Internal combustion engine

Abstract

This internal combustion engine comprises: a combustion chamber; an exhaust passage connected to the combustion chamber; a first exhaust valve which opens / closes the interface between the exhaust passage and the combustion chamber; and a second exhaust valve disposed adjacent to the first exhaust valve and opens / closes the exhaust passage. The valve opening period of the second exhaust valve is shorter than the valve opening period of the first exhaust valve. The first exhaust valve and the second exhaust valve close simultaneously.

Description

Technical Field

[0001] The present disclosure relates to an internal combustion engine.

Background Art

[0002] Conventionally, an internal combustion engine having a plurality of intake valves and a plurality of exhaust valves in one cylinder has been known (see, for example, Patent Document 1). In the internal combustion engine of Patent Document 1, one cylinder has two exhaust valves, and the opening / closing timing of one of the two exhaust valves is retarded with respect to the opening / closing timing of the other exhaust valve. In the internal combustion engine of Patent Document 1, by making the opening / closing timings of the exhaust valves different 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

Summary of the Invention

Problems to be Solved by the Invention

[0004] In the internal combustion engine of Patent Document 1, in order to improve the deterioration of combustibility due to a decrease in intake charge efficiency in the low load region, such an exhaust-side swirl flow is utilized. Specifically, in the internal combustion engine of Patent Document 1, a swirl control valve is used to generate a swirl flow on the intake side as well, mix the swirl flow on the intake side with the swirl flow on the exhaust side, and promote the flow of the air-fuel mixture in the cylinder.

[0005] However, in the high load region, such an exhaust-side swirl flow and an intake-side swirl flow collide with each other, resulting in insufficient scavenging in the cylinder, which may have an adverse effect on the combustion in 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 according to this disclosure comprises a combustion chamber, an exhaust passage connected to the combustion chamber, a first exhaust valve that opens and closes the space between the exhaust passage and the combustion chamber, and a second exhaust valve arranged adjacent to the first exhaust valve that opens and closes the exhaust passage, wherein the opening period of the second exhaust valve is shorter than the opening period of the first exhaust valve, and the first exhaust valve and the second exhaust valve close simultaneously.

[0008] In this internal combustion engine, the peak of the exhaust pressure increase in the exhaust passage opened and closed by the first exhaust valve is offset from the peak of the exhaust pressure increase in the exhaust passage opened and closed by the second exhaust valve. This reduces exhaust interference within the exhaust passage and also suppresses the maximum value of the exhaust pressure. As a result, exhaust backflow is suppressed and combustion efficiency is improved.

[0009] Furthermore, by closing the first and second exhaust valves simultaneously, interference between exhaust gases discharged from different combustion chambers can be suppressed. This further reduces exhaust backflow and improves combustion efficiency. [Effects of the Invention]

[0010] According to this disclosure, it is possible to provide an internal combustion engine with good combustion efficiency. [Brief explanation of the drawing]

[0011] [Figure 1] A system diagram of a vehicle equipped with an internal combustion engine according to one embodiment of the present disclosure. [Figure 2] A cross-sectional view of an internal combustion engine according to one embodiment of the present disclosure. [Figure 3] A view from below of the combustion chamber of an internal combustion engine according to one embodiment of the present disclosure. [Figure 4] A diagram showing a camshaft of an internal combustion engine according to one embodiment of the present disclosure. [Figure 5] A diagram showing the lift curves of the intake valve and exhaust valve of an internal combustion engine according to one embodiment of the present disclosure. [Figure 6] Figure 3 shows a cross-section of aa. [Figure 7] Figure 3 shows the AA section. [Figure 8] Figure 3 shows a cross-section of BB. [Figure 9] This figure shows the cross-section between b and b in Figure 3. [Figure 10] A diagram showing the flow of the air-fuel mixture in the combustion chamber during the intake, compression, and expansion strokes of an internal combustion engine according to one embodiment of the present disclosure. [Figure 11] A diagram showing the exhaust flow in the combustion chamber during the exhaust stroke of an internal combustion engine according to one embodiment of the present disclosure. [Figure 12] A graph showing an example of reduced exhaust interference. [Modes for carrying out the invention]

