JP2008255815A - Diesel engine - Google Patents

Abstract

The present invention relates to a diesel engine, and an object thereof is to control the strength of an in-cylinder swirl with a high degree of freedom with a relatively simple structure.
A diesel engine 10 includes a first intake valve 14a, a second intake valve 14b, a first intake port that communicates with the first intake valve 14a, a second intake port that communicates with the second intake valve 14b, A first valve operating mechanism 20 for driving one intake valve 14a at a fixed operating angle; a minimum operating angle smaller than the fixed operating angle of the first intake valve 14a; and a maximum operating angle larger than the fixed operating angle of the first intake valve 14b. And a second valve mechanism 22 for driving the second intake valve 14b with a variable working angle. The swirl ratio of the first intake port alone is larger than the swirl ratio of the second intake port alone.
[Selection] Figure 1

Description

  The present invention relates to a diesel engine.

  Japanese Patent Application Laid-Open No. 2004-211699 discloses an internal combustion engine that includes a first intake valve in which an opening period and a lift amount are both fixed, and a second intake valve that makes them variable. More specifically, in this internal combustion engine, the opening period of the second intake valve is switched to either the standard opening period or the shortened opening period. The first intake valve is closed at the same timing. And it is supposed that the in-cylinder flow can be strengthened by setting the second intake valve as a shortened opening period and not opening the second intake valve until the vicinity of the time when the piston speed becomes maximum.

Japanese Patent Laid-Open No. 2004-211699 JP 2006-188959 A JP-A-11-229913

  However, the in-cylinder flow is not only required to be strengthened, but depending on the operating region, the in-cylinder flow may need to be weakened.

  This invention is made in view of said point, and it aims at providing the diesel engine which can control the intensity | strength of an in-cylinder swirl with a high degree of freedom with a comparatively simple structure.

In order to achieve the above object, a first invention is a diesel engine,
A first intake valve;
A second intake valve;
A first intake port leading to the first intake valve;
A second intake port leading to the second intake valve;
A first valve mechanism for driving the first intake valve at a fixed operating angle;
A second valve that drives the second intake valve with a variable operating angle between a minimum operating angle smaller than the fixed operating angle of the first intake valve and a maximum operating angle larger than the fixed operating angle of the first intake valve. Mechanism,
With
The swirl ratio of the first intake port alone is larger than the swirl ratio of the second intake port alone.

The second invention is the first invention, wherein
In a predetermined operating region, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve, and after the first intake valve is closed, the air in the cylinder is increased as the piston rises. The second intake valve is closed after a part is caused to flow back to the second intake port through the second intake valve.

The third invention is the first or second invention, wherein
A phase variable mechanism for changing a phase of a valve opening period of the first intake valve and the second intake valve is further provided.

Moreover, 4th invention is set in 3rd invention,
The first intake valve and the second intake valve are driven by a common intake camshaft,
The phase variable mechanism is characterized in that the phase of the intake camshaft is variable.

The fifth invention is the third or fourth invention, wherein
The second valve mechanism changes the operating angle by changing the closing timing while keeping the opening timing of the second intake valve constant,
The opening timing of the first intake valve and the opening timing of the second intake valve are substantially the same.

The sixth invention is the fifth invention, wherein
In the middle-speed intermediate load region, the second intake valve has an operation angle larger than a fixed operation angle of the first intake valve, and includes a control unit that sets the phase variable mechanism in a reference state.

The seventh invention is the fifth invention, wherein
In the high-speed and high-load region, there is provided control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism in an advanced state.

The eighth invention is the fifth invention, wherein
In a light-medium load region, control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism in an advanced state is provided.

The ninth invention is the third or fourth invention, wherein
The second valve mechanism changes the operating angle by changing the opening timing while keeping the closing timing of the second intake valve constant,
The closing timing of the first intake valve and the closing timing of the second intake valve are substantially the same.

  In a tenth aspect based on the ninth aspect, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve in the low speed and light load region, and the phase variable mechanism is in a reference state. It is characterized by comprising the following control means.

  In an eleventh aspect based on the ninth aspect, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve in the high speed and high load region, and the phase variable mechanism is in a reference state. It is characterized by comprising the following control means.

  In a twelfth aspect based on the ninth aspect, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve and the phase variable mechanism is advanced in the low speed and light load region. It is characterized by comprising control means for setting the state.

  In a thirteenth aspect based on the ninth aspect, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve and the phase variable mechanism is advanced in the low speed and high load region. It is characterized by comprising control means for setting the state.

  In the first invention, as the operating angle of the second intake valve is made smaller than the fixed operating angle of the first intake valve, the amount of air flowing into the cylinder through the second intake port having a small swirl ratio decreases. Thus, the amount of air flowing into the cylinder through the first intake port having a large swirl ratio increases. For this reason, the intensity | strength of the swirl formed in a cylinder can be raised. Conversely, as the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve, the amount of air flowing into the cylinder through the second intake port having a small swirl ratio increases, while the swirl ratio The amount of air flowing into the cylinder through the first intake port having a large value decreases. For this reason, it can suppress that a swirl is formed in a pipe | tube, and can suppress the intensity | strength of a swirl. That is, according to the first invention, by changing the operating angle of the second intake valve, the amount of air flowing into the cylinder through the first intake port having a large swirl ratio and the second intake having a small swirl ratio are obtained. The ratio of the amount of air flowing into the cylinder through the port can be freely changed. For this reason, the intensity | strength of the swirl formed in a cylinder can be changed with a high freedom degree, and the optimal swirl intensity according to the driving | running state is realizable. Therefore, good combustion according to the driving state can be performed, and emission, fuel consumption, torque fluctuation, and the like can be improved. Furthermore, it is only necessary to provide the operating angle variable mechanism only in the second intake valve, and it is not necessary to provide the operating angle variable mechanism in the first intake valve. For this reason, compared with the case where an operating angle variable mechanism is provided in both the 1st intake valve and the 2nd intake valve, cost reduction and weight reduction can be aimed at. In addition, it is possible to reduce the size and to be mounted on a vehicle.

