internal combustion engine

Abstract

Provided is an internal combustion engine having excellent combustion efficiency. The internal combustion engine is provided with: a combustion chamber; a suction path that is connected to the combustion chamber; an exhaust path that is connected to the combustion chamber; a first suction valve that opens and closes between the suction path and the combustion chamber; a second suction valve that is disposed adjacent to the first suction valve and opens and closes between the suction path and the combustion chamber; a first exhaust valve that opens and closes between the exhaust path and the combustion chamber; and a second exhaust valve that is disposed adjacent to the first exhaust valve and opens and closes between the exhaust path and the combustion chamber. The second suction valve opens after the first suction valve. The second exhaust valve opens after the first exhaust valve. Respective lift amounts of the first exhaust valve and the second exhaust valve are smaller than those of the first suction valve and the second suction valve.

Description

[Technical Field]

[0001] The present disclosure relates to internal combustion engines. [Background technology]

[0002] Conventionally, internal combustion engines equipped with multiple intake valves and multiple exhaust valves per cylinder have been known (see, for example, Patent Document 1). In the internal combustion engine of Patent Document 1, each cylinder is equipped with two exhaust valves, and the valve opening and closing timing of one of the two exhaust valves is retarded relative to the valve opening and closing timing of the other valve. In the internal combustion engine of Patent Document 1, by differing the valve opening and closing timing of the exhaust valves in this way, a swirl flow is generated in the exhaust gas flowing back into the cylinder. [Prior art documents] [Patent documents]

[0003] [Patent Document 1] Japanese Patent Application Laid-Open No. 2006-161666 Summary of the Invention [Problem to be solved by the invention]

[0004] The internal combustion engine of Patent Document 1 utilizes such exhaust-side swirl flow to improve the deterioration of combustibility caused by a decrease in intake charging efficiency in the low load range. Specifically, the internal combustion engine of Patent Document 1 uses a swirl control valve to generate a swirl flow on the intake side as well, and mixes the swirl flow on the exhaust side with the swirl flow on the intake side to promote the flow of the air-fuel mixture in the cylinder.

[0005] However, in high load ranges, the swirl flow on the exhaust side and the swirl flow on the intake side collide, resulting in insufficient scavenging within the cylinder, which can have a negative impact on combustion within the cylinder.

[0006] An object of the present disclosure is to provide an internal combustion engine with good combustion efficiency. [Means for solving the problem]

[0007] The internal combustion engine of the present disclosure comprises a combustion chamber, an intake passage connected to the combustion chamber, an exhaust passage connected to the combustion chamber, a first intake valve that opens and closes between the intake passage and the combustion chamber, a second intake valve arranged adjacent to the first intake valve and opening and closing between the intake passage and the combustion chamber, a first exhaust valve that opens and closes between the exhaust passage and the combustion chamber, and a second exhaust valve arranged adjacent to the first exhaust valve and opening and closing between the exhaust passage and the combustion chamber, wherein the second intake valve opens slower than the first intake valve, the second exhaust valve opens slower than the first exhaust valve, and the lift amounts of the first exhaust valve and the second exhaust valve are smaller than the lift amounts of the first intake valve and the second intake valve, respectively.

[0008] In this internal combustion engine, the second intake valve opens later than the first intake valve, generating a swirl flow in the intake air. The second exhaust valve also opens later than the first exhaust valve, so the swirl flow generated during intake is more likely to be maintained for a longer period of time than when the first and second exhaust valves are opened simultaneously. This strengthens the flow inside the combustion chamber and improves combustion efficiency.