[0012] Hereinafter, one 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 top side of the electric vehicle C, and DS indicates the bottom 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 includes a 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 a plurality of gears and a clutch 8a. The internal combustion engine 1 is connected to the generator 4 and the axle 10 via the transaxle 8. When the clutch 8a is in the open state, the power transmission between the internal combustion engine 1 and the axle 10 is interrupted, and when the clutch 8a is in the connected state, the power of the internal combustion engine 1 is transmitted to the axle 10. The motor 2 is connected to the axle 10 via the transaxle 8. In addition, the electric vehicle C may have a vehicle control device 12, an engine control device 14 that controls the internal combustion engine 1, an accelerator pedal 16 operated by the 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 modes such as an EV mode, a series mode, a parallel mode, and a charging mode. In the EV mode, the electric vehicle C drives the motor 2 with the power from the drive battery 6. In the series mode, the electric vehicle C drives the generator 4 with the internal combustion engine 1 and drives the motor 2 using the power generated by the generator 4. In the parallel mode, the electric vehicle C connects the clutch 8a and drives the axle 10 using the power of the internal combustion engine 1. In the charging mode, the electric vehicle C drives the generator 4 with the internal combustion engine 1 and stores the power generated by the generator 4 in the drive battery 6. According to the depression state of the accelerator pedal 16 and the operation state of the charging button, the vehicle control device 12 switches the modes, 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] When such an electric vehicle C travels in a low load region, the motor 2 is used. The electric vehicle C uses the internal combustion engine 1 in the series mode, the parallel mode, and the charging mode. In these series mode, parallel mode, and charging mode, the internal combustion engine 1 is mainly operated in a high load operation region. Therefore, the internal combustion engine 1 used in such an electric vehicle C is required to have good combustion efficiency in the high load operation region.

[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 the present embodiment is a horizontally opposed type in which the cylinders 20 of the internal combustion engine 1 are arranged side by side along the direction of the axle 10 of the wheel 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 the present embodiment, four cylinders 20 are arranged side by side. That is, the internal combustion engine 1 of the present embodiment is a four - cylinder in - line engine. As shown in FIG. 2, in each cylinder 20, a piston 34 connected from the crankshaft 36 via a connecting rod is slidably accommodated. 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 Figure 2, in this embodiment, the combustion chamber 20a is a pent-roof type combustion chamber 20a with ridges formed in the left-right direction of the internal combustion engine 1. The intake port 22 is formed in the cylinder head 1a and is connected to the intake side (front side in Figure 2) slope of the combustion chamber 20a. The exhaust port 24 is formed in the cylinder head 1a and is connected to the exhaust side (rear side in Figure 2) slope of the combustion chamber 20a. A spark plug 38 is positioned at the center O of the combustion chamber 20a (see Figure 3). In other words, the internal combustion engine 1 in this embodiment is a gasoline engine.

[0020] As shown in Figures 1 and 2, the 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 Figure 3, the intake port 22 is branched to the left and right by the intake port wall 22a, and has a first intake port 22b located on the right side and a second intake port 22c located on the left side. In this embodiment, the fuel injector 40 is located in front of the branching of the intake port 22. Fuel is supplied to the fuel injector 40 from the fuel tank 54 (see Figure 1), and the fuel injected from the fuel injector 40 is mixed with the air passing through the intake port 22 and supplied to the combustion chamber 20a.

[0021] As shown in Figures 1 and 2, the exhaust ports 24 are provided in each of the four cylinders 20 and are passages for discharging exhaust gas generated after the combustion of the fuel-air mixture in the combustion chamber 20a. As shown in Figure 3, the exhaust ports 24 are branched to the left and right by the exhaust port wall 24a, and have 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 the exhaust manifold 25. The exhaust manifold 25 collects and discharges the exhaust gas flowing through the exhaust ports 24 located in each cylinder 20.