  According to the second invention, in a predetermined operating region, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve, and after the first intake valve is closed, the piston is raised. The second intake valve can be closed after a part of the air in the cylinder flows back to the second intake port through the second intake valve. At this time, the air that is going to flow backward from the cylinder to the second intake port flows into the second intake port while turning in the direction opposite to the swirling direction of the swirl formed in the cylinder in the intake stroke. Acts to counter the swirl inside. For this reason, it is possible to sufficiently suppress the swirl in an operation region where it is necessary to suppress the swirl (for example, a light / medium load region where HC is likely to be restricted).

  According to the third aspect of the invention, various intake valve opening characteristics can be realized by making the valve opening phases of the first intake valve and the second intake valve variable in addition to changing the operating angle of the second intake valve. Therefore, an appropriate combustion state according to the operation state can be realized.

  According to the fourth aspect of the invention, the first intake valve and the second intake valve are driven by a common intake camshaft, and the phase variable mechanism changes the phase of the intake camshaft, thereby The valve opening phase can be changed. According to such 4th invention, a 1st valve mechanism, a 2nd valve mechanism, and a phase variable mechanism can be comprised simply, and further size reduction, weight reduction, and cost reduction can be achieved.

  According to the fifth aspect of the invention, the operation of both intake valves can be controlled so that more appropriate combustion according to the operating state of the diesel engine can be obtained.

  According to the sixth aspect of the present invention, the second intake valve closing timing can be delayed from the standard in the medium-speed intermediate load region, so that the actual compression ratio can be lowered. As a result, the compression end temperature is lowered, so that the combustion temperature can be lowered. Therefore, the NOx emission amount can be reduced. Note that the in-cylinder air amount slightly decreases due to the slow closing of the second intake valve, but there is no risk of smoke because of the medium load.

  According to the seventh aspect, in the high speed and high load region, the opening timing of the first intake valve and the second intake valve can be made earlier than the standard, and the closing timing of the second intake valve can be made later than the standard. This causes a large positive valve overlap in which both the exhaust valve and both intake valves are open. Therefore, at high speeds, the scavenging effect promotes the discharge of burned gas and the introduction of fresh air, and the intake. The exhaust efficiency can be increased. Further, since the closing timing of the second intake valve is relatively late, the intake inertia effect in the high speed range can be utilized to the maximum, and the charging efficiency can be improved. Thus, according to the seventh aspect of the invention, the output performance in the high speed and high load range can be improved.

  According to the eighth aspect of the invention, the positive valve overlap in which the exhaust valve and both intake valves are both open can be increased in the light and medium load region. Thereby, since the amount of internal EGR increases, the in-cylinder temperature can be increased. When the in-cylinder temperature increases, incomplete combustion is reduced, and the amount of HC emission can be reduced. Further, since the operating angle of the second intake valve is larger than the fixed operating angle of the first intake valve, the strength of the swirl can be reduced. Furthermore, since the closing timing of the second intake valve is late, a flow that flows backward from the cylinder to the second intake port is generated, and the swirl formed in the cylinder in the intake stroke can be canceled out. As a result, swirl can be sufficiently suppressed. For this reason, fuel overdiffusion due to swirl can be reliably prevented. Therefore, coupled with the effect of increasing the in-cylinder temperature, the HC emission amount can be sufficiently reduced.

  According to the ninth aspect of the invention, the operation of both intake valves can be controlled so that more appropriate combustion according to the operating state of the diesel engine can be obtained.

  According to the tenth aspect, the opening timing of the second intake valve can be made earlier than the standard in the low speed and light load region. For this reason, the positive valve overlap in which the exhaust valve and the second intake valve are both opened can be increased. In the low speed and light load region, the amount of internal EGR increases due to this positive valve overlap, so that the in-cylinder temperature can be increased. Further, since the operating angle of the second intake valve is larger than the fixed operating angle of the first intake valve, the strength of the swirl can be suppressed. That is, according to the tenth aspect, in the low speed light load region, the in-cylinder temperature can be increased and swirl can be sufficiently suppressed. For this reason, the amount of HC emission can be sufficiently reduced by their synergistic effect.

  According to the eleventh aspect, the opening timing of the second intake valve can be made earlier than the standard in the high speed and high load region. For this reason, the positive valve overlap in which the exhaust valve and the second intake valve are both opened can be increased. If a positive valve overlap is provided in the high speed region, the scavenging effect can promote the discharge of burned gas and the introduction of fresh air, and the intake and exhaust efficiency can be increased. Therefore, according to the eleventh aspect, the output performance in the high speed and high load region can be improved.