[0009] Furthermore, with this internal combustion engine, the second exhaust valve also opens later than the first exhaust valve, resulting in less exhaust interference than when the first and second exhaust valves open simultaneously. Furthermore, the lift amounts of the first and second exhaust valves are smaller than the lift amounts of the first and second intake valves. This allows for a lower maximum exhaust pressure than if the lift amounts of these valves were the same. By reducing exhaust interference and lowering the maximum exhaust pressure, internal EGR, particularly in high-load engines, is reduced, preventing knocking. This results in improved torque and suppression of exhaust backflow. Suppressing exhaust backflow improves combustion efficiency in high-load engines. [Effects of the Invention]

[0010] According to the present disclosure, an internal combustion engine with good combustion efficiency can be provided. [Brief explanation of the drawings]

[0011] [Figure 1] 1 is a system diagram of a vehicle equipped with an internal combustion engine according to one embodiment of the present disclosure. [Figure 2] 1 is a cross-sectional view of an internal combustion engine according to one embodiment of the present disclosure. [Figure 3] 1 is a bottom view of a combustion chamber of an internal combustion engine according to one embodiment of the present disclosure; [Figure 4] 1 illustrates a camshaft of an internal combustion engine according to one embodiment of the present disclosure. [Figure 5] FIG. 2 is a diagram illustrating lift curves of intake and exhaust valves of an internal combustion engine according to one embodiment of the present disclosure. [Figure 6] FIG. 4 is a cross-sectional view of FIG. 3 . [Figure 7] FIG. 4 is a cross-sectional view of FIG. 3 taken along line AA. [Figure 8] FIG. 4 is a cross-sectional view of FIG. 3 . [Figure 9] A diagram showing the b-b cross section of Figure 3. [Figure 10] 2A and 2B are diagrams showing the flow of the mixture in the combustion chamber during the intake, compression, and expansion strokes of an internal combustion engine according to an embodiment of the present disclosure. [Figure 11] FIG. 2 is a diagram illustrating the flow of exhaust gas in a combustion chamber during the exhaust stroke of an internal combustion engine according to one embodiment of the present disclosure. [Figure 12] 6 is a graph showing an example of a state in which exhaust interference is reduced; DETAILED DESCRIPTION OF THE INVENTION

[0012] Hereinafter, an embodiment of the present disclosure will be described with reference to the drawings. In the drawings, FS indicates the front side of the electric vehicle C, BS indicates the rear side of the electric vehicle C, RS indicates the right side of the electric vehicle C, LS indicates the left side of the electric vehicle C, US indicates the upper side of the electric vehicle C, and DS indicates the lower side of the electric vehicle C.

[0013] As shown in FIG. 1, an internal combustion engine 1 is mounted on an electric vehicle C that drives wheels C1 using the internal combustion engine 1 and a motor (FrM) 2 as power sources. The electric vehicle C has the motor 2, a generator (GEN) 4, a drive battery (BT) 6 including a secondary battery such as a lithium-ion battery, and a transaxle 8. The transaxle 8 has multiple gears and a clutch 8a. The internal combustion engine 1 is connected to the generator 4 and axles 10 via the transaxle 8. When the clutch 8a is in a disengaged state, the transaxle 8 cuts off power transmission between the internal combustion engine 1 and the axles 10, and when the clutch 8a is in a engaged state, the power of the internal combustion engine 1 is transmitted to the axles 10. The motor 2 is connected to the axles 10 via the transaxle 8. The electric vehicle C may also have a vehicle control device 12, an engine control device 14 that controls the internal combustion engine 1, an accelerator pedal 16 operated by a user of the electric vehicle C, an inverter 18 that controls the motor 2 and the generator 4, and a charging button (not shown).

[0014] The electric vehicle C of this embodiment has various modes, including an EV mode, a series mode, a parallel mode, and a charge mode. In the EV mode, the electric vehicle C drives the motor 2 using electric power from the drive battery 6. In the series mode, the electric vehicle C drives the generator 4 using the internal combustion engine 1 and drives the motor 2 using electric power generated by the generator 4. In the parallel mode, the electric vehicle C engages the clutch 8a and drives the axle 10 using power from the internal combustion engine 1. In the charge mode, the electric vehicle C drives the generator 4 using the internal combustion engine 1 and stores the electric power generated by the generator 4 in the drive battery 6. In the electric vehicle C, the vehicle control device 12 switches between the various modes depending on the depression state of the accelerator pedal 16 and the operation state of the charge button, and controls the motor 2 and the generator 4 via the inverter 18 and causes the engine control device 14 to control the internal combustion engine 1.