[0022] The intake valve 30 opens and closes the space between the intake port 22 and the combustion chamber 20a. As shown in Figure 3, the intake valve 30 has a first intake valve 30a located on the right side and a second intake valve 30b located adjacent to the first intake valve 30a and on the left side. The first intake valve 30a opens and closes the space between the first intake port 22b and the combustion chamber 20a. The second intake valve 30b opens and closes the space 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 adjacent to the first exhaust valve 32a and on the right side. 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 Figure 2, in this embodiment, the intake valve 30 is a direct-acting valve driven by the cam of the intake camshaft 26 pushing out the intake valve 30. As shown in Figure 4(a), the intake camshaft 26 of 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 such first intake cams 26a and second intake cams 26b arranged in one cylinder 20. For example, if there are four cylinders 20, four first intake cams 26a and four second intake cams 26b are arranged. The first intake valve 30a swings along the cam profile of the first intake cam 26a, resulting in a lift curve as shown in Figure 5. The second intake valve 30b swings along the cam profile of the second intake cam 26b, resulting in a lift curve as shown in Figure 5.

[0025] As shown in Figure 5, the first intake cam 26a and the second intake cam 26b are split-type cams with different phases and cam profiles. Specifically, the second intake cam 26b has a retarded angle compared to the first intake cam 26a. As a result, the second intake valve 30b opens later than the first intake valve 30a. Furthermore, the cam profiles of the first intake cam 26a and the second intake cam 26b are formed so that the timing at which the first intake valve 30a moves from its maximum lift position P1 to its closed position P0 is simultaneous with the timing at which the second intake valve 30b moves from its maximum lift position P1 to its closed position P0. In other words, the first intake valve 30a and the second intake valve 30b close simultaneously. Also, the opening period of the second intake valve 30b is shorter than the opening period of the first intake valve 30a. The cam profile of the second intake cam 26b may be symmetrical or asymmetrical in its upward cam profile from the closed valve position P0 to the maximum lift position P1, and in its downward cam profile from the maximum lift position P1 to the closed valve position P0. An example of an asymmetrical cam profile for the second intake cam 26b is one in which the upward period is shorter than the downward period, or vice versa.

[0026] As shown in Figure 2, in this embodiment, the exhaust valve 32 is a direct-acting valve driven by the cam of the exhaust camshaft 28 pushing out the exhaust valve 32. As shown in Figure 4(b), the exhaust camshaft 28 of 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 such first exhaust cams 28a and second exhaust cams 28b arranged in one cylinder 20. For example, if there are four cylinders 20, four first exhaust cams 28a and four second exhaust cams 28b are arranged. The first exhaust valve 32a oscillates along the cam profile of the first exhaust cam 28a, resulting in a lift curve as shown in Figure 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 Figure 5.

[0027] As shown in Figure 5, the first exhaust cam 28a and the second exhaust cam 28b are split-type cams with different phases and cam profiles. Specifically, the second exhaust cam 28b is retarded in angle compared to the first exhaust cam 28a. This causes the second exhaust valve 32b to open 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 reaches its maximum lift position P2, while the cam profile of the second exhaust cam 28b is formed so that the second exhaust valve 32b reaches 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 profiles of the first exhaust cam 28a and the second exhaust cam 28b are formed such that the timing at which the first exhaust valve 32a moves from the maximum lift position P2 to the closed position P0 and the timing at which the second exhaust valve 32b moves from the maximum lift position P3 to the closed position P0 are simultaneous. That is, the first exhaust valve 32a and the second exhaust valve 32b close simultaneously. Also, the opening period of the second exhaust valve 32b is shorter than the opening period of the first exhaust valve 32a. The cam profile of the second exhaust cam 28b may be symmetrical or asymmetrical in its upward profile from the closed position P0 to the maximum lift position P3 and its downward profile from the maximum lift position P3 to the closed position P0. For example, an asymmetrical cam profile of the second exhaust cam 28b may be made in which the downward period is shorter than the upward period. In this case, the cam profile of the second exhaust cam 28b may be formed such that the rise of the lift curve from the closed valve position P0 is gradual.