  According to the twelfth aspect, the positive valve overlap in which the exhaust valve and the second intake valve are both open can be increased in the low speed and light load region. For this reason, since the amount of internal EGR is further increased, the in-cylinder temperature can be sufficiently increased. According to the twelfth aspect of the invention, the closing timing of the first intake valve and the second intake valve can be advanced to approach the intake bottom dead center. In the low speed range where the intake inertia effect is small, the actual compression ratio is increased and the in-cylinder temperature can be increased by closing both intake valves near the intake bottom dead center. For this reason, according to the twelfth aspect, the in-cylinder temperature can be increased extremely effectively in the low speed light load region. Furthermore, since the operating angle of the second intake valve is larger than the fixed operating angle of the first intake valve, the strength of the swirl can be suppressed. Therefore, the HC emission amount can be sufficiently reduced due to the synergistic effect of increasing the in-cylinder temperature and suppressing swirl.

  According to the thirteenth aspect, the positive valve overlap in which the exhaust valve and the second intake valve are both open can be increased in the low speed and high load region. When the positive valve overlap is increased in the low-speed and high-load region, the scavenging effect can promote the discharge of burned gas and the introduction of fresh air, and the intake and exhaust efficiency can be improved. Further, according to the thirteenth aspect, the closing timing of the first intake valve and the second intake valve can be advanced to approach the intake bottom dead center. As a result, in the low speed range where the intake inertia effect is small, the optimal intake valve closing timing is obtained, and the charging efficiency is improved. Thus, according to the thirteenth aspect, both the intake / exhaust efficiency and the charging efficiency are improved. Therefore, the synergistic effect can improve the output performance in the low speed and high load region.

  Embodiments of the present invention will be described below with reference to the drawings. In addition, the same code | symbol is attached | subjected to the element which is common in each figure, and the overlapping description is abbreviate | omitted.

Embodiment 1 FIG.
[Description of system configuration]
FIG. 1 is a diagram for explaining a system configuration according to the first embodiment of the present invention. As shown in FIG. 1, the system of this embodiment includes a four-cycle diesel engine 10. The diesel engine 10 has a plurality of cylinders, and FIG. 1 shows a cross section of one of the cylinders.

  Each cylinder of the diesel engine 10 is provided with an injector 12 that directly injects fuel into the cylinder, an intake valve 14, an exhaust valve 16, and a piston 18.

  The diesel engine 10 includes two intake valves 14 per cylinder. In order to distinguish the two intake valves 14, hereinafter, one is called a first intake valve 14a and the other is called a second intake valve 14b.

  The diesel engine 10 includes a first valve mechanism 20 that drives the first intake valve 14a, a second valve mechanism 22 that drives the second intake valve 14b, a first intake valve 14a, and a second intake valve 14b. And a phase variable mechanism 24 that makes the valve opening phase variable. These mechanisms will be described later. Further, the exhaust valve 16 may be either driven by a variable valve mechanism that varies its valve opening characteristic or may be driven by a normal valve mechanism having a fixed valve opening characteristic.

  The system of the present embodiment further includes a crank angle sensor 28 that detects the rotation angle of the crankshaft 26 of the diesel engine 10, an accelerator opening sensor 30 that detects the accelerator opening, and an ECU (Electronic Control Unit) 50. I have. The ECU 50 is connected to the various sensors and actuators described above. The ECU 50 controls the operating state of the diesel engine 10 by operating each actuator according to a predetermined program based on the output of each sensor.

  FIG. 2 is a schematic plan view for explaining an intake port of the diesel engine 10 in the system shown in FIG. As shown in FIG. 2, the diesel engine 10 includes two intake ports, a first intake port 32a and a second intake port 32b, per cylinder. The first intake port 32a is opened and closed by the first intake valve 14a, and the second intake port 32b is opened and closed by the second intake valve 14b.

  Such a diesel engine 10 is configured such that the swirl ratio of the first intake port 32a alone is larger than the swirl ratio of the second intake port 32b alone. That is, the swirl ratio when air flows into the cylinder only from the first intake port 32a is made larger than the swirl ratio when air flows into the cylinder only from the second intake port 32b. .

  FIG. 3 is a side view showing the second valve mechanism 22. The second valve mechanism 22 is a variable operating angle mechanism capable of continuously changing the operating angle and lift amount (hereinafter simply referred to as “operating angle”) of the second intake valve 14b. The diesel engine 10 is provided with an intake camshaft 34 that is rotationally driven by a crankshaft 26 via a belt. The second valve mechanism 22 is provided between the cam 36 formed on the intake cam shaft 34 and the second intake valve 14b. The intake camshaft 34 rotates clockwise in FIG.

  On the other hand, the first valve mechanism 20 is provided between the same intake camshaft 34 and the first intake valve 14a, and is a mechanism for driving the first intake valve 14a at a fixed operating angle. That is, the first valve mechanism 20 is a normal valve mechanism that drives the first intake valve 14a at a fixed operating angle corresponding to the profile of the cam provided on the intake cam shaft 34. Therefore, the detailed illustration of the first valve mechanism 20 is omitted.

  Hereinafter, the second valve mechanism 22 will be described in detail with reference to FIG. The second valve mechanism 22 includes a control shaft 38 disposed in parallel with the intake cam shaft 34 and a control shaft drive mechanism (not shown) that can rotate the control shaft 38 within a predetermined angle range. is doing. The configuration of the control shaft drive mechanism is not particularly limited. For example, the control shaft drive mechanism includes a worm wheel fixed to one end of the control shaft 38, a worm gear meshing with the worm wheel, and a servo motor that rotationally drives the worm gear. be able to. In this case, the rotation position (rotation angle) of the control shaft 38 can be controlled by controlling the rotation direction and rotation amount of the servo motor.