[0015] Such an electric vehicle C uses the motor 2 when the electric vehicle C runs in a low load range. The electric vehicle C uses the internal combustion engine 1 in a series mode, a parallel mode, and a charging mode. In these series mode, parallel mode, and charging mode, the internal combustion engine 1 is mainly operated in a high load operating range. For this reason, the internal combustion engine 1 used in such an electric vehicle C is required to have good combustion efficiency in the high load operating range.

[0016] As shown in FIG. 2, the internal combustion engine 1 includes a cylinder head 1a, a cylinder block 1b, a plurality of cylinders 20, an intake port (an example of an intake passage) 22, an exhaust port (an example of an exhaust passage) 24, an intake camshaft 26, an exhaust camshaft 28, a plurality of intake valves 30, a plurality of exhaust valves 32, a piston 34, a crankshaft 36, a spark plug 38, and a fuel injection valve 40.

[0017] As shown in FIG. 1, the internal combustion engine 1 of this embodiment is a horizontally mounted type in which the cylinders 20 of the internal combustion engine 1 are arranged side by side along the direction of the axles 10 of the wheels C1 of the electric vehicle C (the left-right direction of the electric vehicle C).

[0018] A plurality of cylinders 20 are formed in the cylinder block 1b. In this embodiment, four cylinders 20 are formed side by side. That is, the internal combustion engine 1 of this embodiment is a four-cylinder in-line engine. As shown in FIG. 2, a piston 34 connected to a crankshaft 36 via a connecting rod is slidably housed in each cylinder 20. In each cylinder 20, a combustion chamber 20a is formed between the piston 34 and the lower surface of the cylinder head 1a.

[0019] As shown in FIG. 2, in this embodiment, the combustion chamber 20a is a pent roof type combustion chamber 20a with a ridgeline formed in the left-right direction of the internal combustion engine 1. The intake port 22 is formed in the cylinder head 1a and connected to a slope on the intake side (front side in FIG. 2) of the combustion chamber 20a. The exhaust port 24 is formed in the cylinder head 1a and connected to a slope on the exhaust side (rear side in FIG. 2) of the combustion chamber 20a. An ignition plug 38 is disposed at the center O of the combustion chamber 20a (see FIG. 3). That is, the internal combustion engine 1 of this embodiment is a gasoline engine.

[0020] As shown in FIGS. 1 and 2, an intake port 22 is provided in each of the four cylinders 20 and is a passage that supplies intake air to the combustion chamber 20a. As shown in FIG. 3, the intake port 22 is branched into left and right portions by an intake port wall 22a, with a first intake port 22b located on the right side and a second intake port 22c located on the left side. In this embodiment, a fuel injection valve 40 is located forward of the branch of the intake port 22. Fuel is supplied to the fuel injection valve 40 from a fuel tank 54 (see FIG. 1), and the fuel injected from the fuel injection valve 40 and air passing through the intake port 22 are mixed and supplied to the combustion chamber 20a.

[0021] As shown in Figures 1 and 2, an exhaust port 24 is provided in each of the four cylinders 20 and serves as a passage for discharging exhaust gas generated after combustion of the air-fuel mixture in the combustion chamber 20a. As shown in Figure 3, the exhaust port 24 is branched into left and right portions by an exhaust port wall 24a, with a first exhaust port 24b located on the right side and a second exhaust port 24c located on the left side. As shown in Figure 2, the exhaust ports 24 of each cylinder 20 are connected to an exhaust manifold 25. The exhaust manifold 25 collects and discharges the exhaust gases flowing through the exhaust ports 24 located in each cylinder 20.

[0022] The intake valve 30 opens and closes the communication between the intake port 22b and the combustion chamber 20a. As shown in FIG. 3, the intake valve 30 has a first intake valve 30a located on the right side and a second intake valve 30b located on the left side and adjacent to the first intake valve 30a. The first intake valve 30a opens and closes the communication between the first intake port 22b and the combustion chamber 20a. The second intake valve 30b opens and closes the communication between the second intake port 22c and the combustion chamber 20a.