[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.

[0029] As shown in Figure 2, in this embodiment, the internal combustion engine 1 further includes an intake variable valve timing device 26c capable of changing the phase of the intake camshaft 26, and an exhaust variable valve timing device 28c capable of changing 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 Figure 3, the combustion chamber 20a is formed with an asymmetrical shape on the left and right sides, with the center line O1 passing through the center O and the left and right center lines O1 of the combustion chamber 20a being the axis of symmetry. That is, the combustion chamber 20a is asymmetrical in shape on the side where the first exhaust valve 32a is located and the side where the second exhaust valve 32b is located, with the center line O1 of the combustion chamber 20a in between. More specifically, the combustion chamber 20a is formed with the first intake valve 30a and Second exhaust valve 32b The right wall 42 between the second intake valve 30b and First exhaust valve 32a The right wall 42 is formed to bulge outwards from the center O of the combustion chamber 20a, compared to the left wall 44 between it and the cylinder 20. In this embodiment, the right wall 42 is formed to bulge outwards in an arc shape along 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. The intake shroud 45 has an intake shroud protrusion (part of the intake shroud 45) 46 that projects toward the center O of the combustion chamber 20a between the first intake valve 30a and the second intake valve 30b. Furthermore, as shown in Figures 3, 6, and 7, a first shroud wall 48 is formed on the intake shroud 45 around the first intake valve 30a side. The first shroud wall 48 in the AA section is formed to be higher than the height of the first shroud wall 48 in the aa section (see reference line X in Figure 6(a)) (see reference line Y in Figure 7(a)). On the other hand, as shown in Figures 3, 8, and 9, a second shroud wall 50 is formed on the intake shroud 45 around the second intake valve 30b. The height of the second shroud wall 50 in the bb cross section is the same as the height of the first shroud wall 48 in the aa cross section (see reference line X in Figure 9(a)), and the height of the second shroud wall 50 in the BB cross section is the same as the height of the second shroud wall 50 in the AA cross section (see reference line in Figure 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 Figure 8(a)). In this embodiment, the exhaust shroud 51 is formed spanning the first exhaust valve 32a and the second exhaust valve 32b, and the exhaust shroud 51 also has an exhaust shroud protrusion 52 between the first exhaust valve 32a and the second exhaust valve 32b.

[0032] Next, using Figures 5 to 12, the state of the mixture formation in the combustion chamber and the exhaust state in the internal combustion engine 1 configured as described above will be explained.

[0033] As shown in Figure 5, the first intake valve 30a opens first during the intake stroke. In the initial stages of opening the first intake valve 30a, the intake shroud 45 obstructs the flow from the FS side of the first intake valve 30a by the first shroud wall 48, causing the air-fuel mixture to flow into the combustion chamber 20a biased towards the BS side of the first intake valve 30a, as shown by the arrows in Figures 6(b) and 7(b), forming a tumble flow. Furthermore, towards the end of opening the first intake valve 30a (around the time when it reaches 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 that of the first shroud wall 48 other than the intake shroud protrusion 46. This obstructs the flow of the incoming air-fuel mixture, causing a deviation in the flow of the air-fuel mixture 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 along the circumferential direction of the cylinder 20 is also formed within the combustion chamber 20a.