  Further, the second valve mechanism 22 has a swing arm 40. The swing arm 40 is installed so as to be swingable about the control shaft 38. On the swing arm 40, a slider surface 42 is formed on the side facing the cam 36.

  A slider roller 44 is disposed between the swing arm 40 and the cam 36. The slider roller 44 includes a large roller and a small roller arranged on the same axis, and the large roller is in contact with the peripheral surface of the cam 36 and the small roller is in contact with the slider surface 42.

  The slider roller 44 is installed at the tip of the support arm 46 so as to be freely rotatable. The support arm 46 has an arc shape along the outer periphery of the control shaft 38. A control arm 48 that protrudes leftward in FIG. 3 is fixed to the control shaft 38. That is, the control arm 48 rotates integrally with the control shaft 38. A distal end portion of the control arm 48 and a proximal end portion of the support arm 46 are rotatably connected by a pin 52.

  With such a configuration, the slider roller 44 can be moved by rotating the control shaft 38. That is, when the control shaft 38 is rotated counterclockwise from the state shown in FIG. 3, the slider roller 44 is pushed by the support arm 46 and moves toward the distal end of the swing arm 40. When the control shaft 38 is rotated clockwise from that state, the slider roller 44 approaches the control shaft 38. FIG. 3 shows a state where the slider roller 44 is closest to the control shaft 38.

  The slider surface 42 has a curved surface (for example, an arc surface) such that the distance from the center of the cam 36 gradually increases toward the distal end side of the swing arm 40.

  A swing cam surface 54 is formed on the side of the swing arm 40 opposite to the slider surface 42. The rocking cam surface 54 includes a non-working surface (basic circle) 54a formed so that the distance from the rocking center 62 of the rocking arm 40, that is, the center of the control shaft 38 is constant, and the non-working surface. 54a, and a working surface 54b formed so that the distance from the swing center 62 gradually increases.

  Such a swing arm 40 is urged counterclockwise in FIG. 3 by a lost motion spring (not shown). Due to this urging force, the slider surface 42 of the swing arm 40 is pressed against the slider roller 44, and the slider roller 44 is pressed against the cam 36.

  The second valve mechanism 22 further includes a rocker arm 56 that presses the valve shaft end of the second intake valve 14b. The rocker arm 56 is provided with a rocker roller 58 that contacts the rocking cam surface 54. The rocker roller 58 is rotatably attached to an intermediate portion of the rocker arm 56. One end of the rocker arm 56 is in contact with the valve shaft end of the second intake valve 14 b, and the other end of the rocker arm 56 is supported by the hydraulic lash adjuster 60. The second intake valve 14b is urged in a closing direction, that is, a direction in which the rocker arm 56 is pushed up by a valve spring (not shown). The rocker roller 58 is pressed against the swing cam surface 54 of the swing arm 40 by this urging force and the hydraulic lash adjuster 60.

  In such a second valve mechanism 22, when the cam 36 rotates, the cam lift of the cam 36 is transmitted to the swing arm 40 via the slider roller 44, so that the swing arm 40 swings. When the swing arm 40 swings, the contact point between the swing cam surface 54 and the rocker roller 58 moves back and forth between the non-working surface 54a and the working surface 54b. When the rocker roller 58 is in contact with the non-operation surface 54a, the second intake valve 14 is closed without being lifted. When the rocker roller 58 is in contact with the working surface 54b, the second intake valve 14 is lifted and opened.

  When changing the operating angle of the second intake valve 14b, the control shaft 38 is rotated and the slider roller 44 is moved. In the state shown in FIG. 3, that is, the state in which the slider roller 44 is closest to the swing center 62, the swing width of the swing arm 40 is maximized, so that the operating angle of the second intake valve 14 b is maximum. Become.

  In contrast, as the control shaft 38 is rotated counterclockwise in FIG. 3 and the slider roller 44 is moved to the distal end side of the swing arm 40, the swing width of the swing arm 40 becomes smaller. This is one of the reasons for reducing the operating angle of the second intake valve 14b.

  As described above, the distance between the center of the cam 36 and the slider surface 42 decreases as the distance from the swing center 62 increases. Therefore, when the slider roller 44 is moved away from the swing center 62, the swing start position of the swing arm 40 moves in the clockwise direction in FIG. Therefore, after the cam crest of the cam 36 starts to contact the slider roller 44 and the swing arm 40 starts swinging, the contact point between the rocker roller 58 and the swing cam surface 54 shifts to the non-operation surface 54b, that is, The amount of rotation of the swing arm 40 required until the second intake valve 14b starts to lift increases as the slider roller 44 moves away from the swing center 62. This is also one of the reasons for reducing the operating angle of the second intake valve 14b.

  In this way, in the second valve mechanism 22, the slider roller 44 is moved to the distal end side of the swing arm 40 by rotating the control shaft 38 counterclockwise from the state of FIG. 3 for the above two reasons. As the operating angle of the second intake valve 14b decreases.