[0023] The exhaust valve 32 opens and closes the gap between the exhaust port 24 and the combustion chamber 20a. As shown in Figure 3, the exhaust valve 32 has a first exhaust valve 32a located on the left side and a second exhaust valve 32b located on the right side and adjacent to the first exhaust valve 32a. The first exhaust valve 32a opens and closes the gap between the second exhaust port 24c and the combustion chamber 20a. The second exhaust valve 32b opens and closes the gap between the first exhaust port 24b and the combustion chamber 20a.

[0024] As shown in FIG. 2, in this embodiment, the intake valve 30 is a direct-stroke valve that is driven by the cam of the intake camshaft 26 pushing out the intake valve 30. As shown in FIG. 4(a), the intake camshaft 26 in this embodiment has a first intake cam 26a that drives the first intake valve 30a and a second intake cam 26b that drives the second intake valve 30b. The intake camshaft 26 has one set of the first intake cam 26a and the second intake cam 26b for each cylinder 20. For example, if there are four cylinders 20, four first intake cams 26a and four second intake cams 26b are provided. The first intake valve 30a oscillates along the cam profile of the first intake cam 26a, resulting in a lift curve as shown in FIG. 5. The second intake valve 30b oscillates along the cam profile of the second intake cam 26b, resulting in a lift curve as shown in FIG. 5.

[0025] As shown in FIG. 5, the first intake cam 26a and the second intake cam 26b are split cams with different phases and cam profiles. Specifically, the second intake cam 26b is retarded more than the first intake cam 26a. This causes the second intake valve 30b to open later than the first intake valve 30a. Furthermore, the cam profiles of the first intake cam 26a and the second intake cam 26b are designed so that the first intake valve 30a changes from the maximum lift position P1 to the closed position P0 at the same time as the second intake valve 30b changes from the maximum lift position P1 to the closed position P0. In other words, the first intake valve 30a and the second intake valve 30b close simultaneously. The opening period of the second intake valve 30b is shorter than that of the first intake valve 30a. The cam profile of the second intake cam 26b may be either linearly symmetrical or asymmetrical, with the upward profile of the cam from the valve-closed position P0 to the maximum lift position P1 and the downward profile of the cam from the maximum lift position P1 to the valve-closed position P0. Examples of an asymmetrical cam profile of the second intake cam 26b include one in which the upward period is shorter than the downward period, and one in which the downward period is shorter than the upward period.

[0026] As shown in FIG. 2, in this embodiment, the exhaust valves 32 are direct-stroke valves that are driven by the cams of the exhaust camshaft 28 pushing out the exhaust valves 32. As shown in FIG. 4(b), the exhaust camshaft 28 in this embodiment has a first exhaust cam 28a that drives the first exhaust valve 32a and a second exhaust cam 28b that drives the second exhaust valve 32b. The exhaust camshaft 28 has one set of the first exhaust cam 28a and the second exhaust cam 28b for each cylinder 20. For example, if there are four cylinders 20, four first exhaust cams 28a and four second exhaust cams 28b are provided. The first exhaust valve 32a oscillates along the cam profile of the first exhaust cam 28a, resulting in a lift curve as shown in FIG. 5. The second exhaust valve 32b oscillates along the cam profile of the second exhaust cam 28b, resulting in a lift curve as shown in FIG. 5.