[0034] Next, as shown in Figure 5, the second intake valve 30b opens. In the initial stage of opening the second intake valve 30b, the intake shroud 45 obstructs the flow from the FS side of the second intake valve 30b by the second shroud wall 50, causing the air-fuel mixture to flow into the combustion chamber 20a biased towards the BS side of the second intake valve 30b, as shown by the arrows in Figures 8(b) and 9(b), forming a tumble flow. Furthermore, in the final stage of opening the second intake valve 30b, as shown by the arrows in Figures 8(c) and 9(c), the air-fuel mixture also flows into the combustion chamber 20a from the FS side of the valve. At this time, the height of the second shroud wall 50 is the same as that of the first intake valve side due to the intake shroud protrusion 46 on the second intake valve 30b side, but a gap D is provided between the second intake valve 30b and the second shroud wall 50, allowing air-fuel mixture that does not form a swirl flow to flow into the combustion chamber 20a. 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 shape moves away from the second intake valve 30b as it approaches the center O, see also Figure 3), a uniform flow rate can be made to flow into the combustion chamber 20a from the entire circumference of the second intake valve 30b. As a result, the amount of air-fuel mixture flowing into the combustion chamber 20a can be increased compared to the first intake valve 30a, and the swirl flow formed by the opening of the first intake valve 30a is maintained without being obstructed even when the second intake valve 30b opens. In addition, since the first intake valve 30a and the second intake valve 30b close simultaneously, backflow of the swirl flow from the second intake valve 30b can also be suppressed.

[0035] As shown by the arrow in Figure 10(b), the swirl flow of the air-fuel mixture formed in the combustion chamber 20a becomes a flow that is centered towards the first intake valve 30a during the compression stroke after the first intake valve 30a and the second intake valve 30b are closed. This air-fuel mixture with such a swirl flow is ignited and combusted by the spark plug 38. As a result, the flame generated after ignition propagates vigorously throughout the combustion chamber 20a over the unburned mixture, which has become more turbulent due to the collapse of the tumble flow before ignition. At the same time, the maintained centered swirl flow helps to equalize the uneven flame propagation, improving volumetricity and preventing knocking caused by end gases from the unburned 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 towards the first intake valve 30a is maintained for a longer period compared to the case where the first exhaust valve 32a opens before the second exhaust valve 32b. This shortens the flame propagation during the expansion stroke, improves isovolume, and increases combustion efficiency.

[0037] Furthermore, with this internal combustion engine 1, since the second exhaust valve 32b opens later than the first exhaust valve 32a, there is less exhaust interference than when the first exhaust valve 32a and the second exhaust valve 32b open simultaneously. Specifically, as shown by the arrow in Figure 11(b), first the first exhaust valve 32a opens, and exhaust flows into the second exhaust port 24c. Then, as shown by the arrow in Figure 11(c), the second exhaust valve 32b opens later than the first exhaust valve 32a, and exhaust flows 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 with a delay. As a result, exhaust interference at the junction of the first exhaust port 24b and the second exhaust port 24c is reduced, and the maximum value of the exhaust pressure is also reduced.

[0038] Furthermore, the opening period of the second exhaust valve 32b is shorter than that of the first exhaust valve 32a, and the lift amount of the second exhaust valve 32b is also shorter than that 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. This not only suppresses knocking but also facilitates scavenging within the combustion chamber 20a.

[0039] Figures 12(a) to 12(d) are graphs showing an example of reduced exhaust interference. The solid line, with the exhaust split EC aligned, represents the graph when the first exhaust valve 32a and second exhaust valve 32b are used according to this embodiment, while the dashed line, representing a normal 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 makes it possible to lower the maximum exhaust pressure M2 to a lower value than the maximum value M1 of a normal cam, as shown in Figure 12(b), compared to making the lift amounts of the intake valve 30 and exhaust valve 32 the same. As a result, the internal combustion engine 1 can further reduce exhaust interference, and in particular, it is easier to suppress exhaust backflow in the high-load region.

[0040] Furthermore, during the exhaust stroke, the first exhaust valve 32a and the second exhaust valve 32b close simultaneously. That is, as shown in Figure 12(b), by delaying the increase in exhaust pressure at the first exhaust port 24b compared to the second exhaust port 24c, it becomes possible to delay the timing N2 of reaching the maximum value of the exhaust pressure increase compared to the timing N1 of reaching the maximum value of a normal cam. As a result, the internal combustion engine 1 can stagger the timing of the increase in exhaust pressure and its maximum value during the valve overlap period (VOL period in Figure 12) in another cylinder 20. This allows the internal combustion engine 1 to suppress backflow from the exhaust and further suppress exhaust interference to another cylinder 20. As a result, as shown in Figure 12(d), backflow of exhaust due to exhaust interference within the exhaust manifold 25 is suppressed in cylinder 20 as well, making it possible to shorten the opening period of the exhaust valve and avoid deterioration of exhaust efficiency due to a smaller lift amount.

