JP2019214995A - Internal combustion engine - Google Patents

Abstract

To increase a thermal efficiency to a maximum while securing a sufficient out at an output point of an internal combustion engine.SOLUTION: An internal combustion engine 100 with a mechanical compression ratio of 15 and up includes: a piston reciprocating in a cylinder; an intake valve 9 with a working angle of 180 degrees or higher; an exhaust valve 10 with a working angle of 180 degrees or higher; and an intake variable valve timing mechanism for changing a phase angle of the intake valve. The internal combustion engine is configured so that a pressure difference produced by subtracting pressure of an exhaust gas from that of an intake gas is -10kPa or higher. With the intake variable valve timing mechanism, valve-opening time of the intake valve can be set in advance of an intake top dead point. The exhaust valve can be driven to open on a delay side from the intake top dead point. The piston is configured to have a maximum depth of an exhaust valve recess 54 to be equal to or less than a maximum depth of an intake valve recess 53.SELECTED DRAWING: Figure 4

Description

  The present invention relates to an internal combustion engine.

  In an internal combustion engine having a high compression ratio (for example, a compression ratio of 15 or more), the top surface of the piston is provided with an intake valve to prevent interference between the piston and the intake and exhaust valves when the piston is at the intake top dead center. It is known to provide a recess and a recess for an exhaust valve (for example, Patent Document 1).

Japanese Patent No. 5692462

  By the way, in an internal combustion engine equipped with a valve timing mechanism of an intake valve that can change the phase angle, when the phase angle of the intake valve is advanced, the closing timing of the intake valve approaches the intake bottom dead center and the charging efficiency is high. As a result, the engine output can be increased. On the other hand, when the phase angle of the intake valve is retarded, the closing timing of the intake valve is away from the bottom dead center of the intake, so that the pumping loss is reduced, and as a result, the thermal efficiency of the internal combustion engine can be increased.

  On the other hand, in an internal combustion engine provided with a valve timing mechanism of an exhaust valve capable of changing a phase angle, when the phase angle of the exhaust valve is advanced, the exhaust valve opening timing is advanced, thereby improving the exhaust efficiency. Therefore, the engine output can be increased. On the other hand, when the phase angle of the exhaust valve is retarded, the actual expansion ratio is increased by delaying the valve closing timing of the exhaust valve, so that the thermal efficiency can be increased. In addition, when the phase angle of the exhaust valve is retarded, the valve closing time is retarded to increase the valve overlap, thereby reducing pumping loss and, as a result, increasing the thermal efficiency of the internal combustion engine. Can be.

  However, when the phase angle of the exhaust valve is retarded, the lift amount of the exhaust valve at the intake top dead center increases accordingly. In the internal combustion engine having a high compression ratio as described above, since the combustion chamber volume at the top dead center is small, it is necessary to deepen the exhaust valve recess in order to increase the lift amount of the exhaust valve at the intake top dead center. Become.

  Here, when the area of the intake valve or the exhaust valve is increased to increase the intake efficiency or the exhaust efficiency, the intake valve recess or the exhaust valve recess is formed on the top surface of the piston so as to extend to the vicinity of the outer periphery of the piston. Therefore, in order to secure the strength of the piston, it is necessary to provide the groove for the top ring at a position farther from the top surface of the piston than the depth position of the recess for the intake valve or the recess for the exhaust valve. In other words, the height of the top land of the piston needs to be higher than the depth of the intake valve recess and the exhaust valve recess.

  However, as the height of the top land of the piston increases, the clevis area increases accordingly. When the clevis region becomes large, the air-fuel mixture in the clevis region does not burn, so that the unburned loss increases, and as a result, the thermal efficiency of the internal combustion engine decreases.

  The present invention has been made in view of the above problems, and an object of the present invention is to increase thermal efficiency as much as possible while ensuring a sufficient output at an output point of an internal combustion engine in an internal combustion engine having a compression ratio of 15 or more. is there.

  The present invention has been made to solve the above problems, and the gist thereof is as follows.