[0027] As shown in Fig. 5, the first exhaust cam 28a and the second exhaust cam 28b are split cams with different phases and cam profiles. Specifically, the second exhaust cam 28b is retarded more than the first exhaust cam 28a. As a result, the second exhaust valve 32b opens later than the first exhaust valve 32a. Furthermore, the cam profile of the first exhaust cam 28a is formed so that the first exhaust valve 32a is at a maximum lift position P2, while the cam profile of the second exhaust cam 28b is formed so that the second exhaust valve 32b is at a maximum lift position P3 that is lower than the maximum lift position P2. In other words, the maximum lift position P3 of the second exhaust valve 32b is lower than the maximum lift position P2 of the first exhaust valve 32a. On the other hand, the cam profile of the first exhaust cam 28a and the cam profile of the second exhaust cam 28b are formed so that the first exhaust valve 32a changes from the maximum lift position P2 to the closed valve position P0 at the same time as the second exhaust valve 32b changes from the maximum lift position P3 to the closed valve position P0. That is, the first exhaust valve 32a and the second exhaust valve 32b close simultaneously. The open period of the second exhaust valve 32b is shorter than the open period of the first exhaust valve 32a. The cam profile of the second exhaust cam 28b may be either linearly symmetrical or asymmetrical in the ascending cam profile from the closed valve position P0 to the maximum lift position P3 and the descending cam profile from the maximum lift position P3 to the closed valve position P0. An example of an asymmetric cam profile of the second exhaust cam 28b is one in which the descending period is shorter than the ascending period. In this case, the cam profile of the second exhaust cam 28b may be formed so that the lift curve rises gradually from the valve closing position P0.

[0028] Furthermore, the maximum lift position P1 of the first intake valve 30a and the second intake valve 30b is higher than the maximum lift position P2 of the first exhaust valve 32a and the maximum lift position P3 of the second exhaust valve 32b. In other words, the lift amounts of the first intake valve 30a and the second intake valve 30b are higher than the lift amounts of the first exhaust valve 32a and the second exhaust valve 32b, respectively.

[0029] 2, in this embodiment, the internal combustion engine 1 further includes an intake variable valve timing device 26c that can change the phase of the intake camshaft 26, and an exhaust variable valve timing device 28c that can change the phase of the exhaust camshaft 28. The intake variable valve timing device 26c and the exhaust variable valve timing device 28c can adjust the valve overlap amount between the first intake valve 30a and the second intake valve 30b and the first exhaust valve 32a and the second exhaust valve 32b by changing the phase of each camshaft.

[0030] As shown in FIG. 3 , the combustion chamber 20a is formed asymmetrically with respect to a symmetry axis O1, which passes through the center O and is a lateral center line O1 of the combustion chamber 20a. That is, the shape of the combustion chamber 20a is asymmetric between the side where the first exhaust valve 32a is disposed and the side where the second exhaust valve 32b is disposed, with the center line O1 of the combustion chamber 20a sandwiched between them. More specifically, the combustion chamber 20a is formed such that a right wall 42 between the first intake valve 30a and the first exhaust valve 32a bulges out in a direction away from the center O of the combustion chamber 20a more than a left wall 44 between the second intake valve 30b and the second exhaust valve 32b. In this embodiment, the right wall 42 is formed bulging in an arc shape to fit the cylindrical shape of the cylinder 20.

[0031] In the combustion chamber 20a, an intake shroud 45 is formed spanning the first intake valve 30a and the second intake valve 30b, and the intake shroud 45 has an intake shroud protrusion (part of the intake shroud 45) 46 between the first intake valve 30a and the second intake valve 30b that protrudes toward the center O of the combustion chamber 20a. Furthermore, as shown in FIGS. 3, 6, and 7, a first shroud wall 48 is formed on the intake shroud 45 around the first intake valve 30a side, and is formed so that the height of the first shroud wall 48 on the AA cross section is higher (see reference line Y in FIG. 7(a)) when the height of the first shroud wall 48 on the aa cross section is taken as the reference (see reference line X in FIG. 6(a)). 3, 8, and 9, a second shroud wall 50 is formed on the intake shroud 45 around the second intake valve 30b side, and the height of the second shroud wall 50 at cross section bb is the same as the height of the first shroud wall 48 at cross section aa (see reference line X in FIG. 9(a)), and the height of the second shroud wall 50 at cross section BB is the same as the height of the second shroud wall 50 at cross section AA (see reference line D in FIG. 8(a)). However, the gap between the second intake valve 30b and the second shroud wall 50 is formed to be larger than the gap between the first intake valve 30a and the first shroud wall 48 (see gap D in FIG. 8(a)). In this embodiment, an exhaust shroud 51 is formed across the first exhaust valve 32a and the second exhaust valve 32b, and an exhaust shroud protrusion 52 is also formed on the exhaust shroud 51 between the first exhaust valve 32a and the second exhaust valve 32b.