  Incidentally, since the cam 36 rotates clockwise in FIG. 3, the timing at which the cam crest of the cam 36 starts to contact the slider roller 44 becomes earlier as the slider roller 44 is moved to the tip side of the swing arm 40. Become. That is, as the operating angle of the second intake valve 14b is decreased, the timing at which the swing arm 40 starts to swing is earlier. On the other hand, as described above, the smaller the operating angle of the second intake valve 14b, the more the rotation of the swing arm 40 required from when the swing arm 40 starts to swing until the second intake valve 14b starts to lift. The amount gets bigger. Therefore, in the second valve mechanism 22 of the present embodiment, when the operating angle of the second intake valve 14b is decreased, the swing arm 40 starts swinging, but the swing arm 40 swings. Since the time required from the start until the second intake valve 14b starts to be lifted becomes longer, they are offset and the opening timing of the second intake valve 14b is not changed. In other words, when the operating angle of the second intake valve 14b is changed, the second valve mechanism 22 of the present embodiment does not change its opening timing but only the closing timing.

  FIG. 4 is a diagram schematically showing lift curves of the first intake valve 14a and the second intake valve 14b in the present embodiment. As shown in FIG. 4, in the present embodiment, the opening timing of the first intake valve 14a and the opening timing of the second intake valve 14b are configured to be substantially the same. The fixed operating angle of the first intake valve 14a is a standard size. The operating angle of the second intake valve 14b is set to a minimum operating angle smaller than the fixed operating angle of the first intake valve 14a and a maximum operating angle larger than the fixed operating angle of the first intake valve 14a by the second valve mechanism 22 described above. It can be continuously changed between the working angles. At this time, since the opening timing of the second intake valve 14b does not change, the opening timing of the first intake valve 14a and the opening timing of the second intake valve 14b are always substantially the same.

  As the operating angle of the second intake valve 14b is smaller than the fixed operating angle of the first intake valve 14a, the amount of air flowing into the cylinder through the second intake port 32b having a small swirl ratio decreases, while the swirl ratio The amount of air flowing into the cylinder through the large first intake port 32a increases. For this reason, the swirl in a cylinder can be strengthened.

  In contrast, as the operating angle of the second intake valve 14b is made larger than the fixed operating angle of the first intake valve 14a, the amount of air flowing into the cylinder through the second intake port 32b having a small swirl ratio increases. Thus, the amount of air flowing into the cylinder through the first intake port 32a having a large swirl ratio decreases. For this reason, the swirl in a cylinder can be weakened.

  In this way, in the diesel engine 10 of the present embodiment, by changing the operating angle of the second intake valve 14b, the amount of air flowing into the cylinder through the first intake port 32a having a large swirl ratio, and the swirl The ratio of the amount of air flowing into the cylinder through the second intake port 32b having a small ratio can be freely changed. For this reason, the intensity | strength of the swirl formed in a cylinder can be changed with a high freedom degree, and the optimal swirl intensity according to the driving | running state is realizable. Therefore, good combustion according to the driving state can be performed, and emission, fuel consumption, torque fluctuation, and the like can be improved.

  Further, in the diesel engine 10 of the present embodiment, a working angle variable mechanism such as the second valve mechanism 22 may be provided only in the second intake valve 14b. That is, the above-described effect can be obtained without providing a variable operating angle mechanism in the first intake valve 14a. For this reason, compared with the case where a working angle variable mechanism is provided in both the 1st intake valve 14a and the 2nd intake valve 14b, cost reduction and weight reduction can be aimed at. In addition, it is possible to reduce the size and to be mounted on a vehicle.

  Next, the phase variable mechanism 24 will be described. The phase variable mechanism 24 is provided between a pulley (not shown) provided at an end of the intake camshaft 34 and the intake camshaft 34, and the intake camshaft 34 is rotated with respect to this pulley. This is a variable valve timing mechanism capable of changing the relative phase of the intake camshaft 34 with respect to the pulley. Since such a mechanism is publicly known, detailed illustration and description thereof will be omitted.

  According to the phase variable mechanism 24, the phase of the valve opening period of the first intake valve 14a and the second intake valve 14b can be changed (advanced or retarded). As described above, the intake camshaft 34 is common to the first intake valve 14a and the second intake valve 14b. Therefore, when the relative phase of the intake camshaft 34 is changed by the phase variable mechanism 24, the phase of the valve opening period of the first intake valve 14a and the phase of the valve opening period of the second intake valve 14b are integrated. Change (overall).

  In the present embodiment, the diesel engine is obtained by combining the variable operating angle of the second intake valve 14b by the second valve mechanism 22 and the variable phase of the first intake valve 14a and the second intake valve 14b by the phase variable mechanism 24. Various combustion controls according to the ten operating states are possible. For this reason, an optimal combustion state can be realized according to the operating state of the diesel engine 10. Specific examples thereof will be described below.

  FIG. 5 shows the lift curves of both intake valves when the operating angle of the second intake valve 14b is a large operating angle (the operating angle larger than the operating angle of the first intake valve 14a) and the phase variable mechanism 24 is in the reference state. It is a figure shown typically. On the other hand, FIG. 6 is a diagram schematically showing lift curves of both intake valves when the operating angle of the second intake valve 14b is set to the same large operating angle and the phase variable mechanism 24 is in the advanced state.

  As shown in FIG. 5, when the phase variable mechanism 24 is in the reference state, the opening timing of the first intake valve 14a and the second intake valve 14b is in the vicinity of the top dead center (before the top dead center).

  In the present embodiment, when the diesel engine 10 is operated in the middle speed / medium load region, it is preferable to perform control so that the intake valve opening characteristics as shown in FIG. 5 are obtained. Thereby, there is an advantage that the amount of NOx emission is reduced. The reason is as follows.