  (1) An internal combustion engine having a mechanical compression ratio of 15 or more, a piston reciprocating in a cylinder, an intake valve opening and closing an intake port communicating with the cylinder, and having an operating angle of 180 ° or more, and the cylinder An exhaust valve that opens and closes an exhaust port communicating with the exhaust valve and has an operation angle of 180 ° or more, and an intake variable valve timing mechanism that changes a phase angle of the intake valve; a combustion chamber formed in the cylinder; The pressure of the intake gas is controlled so that the pressure difference obtained by subtracting the pressure of the exhaust gas discharged from the combustion chamber from the pressure of the supplied intake gas is -10 kPa or more. The valve timing mechanism is configured such that the opening timing of the intake valve can be set to a timing at which the charging efficiency of the intake gas on the advanced side of the intake top dead center is maximized, The exhaust valve can be driven so as to close on a more retarded side than the intake top dead center, and the piston is so arranged that it does not interfere with the intake valve when the piston is at the intake top dead center. An intake valve recess provided on the top surface of the piston, and an exhaust valve recess provided so as not to interfere with the exhaust valve when the piston is at the intake top dead center. An internal combustion engine, wherein a maximum depth is equal to or less than a maximum depth of the intake valve recess.

  Advantageous Effects of Invention According to the present invention, in an internal combustion engine having a compression ratio of 15 or more, it is possible to increase the thermal efficiency as much as possible while ensuring a sufficient output at the output point of the internal combustion engine.

FIG. 1 is a diagram schematically showing an internal combustion engine in which the exhaust emission control device according to one embodiment is used. FIG. 2 is a perspective view schematically showing a piston. FIG. 3 is a plan view of the piston as viewed from the combustion chamber side. FIG. 4 is a cross-sectional view of the vicinity of the top surface of the piston, taken along IV-IV in FIG. 3. FIG. 5 is a diagram showing the phase angles of the intake valve and the exhaust valve. FIG. 6 is a diagram showing the phase angles of the intake valve and the exhaust valve. FIG. 7 is a diagram illustrating the relationship between the mechanical compression ratio and the thermal efficiency. FIG. 8 is a diagram illustrating the relationship between the mechanical compression ratio and the thermal efficiency.

  Hereinafter, embodiments of the present invention will be described in detail with reference to the drawings. In the following description, the same reference numerals are assigned to similar components.

<Description of the entire internal combustion engine>
FIG. 1 is a diagram schematically showing an internal combustion engine 100 in which an exhaust emission control device according to one embodiment is used. As shown in FIG. 1, the internal combustion engine 100 includes an engine body 1 including a cylinder block 2 and a cylinder head 3 fixed on the cylinder block 2. A plurality of cylinders 4 are formed in the cylinder block 2, and a piston 5 reciprocates in these cylinders 4. The cylinder 4, the piston 5, and the cylinder head 3 define a combustion chamber 6 in which the air-fuel mixture burns.

  In the present embodiment, the combustion chamber 6 is configured such that the mechanical compression ratio has a relatively high value of 15 or more. Here, the mechanical compression ratio is defined as the volume of the combustion chamber 6 when the piston 5 is at the bottom dead center (the volume obtained by adding the stroke volume to the volume of the combustion chamber 6 when the piston 5 is at the top dead center). Is a value obtained by dividing by the volume of the combustion chamber 6 at the top dead center.

  An intake port 7 communicating with the cylinder 4 and an exhaust port 8 communicating with the cylinder 4 are formed in the cylinder head 3. The cylinder head 3 is provided with an intake valve 9 that opens and closes an intake port 7 with respect to the combustion chamber 6 and an exhaust valve 10 that opens and closes an exhaust port 8 with respect to the combustion chamber 6.

  Further, as shown in FIG. 1, a spark plug 11 is disposed at the center of the inner wall surface of each cylinder head 3 facing each cylinder 4. The spark plug 11 is configured to generate a spark according to an ignition signal. In addition, a fuel injection valve 12 is arranged around the intake port 7 of the cylinder head 3. The fuel injection valve 12 injects a predetermined amount of fuel into the intake port 7 according to an injection signal. Note that the fuel injection valve 12 may be arranged facing the combustion chamber 6 so as to directly inject fuel into the combustion chamber 6.

  The intake port 7 of each cylinder is connected to a surge tank 14 via a corresponding intake branch pipe 13, and the surge tank 14 is connected to an air cleaner 16 via an intake pipe 15. The intake port 7, the intake branch pipe 13, the surge tank 14, and the intake pipe 15 form an intake passage. A throttle valve 18 driven by a throttle valve drive actuator 17 is arranged in the intake pipe 15. The throttle valve 18 is rotated by the throttle valve drive actuator 17 so that the opening area of the intake passage can be changed.

  On the other hand, the exhaust port 8 of each cylinder is connected to an exhaust manifold 19. The exhaust manifold 19 is connected to an upstream casing 21 containing an exhaust purification catalyst 20. The outlet of the upstream casing 21 is connected to the exhaust pipe 22. The exhaust port 8, the exhaust manifold 19, the upstream casing 21, and the exhaust pipe 22 form an exhaust passage.