[0032] Next, with reference to FIGS. 5 to 12, the formation of the air-fuel mixture in the combustion chamber and the exhaust state in the internal combustion engine 1 configured as above will be described.

[0033] As shown in Figure 5, the first intake valve 30a opens first during the intake stroke. In the early stages of opening of the first intake valve 30a, the intake shroud 45 prevents the first shroud wall 48 from blocking the flow from the FS side of the first intake valve 30a. As shown by the arrows in Figures 6(b) and 7(b), the air-fuel mixture flows into the combustion chamber 20a biased toward the BS side of the first intake valve 30a, forming a tumble flow. Furthermore, toward the end of the opening of the first intake valve 30a (when the valve approaches the maximum lift position P1 in Figure 5), the air-fuel mixture also flows into the combustion chamber 20a from the FS side of the valve, as shown by the arrows in Figures 6(c) and 7(c). The height of the first shroud wall 48 around the intake shroud protrusion 46 is set higher than the first shroud wall 48 other than the intake shroud protrusion 46. This obstructs the flow of the incoming mixture, causing a deviation in the mixture flow on the left and right sides. As shown by the arrows in Figure 10(a), this strengthens the flow along the first shroud wall 48 and the right wall 42, and in addition to the tumble flow, a swirl flow is also formed inside the combustion chamber 20a along the circumferential direction of the cylinder 20.

[0034] Next, the second intake valve 30b opens as shown in FIG. 5. In the initial stage of opening of the second intake valve 30b, the intake shroud 45 prevents the airflow from the front-side end of the second intake valve 30b by the second shroud wall 50. As shown by the arrows in FIGS. 8(b) and 9(b), the air-fuel mixture flows into the combustion chamber 20a biased toward the back-side end of the second intake valve 30b, forming a tumble flow. Furthermore, in the final stage of opening of the second intake valve 30b, the air-fuel mixture also flows into the combustion chamber 20a from the front-side end of the valve as shown by the arrows in FIGS. 8(c) and 9(c). At this time, the height of the second shroud wall 50 on the second intake valve 30b side is the same as that on the first intake valve side due to the intake shroud protrusion 46. However, by providing a gap D between the second intake valve 30b and the second shroud wall 50, the air-fuel mixture flows into the combustion chamber 20a without forming a swirl flow. Furthermore, by narrowing the gap D between the second intake valve 30b and the second shroud wall 50 as the height of the shroud wall decreases (i.e., the gap D becomes increasingly distant from the second intake valve 30b as it approaches the center O; see also Figure 3), a uniform flow rate can be allowed to flow into the combustion chamber 20a from the entire circumferential surface of the second intake valve 30b. This allows a larger amount of mixture to flow into the combustion chamber 20a than with the first intake valve 30a, and the swirl flow formed when the first intake valve 30a opens is maintained without being obstructed even when the second intake valve 30b opens. Furthermore, because the first intake valve 30a and the second intake valve 30b close simultaneously, reverse flow of the swirl flow from the second intake valve 30b is also suppressed.

[0035] As shown by the arrows in Figure 10(b), the swirl flow of the air-fuel mixture formed in the combustion chamber 20a becomes biased toward the first intake valve 30a during the compression stroke after the first and second intake valves 30a and 30b are closed. The air-fuel mixture with this swirl flow is ignited by the spark plug 38 and burns. As a result, the flame generated after the air-fuel mixture is ignited propagates vigorously throughout the combustion chamber 20a over the unburned air-fuel mixture, which has become more turbulent due to the collapse of the tumble flow before ignition. The maintained biased swirl flow also helps to uniformize the uneven propagation of the flame, improving the degree of constant volume and preventing knocking caused by the end gas of the unburned air-fuel mixture.