  The substantial compression stroke starts when the intake valve is closed and the cylinder is sealed. If the intake valve opening characteristic as shown in FIG. 5 is used, the closing timing of the second intake valve 14b is delayed from the standard timing, so that the substantial compression ratio (hereinafter referred to as “actual compression ratio”). Becomes smaller. For this reason, since the compression end temperature is lowered, the ignition delay is expanded, and the combustion period shifts to the expansion stroke side. As a result, the combustion temperature is lowered, so that the amount of NOx emissions can be reduced. In this case, since the part of the air once sucked into the cylinder returns to the second intake port 32b due to the slow closing of the second intake valve 14b, the in-cylinder air amount is somewhat reduced. Since the fuel injection amount is not so large, the generation of smoke can be sufficiently suppressed.

  Further, in the present embodiment, when the diesel engine 10 is operated in the high speed and high load region, it is preferable to perform control so that the intake valve opening characteristics as shown in FIG. 6 are obtained. Thereby, there is an advantage that the output performance in the high speed and high load region is improved. The reason is as follows.

  If the intake valve opening characteristic is as shown in FIG. 6, the opening timing of the first intake valve 14a and the second intake valve 14b is earlier than the standard timing (around the top dead center). As a result, a large positive valve overlap occurs in which the exhaust valve 16 and the first intake valve 14a and the second intake valve 14b are both open. If a positive valve overlap is provided in the high speed region, the scavenging effect can promote the discharge of burned gas and the introduction of fresh air, and the intake and exhaust efficiency can be increased. Further, in the intake valve opening characteristics as shown in FIG. 6, since the second intake valve 14b has a large operating angle, the closing timing of the second intake valve 14b is later than the standard timing (near bottom dead center). . For this reason, in the high speed range, the intake inertia effect can be fully utilized, and the charging efficiency can be improved. In this way, the intake / exhaust efficiency and the charging efficiency can be increased in the high-speed and high-load region, so that the output performance can be improved.

  Further, in the present embodiment, when the load of the diesel engine 10 is in a light / medium load region, it is preferable to perform control so that the intake valve opening characteristic as shown in FIG. 6 is obtained. This has the advantage that the amount of HC emission is reduced. The reason is as follows.

  When the intake valve opening characteristic as shown in FIG. 6 is used, as described above, a large positive valve overlap occurs. In the light and medium load region, the amount of internal EGR increases due to this positive valve overlap, so that the in-cylinder temperature can be increased. When the in-cylinder temperature rises, incomplete combustion is reduced and HC emissions are reduced.

  By the way, in the light and medium load region, if the swirl is strong, the fuel is excessively diffused by the swirl, so that the air-fuel mixture becomes too lean and incomplete combustion is likely to occur, and the HC emission amount is likely to increase. On the other hand, when the intake valve opening characteristic as shown in FIG. 6 is used, since the operating angle of the second intake valve 14b is larger than the fixed operating angle of the first intake valve 14a, the second intake port 32b having a small swirl ratio is provided. The ratio of the amount of air flowing through and into the cylinder is large, and the ratio of the amount of air flowing into the cylinder through the first intake port 32a having a large swirl ratio is small. Therefore, the strength of the swirl can be suppressed. In the case of the intake valve opening characteristics as shown in FIG. 6, the swirl is further suppressed for the following reason.

  In the intake valve opening characteristics as shown in FIG. 6, the closing timing of the second intake valve 14b is delayed, so that the second intake valve 14b is opened even after the piston 18 starts to rise after passing through the bottom dead center. There is a period. For this reason, in the light and medium load region, a phenomenon occurs in which the air in the cylinder flows backward to the second intake port 32b from when the first intake valve 14a is closed to when the second intake valve 14b is closed. At this time, the air that is going to flow backward from the cylinder to the second intake port 32b flows into the second intake port 32b while turning in the direction opposite to the turning direction of the swirl formed in the cylinder in the intake stroke. For this reason, since the swirl formed in the cylinder in the intake stroke is canceled, the swirl can be sufficiently suppressed.

  As described above, when the intake valve opening characteristic is as shown in FIG. 6 in the light and medium load region, the internal EGR increases and the in-cylinder temperature increases, and the swirl can be sufficiently suppressed. These synergistic effects can sufficiently reduce HC emissions.

Embodiment 2. FIG.
Next, the second embodiment of the present invention will be described with reference to FIG. 7 to FIG. 9. The difference from the above-described first embodiment will be mainly described, and the same matters will be described. Simplify or omit.

  In the present embodiment, the second valve mechanism 22 is configured such that when the operating angle of the second intake valve 14b is changed, the closing timing does not change, but only the opening timing changes. To do. Note that the second valve mechanism 22 of this embodiment can be realized by changing the arrangement or cam shape of the parts of the second valve mechanism 22 described in the first embodiment. Detailed illustration and description thereof will be omitted.

  FIG. 7 is a diagram schematically showing lift curves of the first intake valve 14a and the second intake valve 14b in the present embodiment. As shown in FIG. 7, in the present embodiment, the closing timing of the first intake valve 14a and the closing timing of the second intake valve 14b are configured to be substantially the same. The fixed operating angle of the first intake valve 14a is a standard size. The operating angle of the second intake valve 14b is set by the second valve mechanism 22 so that the minimum operating angle is smaller than the fixed operating angle of the first intake valve 14a and the maximum operating angle is larger than the fixed operating angle of the first intake valve 14a. And can be changed continuously. At this time, since the closing timing of the second intake valve 14b does not change, the closing timing of the first intake valve 14a and the closing timing of the second intake valve 14b are always substantially the same.