  The internal combustion engine 100 includes an EGR mechanism. The EGR mechanism includes an exhaust gas recirculation (EGR) passage 25 that connects the exhaust manifold 19 and the surge tank 14 to each other, and an electrically controlled EGR control valve 26 disposed in the EGR passage 25. A cooling device 27 for cooling the EGR gas flowing in the EGR passage 25 is disposed around the EGR passage 25. By adjusting the opening of the EGR control valve 26, the flow rate of the exhaust gas supplied again into the combustion chamber 6 can be controlled. As a result, the EGR rate of the intake gas sucked into the combustion chamber 6 can be reduced. Can be controlled. Here, the EGR rate means the ratio of the amount of EGR gas to the total amount of intake gas sucked into the combustion chamber 6.

  The internal combustion engine 100 of the present embodiment further includes an intake variable valve timing mechanism 28 and an exhaust variable valve timing mechanism 29. The variable intake valve timing mechanism 28 is configured to change the phase angle of the intake valve 9. Further, the intake variable valve timing mechanism 28 of the present embodiment is configured such that the operating angle of the intake valve 9 is maintained at a predetermined angle of 180 ° or more and is not changed. Therefore, according to the intake variable valve timing mechanism 28, the valve opening timing and the valve closing timing of the intake valve 9 can be changed while keeping their angular intervals constant.

  Similarly, the exhaust variable valve timing mechanism 29 is configured so that the phase angle of the exhaust valve 10 can be changed. Further, the exhaust variable valve timing mechanism 29 of the present embodiment is configured such that the operating angle of the exhaust valve 10 is maintained at a predetermined angle of 180 ° or more and is not changed. Therefore, according to the exhaust variable valve timing mechanism 29, the valve opening timing and the valve closing timing of the exhaust valve 10 can be changed while keeping their angular intervals constant. The internal combustion engine 100 does not necessarily need to include the exhaust variable valve timing mechanism 29.

  The internal combustion engine 100 of the present embodiment includes an electronic control unit (ECU) 31 and a control device including various sensors. The ECU 31 is composed of a digital computer and is connected to each other via a bidirectional bus 32. A RAM (random access memory) 33, a ROM (read only memory) 34, a CPU (microprocessor) 35, an input port 36, and an output port 37 Is provided. An air flow meter 39 for detecting the flow rate of air flowing through the intake pipe 15 is arranged in the intake pipe 15. In addition, an air-fuel ratio sensor 40 that detects an air-fuel ratio of exhaust gas is disposed in the exhaust manifold 19. The outputs of the air flow meter 39 and the air-fuel ratio sensor 40 are input to the input port 36 via the corresponding AD converter 38.

  A load sensor 43 that generates an output voltage proportional to the amount of depression of the accelerator pedal 42 is connected to the accelerator pedal 42, and the output voltage of the load sensor 43 is input to the input port 36 via the corresponding AD converter 38. You. The crank angle sensor 44 generates an output pulse every time the crankshaft rotates 15 degrees, for example, and this output pulse is input to the input port 36. The CPU 35 calculates the engine speed from the output pulse of the crank angle sensor 44. On the other hand, the output port 37 is connected to the ignition plug 11, the fuel injection valve 12, the throttle valve drive actuator 17 and the EGR control valve 26 via the corresponding drive circuit 45.

  Further, the internal combustion engine 100 of the present embodiment is a lean burn internal combustion engine. Therefore, when the engine load is low, an air-fuel mixture having an air-fuel ratio of 40 or more is supplied to the combustion chamber 6. Therefore, even when the engine load is low, the opening of the throttle valve 18 is set to a relatively large opening, and the pressure of the intake gas in the intake passage downstream of the throttle valve 18 is maintained at a relatively high pressure. Therefore, at this time, the pressure difference obtained by subtracting the pressure of the exhaust gas discharged from the combustion chamber 6 from the pressure of the intake gas supplied to the combustion chamber 6 is −10 kPa or more. As a result, pumping loss caused by reducing the opening of the throttle valve 18 can be reduced.

  When the engine load is high, the combustion chamber 6 is supplied with an air-fuel mixture whose air-fuel ratio is close to the stoichiometric air-fuel ratio. Also in this case, the opening of the throttle valve 18 is set to a relatively large opening, so that the pressure of the intake gas in the intake passage downstream of the throttle valve 18 is maintained at a relatively high pressure.

