force to the exhaust valve 5, the opening / closing timing of the exhaust valve 5 can be changed easily. In the present embodiment, two combinations of the main air cylinder 8a and the main air piston 10a and the sub air cylinder 8b and the sub air piston 10b are used. The opening / closing timing of the valve 5 may be changed.

Note that the exhaust valve driving devices 1 and 1 ′ of the above-described embodiments may be provided for each cylinder of the diesel engine, or the piston 7, the first cylinder 15, the cam 13 and the plunger 11, and the check valve 14. The buffer tank 16 may be shared by a plurality of cylinders.
Further, by combining the reference embodiment and the first embodiment, the air supply pressure may be changed as in the reference embodiment when the pressure receiving area shown in FIG. 3 is large, or the pressure receiving area shown in FIG. However, the air supply pressure may be changed as in the reference embodiment.
Moreover, in each said embodiment, although demonstrated using air (air) as an example of a compressive fluid, other compressive fluids, such as nitrogen, may be sufficient, for example.

1, 1 'exhaust valve drive device 3 cylinder cover 5 exhaust valve 6, 6' air spring (pressing means)
7 Piston 8 Air cylinder 8a Main air cylinder 8b Sub air cylinder 9 Hydraulic path 10 Air piston 10a Main air piston 10b Sub air piston 11 Plunger 12 Air supply path 13 Cam 14 Check valve 15 First cylinder 16 Buffer tank 17 Hydraulic chamber 18 Air compressor (pressure change means)
19 Orifice path 21 Orifice 23 Low pressure hydraulic oil supply path 25 Check valve 27 Second cylinder 29 Pressurizing chamber 35 Connecting shaft 37 Cam roller 40 Connecting member 42 Solenoid valve (pressure receiving area changing means)
44 1st communication path 46 2nd communication path 48 3rd communication path

Claims (4)

An actuator for operating an exhaust valve of an internal combustion engine;
A hydraulic path for supplying hydraulic oil to the actuator;
A plunger connected to the hydraulic path;
A cylinder containing the plunger;
A cam for reciprocating the plunger;
Pressing means for pressing the exhaust valve in the closing direction by the supplied fluid;
With
In the exhaust valve driving device that opens the exhaust valve by operating the actuator by the hydraulic oil pressurized by the plunger,
The pressing means receives a pressure from the fluid and transmits a pressing force to the exhaust valve;
Pressure receiving area changing means for changing the pressure receiving area of the pressure receiving member;
An exhaust valve driving device comprising: The exhaust valve driving device according to claim 1 , wherein the pressure receiving area changing means increases the pressure receiving area as the load of the internal combustion engine decreases. The pressing means has a plurality of the pressure receiving members,
The pressure receiving area changing means, the exhaust valve driving apparatus according to claim 2, wherein changing the number of the pressure-receiving member for transmitting the pressing force to the exhaust valve. An exhaust valve driving device according to any one of claims 1 to 3 ,
The exhaust valve driven by the exhaust valve driving device;
A combustion chamber containing the exhaust valve;
An internal combustion engine comprising:

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