  In the diesel engine 10 according to this embodiment, by changing the operating angle of the second intake valve 14b, it flows into the cylinder through the first intake port 32a having a large swirl ratio, as in the first embodiment. It is possible to freely change the ratio between the amount of air that flows and the amount of air that flows into the cylinder through the second intake port 32b having a small swirl ratio. For this reason, the intensity | strength of the swirl formed in a cylinder can be changed with a high freedom degree, and the optimal swirl intensity according to the driving | running state is realizable. Therefore, good combustion according to the driving state can be performed, and emission, fuel consumption, torque fluctuation, and the like can be improved.

  Further, in the diesel engine 10 of the present embodiment, similarly to the first embodiment, a working angle variable mechanism such as the second valve mechanism 22 may be provided only in the second intake valve 14b. That is, the above-described effect can be obtained without providing a variable operating angle mechanism in the first intake valve 14a. For this reason, compared with the case where a working angle variable mechanism is provided in both the 1st intake valve 14a and the 2nd intake valve 14b, cost reduction and weight reduction can be aimed at. In addition, it is possible to reduce the size and to be mounted on a vehicle.

  Further, in the present embodiment as well, as in the first embodiment, the operating angle of the second intake valve 14b by the second valve mechanism 22 and the first intake valve 14a and the second intake valve 14b by the phase variable mechanism 24 are changed. In combination with the variable phase, various combustion controls according to the operating state of the diesel engine 10 are possible. For this reason, an optimal combustion state can be realized according to the operating state of the diesel engine 10. Specific examples thereof will be described below.

  FIG. 8 shows the lift curves of both intake valves when the operating angle of the second intake valve 14b is a large operating angle (the operating angle larger than the operating angle of the first intake valve 14a) and the phase variable mechanism 24 is in the reference state. It is a figure shown typically. On the other hand, FIG. 9 is a diagram schematically showing lift curves of both intake valves when the operating angle of the second intake valve 14b is set to the same large operating angle and the phase variable mechanism 24 is in the advanced state.

  As shown in FIG. 8, when the phase variable mechanism 24 is in the reference state, the opening timing of the first intake valve 14a is in the vicinity of the top dead center (before the top dead center).

  In the present embodiment, when the diesel engine 10 is operated in the low speed and light load region, it is preferable to perform control so that the intake valve opening characteristics as shown in FIG. 8 are obtained. This has the advantage that the amount of HC emission is reduced. The reason is as follows.

  If the intake valve opening characteristic is as shown in FIG. 8, the opening timing of the second intake valve 14b is earlier than the standard timing (near top dead center). As a result, a positive valve overlap in which the exhaust valve 16 and the second intake valve 14b are both opened is generated. In the low speed and light load region, the amount of internal EGR increases due to this positive valve overlap, so that the in-cylinder temperature can be increased. Further, since the operating angle of the second intake valve 14b is larger than the fixed operating angle of the first intake valve 14a, the ratio of the amount of air flowing into the cylinder through the second intake port 32b having a small swirl ratio is large, and the swirl ratio Since the ratio of the amount of air flowing into the cylinder through the large first intake port 32a is small, the strength of the swirl can be suppressed. As described above, when the intake valve opening characteristic is as shown in FIG. 8 in the low speed and light load region, the internal EGR increases and the in-cylinder temperature rises, and the swirl can be sufficiently suppressed. For this reason, the amount of HC emission can be sufficiently reduced by their synergistic effect.

  Further, in the present embodiment, when the diesel engine 10 is operated in the high speed and high load region, it is preferable to perform control so that the intake valve opening characteristic as shown in FIG. 8 is obtained. Thereby, there is an advantage that the output performance in the high speed and high load region is improved. The reason is as follows.

  When the intake valve opening characteristic as shown in FIG. 8 is used, as described above, there is a large positive valve overlap in which the exhaust valve 16 and the second intake valve 14b are both open. If a positive valve overlap is provided in the high speed region, the scavenging effect can promote the discharge of burned gas and the introduction of fresh air, and the intake and exhaust efficiency can be increased. As a result, the output performance in the high speed and high load region can be improved.

  Further, in the present embodiment, when the diesel engine 10 is operated in the low speed and light load region, it is preferable to perform control so that the intake valve opening characteristics as shown in FIG. 9 are obtained. This has the advantage that the amount of HC emission is reduced. The reason is as follows.

  When the intake valve opening characteristic as shown in FIG. 9 is adopted, a positive valve overlap in which both the exhaust valve 16 and the second intake valve 14b are opened occurs more greatly than in the case of FIG. For this reason, since the amount of internal EGR is further increased, the in-cylinder temperature can be further increased. Further, the in-cylinder temperature is increased for the following reason.

  With the intake valve opening characteristics as shown in FIG. 9, the closing timing of the first intake valve 14a and the second intake valve 14b can be advanced to approach the intake bottom dead center. In the low speed range where the intake inertia effect is small, the actual compression ratio is increased and the in-cylinder temperature can be increased by closing both intake valves near the intake bottom dead center.

  For this reason, the in-cylinder temperature can be increased extremely effectively by adopting the intake valve opening characteristics as shown in FIG. 9 in the low speed and light load region. Furthermore, since the operating angle of the second intake valve 14b is larger than the fixed operating angle of the first intake valve 14a, the strength of the swirl can be suppressed. Therefore, the HC emission amount can be further reduced as compared with the case of FIG.