[0036] As shown in Figure 5, during the exhaust stroke, the first exhaust valve 32a opens first, followed by the second exhaust valve 32b. The first exhaust valve 32a is located diagonally opposite the first intake valve 30a across the center O. Therefore, as shown by the arrow in Figure 11(a), the swirl flow biased toward the first intake valve 30a is maintained for a longer period of time than when the first exhaust valve 32a opens before the second exhaust valve 32b. This shortens the flame propagation during the expansion stroke, improves the degree of constant volume, and improves combustion efficiency.

[0037] Furthermore, with this internal combustion engine 1, the second exhaust valve 32b opens later than the first exhaust valve 32a, resulting in less exhaust interference than when the first exhaust valve 32a and the second exhaust valve 32b open simultaneously. Specifically, as shown by the arrows in FIG. 11(b), the first exhaust valve 32a opens first, allowing exhaust gas to flow into the second exhaust port 24c. Then, as shown by the arrows in FIG. 11(c), the second exhaust valve 32b opens later than the first exhaust valve 32a, allowing exhaust gas to flow into the first exhaust port 24b. In this way, the exhaust pressure in the second exhaust port 24c increases first, and the exhaust pressure generated in the first exhaust port 24b increases later. This reduces exhaust interference at the junction of the first exhaust port 24b and the second exhaust port 24c, and also reduces the maximum value of the exhaust pressure.

[0038] Furthermore, the opening period of the second exhaust valve 32b is shorter than the opening period of the first exhaust valve 32a, and the lift amount of the second exhaust valve 32b is also smaller than the lift amount of the first exhaust valve 32a. This further reduces exhaust interference. As a result, the amount of exhaust gas (internal EGR) that flows back into the combustion chamber 20a due to exhaust interference is reduced. As a result, not only is knocking suppressed, but scavenging of the combustion chamber 20a is also facilitated.

[0039] 12(a) to 12(d) are graphs showing an example of reduced exhaust interference. The solid line, representing the exhaust split EC alignment, represents the graph when the first exhaust valve 32a and second exhaust valve 32b of this embodiment are used, while the dashed line, representing the standard cam, represents the graph of a conventional cam profile. As described above, in this embodiment, the lift amounts of the first exhaust valve 32a and second exhaust valve 32b are smaller than the lift amounts of the first intake valve 30a and second intake valve 30b. This allows the maximum exhaust pressure M2 to be lower than the maximum exhaust pressure M1 for the standard cam, as shown in FIG. 12(b), compared to when the lift amounts of the intake valve 30 and the exhaust valve 32 are the same. As a result, the internal combustion engine 1 can further reduce exhaust interference, making it easier to suppress exhaust backflow, especially in high-load ranges.

[0040] Furthermore, the first exhaust valve 32a and the second exhaust valve 32b close simultaneously during the exhaust stroke. That is, as shown in FIG. 12(b), by making the exhaust pressure rise in the first exhaust port 24b later than in the second exhaust port 24c, the timing N2 at which the exhaust pressure rise reaches its maximum value can be made later than the timing N1 at which the exhaust pressure reaches its maximum value for the normal cam. This allows the internal combustion engine 1 to shift the timing at which the exhaust pressure rises and reaches its maximum value during the valve overlap period (VOL period in FIG. 12) in another cylinder 20. This allows the internal combustion engine 1 to suppress backflow from the exhaust and further suppress exhaust interference with another cylinder 20. As a result, as shown in FIG. 12(d), it is possible to suppress backflow of exhaust gas due to exhaust interference in the exhaust manifold 25 in the cylinder 20 as well, shortening the exhaust valve opening period and avoiding deterioration of exhaust efficiency due to a reduced lift amount.