[0041] Furthermore, when the first exhaust valve 32a and the second exhaust valve 32b close 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 closes later than the first exhaust valve 32a. As a result, at high rotational speeds and high loads with high exhaust pressure, the amount of immediately combustible gas (internal EGR gas) flowing back from the exhaust into the combustion chamber 20a can be reduced, making it possible to avoid knocking while maintaining swirl flow. Also, at low rotational speeds and high loads with relatively low exhaust pressure, it is possible to suppress the air-fuel mixture flowing into the combustion chamber 20a when the first intake valve 30a opens from blowing out through the second exhaust valve 32b. As a result, it becomes easier to generate swirl flow, and the charging efficiency is also improved.

[0042] In such an internal combustion engine 1, the amount of air drawn into the cylinder 20 is greater in the high-load operating range than in the low-load operating range, and the amount of exhaust gas discharged after combustion is also greater. Electric vehicles C frequently utilize such high-load operating ranges. Therefore, in such high-load operating ranges, 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 further engine performance can be improved by enhancing combustion efficiency through swirl flow enhancement. For this reason, such an internal combustion engine 1 is more suitable for electric vehicles C.

[0043] As explained above, this disclosure provides an internal combustion engine with good combustion efficiency.

[0044] <Other Embodiments> Although embodiments of the present disclosure have been described above, the present disclosure is not limited to the embodiments described above, and various modifications are possible without departing from the spirit of the invention. In particular, the various modifications described herein can be combined as needed.

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

[0046] (b) In the embodiments described above, a gasoline engine was used as an example, but the disclosure is not limited thereto. The internal combustion engine 1 may be an auto-ignition type engine such as a diesel engine.

[0047] (c) In the above embodiment, an example was described in which the intake valve 30 and exhaust valve 32 are driven by a split-type direct-acting cam, but the disclosure is not limited thereto. The intake valve 30 and exhaust valve 32 may be, for example, electrically operated valves.

[0048] (d) In the above embodiment, an example was described in which a swirl flow is generated by opening the second intake valve 30b later than the first intake valve 30a, but the disclosure is not limited thereto. The swirl flow may be generated using, for example, a swirl control valve provided in the second intake port 22c. [Explanation of symbols]

[0049] 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. The combustion chamber and An exhaust passage connected to the combustion chamber, A first exhaust valve that opens and closes the exhaust passage and the combustion chamber, A second exhaust valve is positioned adjacent to the first exhaust valve and opens and closes the exhaust passage, An intake passage connected to the combustion chamber, A first intake valve that opens and closes the intake passage and the combustion chamber, A second intake valve is positioned adjacent to the first intake valve and opens and closes the space between the intake passage and the combustion chamber, Equipped with, The opening period of the second exhaust valve is shorter than the opening period of the first exhaust valve. The first exhaust valve and the second exhaust valve close simultaneously. The combustion chamber is formed spanning between the first intake valve and the second intake valve, and has a shroud that protrudes from the wall of the combustion chamber toward the center of the combustion chamber. The shroud is formed such that the wall on the first intake valve side extends along the first intake valve, and the wall on the second intake valve side moves away from the second intake valve as it approaches the center of the combustion chamber. Internal combustion engine.

2. On either side of the centerline of the combustion chamber, the first exhaust valve is positioned, The shape of the combustion chamber is asymmetrical on the side where the second exhaust valve is located. The internal combustion engine according to claim 1.

3. The combustion chamber is formed such that the wall between the first intake valve and the second exhaust valve bulges outward from the center of the combustion chamber more than the wall between the second intake valve and the first exhaust valve. The internal combustion engine according to claim 1.

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