  In the present embodiment, the low-speed light load region is divided into a normal light load region and an extremely light load region having a smaller load, and the intake valve opening as shown in FIG. 8 is performed in the normal light load region. It is more preferable to set the intake valve opening characteristic as shown in FIG. 9 in the extremely light load region.

  Further, in the present embodiment, when the diesel engine 10 is operated in the low speed and high load region, it is preferable to perform control so that the intake valve opening characteristic as shown in FIG. 9 is obtained. Thereby, there is an advantage that the output performance in the low speed and high load region is improved. The reason is as follows.

  When the intake valve opening characteristic as shown in FIG. 9 is used, as described above, a large positive valve overlap occurs. When the positive valve overlap is increased in the low-speed and high-load region, the scavenging effect can promote the discharge of burned gas and the introduction of fresh air, and the intake and exhaust efficiency can be improved. Further, as described above, the closing timing of the first intake valve 14a and the second intake valve 14b approaches the intake bottom dead center, so that the optimal intake valve closing timing is obtained in the low speed region where the intake inertia effect is small, and the charging efficiency is increased. improves. Thus, since both the intake and exhaust efficiency and the charging efficiency are improved, the output performance in the low speed and high load region can be improved by the synergistic effect.

It is a figure for demonstrating the system configuration | structure of Embodiment 1 of this invention. It is a typical top view for demonstrating the intake port of the diesel engine in the system shown in FIG. It is a side view which shows a 2nd valve mechanism. It is a figure which shows typically the lift curve of the 1st intake valve and the 2nd intake valve in Embodiment 1 of this invention. It is a figure which shows typically the lift curve of the 1st intake valve and the 2nd intake valve in Embodiment 1 of this invention. It is a figure which shows typically the lift curve of the 1st intake valve and the 2nd intake valve in Embodiment 1 of this invention. It is a figure which shows typically the lift curve of the 1st intake valve and the 2nd intake valve in Embodiment 2 of this invention. It is a figure which shows typically the lift curve of the 1st intake valve and the 2nd intake valve in Embodiment 2 of this invention. It is a figure which shows typically the lift curve of the 1st intake valve and the 2nd intake valve in Embodiment 2 of this invention.

Explanation of symbols

10 Diesel Engine 12 Injector 14a First Intake Valve 14b Second Intake Valve 16 Exhaust Valve 18 Piston 20 First Valve Mechanism 22 Second Valve Mechanism 24 Phase Variable Mechanism 26 Crankshaft 28 Crank Angle Sensor 30 Accelerator Opening Sensor 32a First 1 intake port 32b second intake port 34 intake cam shaft 36 cam 38 control shaft 40 swing arm 42 slider surface 44 slider roller 46 support arm 48 control arm 50 ECU
54 rocking cam surface 54a non-working surface 54b working surface 56 rocker arm 58 rocker roller 62 rocking center

Claims (13)

A first intake valve;
A second intake valve;
A first intake port leading to the first intake valve;
A second intake port leading to the second intake valve;
A first valve mechanism for driving the first intake valve at a fixed operating angle;
A second valve that drives the second intake valve with a variable operating angle between a minimum operating angle smaller than the fixed operating angle of the first intake valve and a maximum operating angle larger than the fixed operating angle of the first intake valve. Mechanism,
With
A diesel engine characterized in that a swirl ratio of the first intake port alone is larger than a swirl ratio of the second intake port alone.   In a predetermined operating region, the operating angle of the second intake valve is made larger than the fixed operating angle of the first intake valve, and after the first intake valve is closed, the air in the cylinder is increased as the piston rises. 2. The diesel engine according to claim 1, wherein a part of the second intake valve is caused to flow backward to the second intake port through the second intake valve and then the second intake valve is closed.   The diesel engine according to claim 1 or 2, further comprising a phase variable mechanism that changes a phase of a valve opening period of the first intake valve and the second intake valve. The first intake valve and the second intake valve are driven by a common intake camshaft,
The diesel engine according to claim 3, wherein the phase variable mechanism is configured to change a phase of the intake camshaft. The second valve mechanism changes the operating angle by changing the closing timing while keeping the opening timing of the second intake valve constant,
The diesel engine according to claim 3 or 4, wherein the opening timing of the first intake valve and the opening timing of the second intake valve are substantially the same.   The control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism in a reference state in a medium-speed intermediate load region. 5. The diesel engine according to 5.   The high-speed and high-load region includes control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism to an advanced state. 5. The diesel engine according to 5.   The control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism to an advanced angle state in a light and medium load region. 5. The diesel engine according to 5. The second valve mechanism changes the operating angle by changing the opening timing while keeping the closing timing of the second intake valve constant,
The diesel engine according to claim 3 or 4, wherein the closing timing of the first intake valve and the closing timing of the second intake valve are substantially the same.   The low-speed and light-load region includes control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism as a reference state. The listed diesel engine.   The high-speed and high-load region includes control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism as a reference state. The listed diesel engine.   The low-speed light load region includes control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism to an advanced state. 9. The diesel engine according to 9.   The low-speed and high-load region includes control means for making the operating angle of the second intake valve larger than the fixed operating angle of the first intake valve and setting the phase variable mechanism to an advanced state. 9. The diesel engine according to 9.

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