[0041] Furthermore, by closing the first exhaust valve 32a and the second exhaust valve 32b simultaneously during the exhaust stroke, the valve overlap between the first intake valve 30a and the second exhaust valve 32b is reduced compared to when the second exhaust valve 32b is closed later than the first exhaust valve 32a. This reduces the amount of immediately combustible gas (internal EGR gas) that flows back from the exhaust to the combustion chamber 20a at high speeds and high loads where the exhaust pressure is high, making it possible to avoid knocking while maintaining a swirl flow. Furthermore, at low speeds and high loads where the exhaust pressure is relatively low, it is possible to prevent the air-fuel mixture flowing into the combustion chamber 20a when the first intake valve 30a opens from blowing through the second exhaust valve 32b. As a result, it becomes easier to generate a swirl flow and improves charging efficiency.

[0042] In this type of internal combustion engine 1, a larger amount of air is taken into the cylinder 20 and a larger amount of exhaust gas is discharged after combustion in a high-load operating range than in a low-load operating range. The electric vehicle C frequently uses this high-load operating range. Therefore, in this high-load operating range, by reducing exhaust interference and suppressing exhaust backflow, intake and exhaust performance can be made equal to or better than that of a normal cam, while engine performance can be improved by improving combustion efficiency through enhanced swirl flow. For this reason, this type of internal combustion engine 1 is particularly suitable for the electric vehicle C.

[0043] As described above, according to the present disclosure, an internal combustion engine with good combustion efficiency can be provided.

[0044] <Other embodiments> Although the embodiments of the present disclosure have been described above, the present disclosure is not limited to the above embodiments, and various modifications are possible within the scope of the gist of the invention. In particular, the multiple modifications described in this specification can be arbitrarily combined as necessary.

[0045] (a) In the above embodiment, an example of an electric vehicle C having modes such as an EV mode, a series mode, a parallel mode, and a charging mode has been described, but the present disclosure is not limited to this. The electric vehicle C is not limited to these modes, and may have any mode in which the internal combustion engine 1 is used.

[0046] (b) In the above embodiment, a gasoline engine has been described as an example, but the present disclosure is not limited to this. The internal combustion engine 1 may be a self-ignition engine such as a diesel engine.

[0047] (c) In the above embodiment, the intake valve 30 and the exhaust valve 32 are driven by a split-type direct-acting cam, but the present disclosure is not limited to this. The intake valve 30 and the exhaust valve 32 may be, for example, electric valves. [Explanation of symbols]

[0048] 1: internal combustion engine, 20a: combustion chamber, 22: Intake port, 24: Exhaust port, 30: intake valve, 30a: first intake valve, 30b: second intake valve 32: Exhaust valve, 32a: First exhaust valve, 32b: Second exhaust valve 42: Right wall, 44: Left wall 46: Intake shroud protrusion 48: First shroud wall, 50: Second shroud wall O: Center, O1: Center line P0: Valve closed position P1, P2, P3: Maximum lift position

Claims

1. A combustion chamber; an intake passage connected to the combustion chamber; an exhaust passage connected to the combustion chamber; a first intake valve that opens and closes between the intake passage and the combustion chamber; a second intake valve disposed adjacent to the first intake valve and configured to open and close the intake passage and the combustion chamber; a first exhaust valve that opens and closes between the exhaust passage and the combustion chamber; a second exhaust valve disposed adjacent to the first exhaust valve and configured to open and close the exhaust passage and the combustion chamber; Equipped with the second intake valve opens later than the first intake valve; the second exhaust valve opens later than the first exhaust valve; the lift amounts of the first exhaust valve and the second exhaust valve are smaller than the lift amounts of the first intake valve and the second intake valve, The lift amounts of the first intake valve and the second intake valve are the same. Internal combustion engine.

2. The first intake valve and the second intake valve simultaneously close the intake passage.

2. The internal combustion engine according to claim 1.

3. the first exhaust valve and the second exhaust valve simultaneously close the exhaust passage; 2. The internal combustion engine according to claim 1.

4. The lift amount of the second exhaust valve is smaller than the lift amount of the first exhaust valve.

2. The internal combustion engine according to claim 1.

5. The first intake valve and the first exhaust valve are disposed at diagonal positions across the center of the combustion chamber.

2. The internal combustion engine according to claim 1.

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