  Although the internal combustion engine 100 of the present embodiment is a lean burn internal combustion engine, if the pumping loss can be reduced by increasing the pressure of the intake gas in the intake passage downstream of the throttle valve 18, a different internal combustion engine will be used. It may be an institution. For example, the pressure of the intake gas in the intake passage downstream of the throttle valve 18 can also be increased by introducing the EGR gas into the intake passage by the EGR mechanism. In this case, when the load on the internal combustion engine is low, the pressure difference obtained by subtracting the pressure of the exhaust gas discharged from the combustion chamber 6 from the pressure of the intake gas supplied to the combustion chamber 6 is -10 kPa or more. , The EGR rate by the EGR mechanism is controlled.

  In any case, when the load on the internal combustion engine 100 is low, the pressure difference obtained by subtracting the pressure of the exhaust gas discharged from the combustion chamber 6 from the pressure of the intake gas supplied to the combustion chamber 6 is −10 kPa or more. Any configuration may be used as long as the pressure of the intake gas can be controlled so that

<Structure of piston>
Next, the configuration of the piston 5 used in the internal combustion engine 100 of the present embodiment will be described with reference to FIGS. FIG. 2 is a perspective view schematically showing the piston 5, and FIG. 3 is a plan view of the piston 5 viewed from the combustion chamber 6 side. FIG. 4 is a cross-sectional view of the vicinity of the top surface of the piston 5 taken along the line IV-IV in FIG.

  As can be seen from FIGS. 2 to 4, a flat portion 51, a raised portion 52, an intake valve recess 53, and an exhaust valve recess 54 are provided on the top surface of the piston 5. In this specification, the direction from the piston 5 toward the combustion chamber 6 in the direction of the axis X of the piston 5 is referred to as upward, and the direction from the piston 5 toward the crankshaft (not shown) in the direction of the axis X of the piston 5 is referred to as downward. . However, the engine body 1 does not necessarily need to be arranged so that the axis X of the piston 5 extends in the vertical direction, and may be arranged such that the axis X of the piston 5 extends inclining with respect to the vertical direction, or may be arranged in the horizontal direction. May be arranged so as to extend.

  The flat portion 51 is formed on the top surface of the piston 5 and extends perpendicular to the axis X of the piston 5. The raised portion 52 is configured to protrude upward from the flat portion 51. The raised portion 52 has an intake-side inclined surface 52a arranged so as to face a part of the valve element 9a of the intake valve 9 when the piston 5 is at the top dead center, and when the piston 5 is at the top dead center. The exhaust valve 10 includes an exhaust-side inclined surface 52b disposed so as to face a part of the valve body 10a of the exhaust valve 10, and a protruding surface 52c provided between the inclined surfaces 52a and 52b.

  The intake-side inclined surface 52a is formed so as to extend substantially perpendicularly to the axis of the intake valve 9, and to extend in parallel with the plane of the valve body 9a of the intake valve 9 on the combustion chamber 6 side. The exhaust-side inclined surface 52b is formed so as to extend substantially perpendicularly to the axis of the exhaust valve 10, and thus to extend in parallel with the plane of the valve body 10a of the exhaust valve 10 on the combustion chamber 6 side. The protruding surface 52c extends parallel to the plane of the flat portion 51 and perpendicular to the axis X of the piston 5.

  The intake valve recess 53 is formed so as to be recessed downward from the flat portion 51. The intake valve recess 53 is formed at a position facing a part of the valve element 9a of the intake valve 9 when the piston 5 is at the top dead center. In the present embodiment, since two intake valves 9 are provided for one cylinder 4, two intake valve recesses 53 are provided for each piston 5. The intake valve recess 53 has an inclined surface extending substantially perpendicularly to the axis of the intake valve 9, and the inclined surface is continuous from the intake side inclined surface 52 a in parallel with the intake side inclined surface 52 a of the raised portion 52. It is configured to extend. Therefore, the intake valve recess 53 is provided on the top surface of the piston 5 so that the intake valve 9 and the piston 5 do not interfere with each other when the piston 5 is at the intake top dead center.

  The exhaust valve recess 54 is formed so as to be recessed downward from the flat portion 51. The exhaust valve recess 54 is formed at a position facing a part of the valve element 10a of the exhaust valve 10 when the piston 5 is at the top dead center. In the present embodiment, since two exhaust valves 10 are provided for one cylinder 4, each piston 5 is provided with two exhaust valve recesses 54. The exhaust valve recess 54 has an inclined surface extending substantially perpendicular to the axis of the exhaust valve 10, and this inclined surface is continuous from the exhaust side inclined surface 52 b in parallel with the exhaust side inclined surface 52 b of the raised portion 52. It is configured to extend. Therefore, the exhaust valve recess 54 is provided on the top surface of the piston 5 so that the exhaust valve 10 does not interfere with the piston 5 when the piston 5 is at the intake top dead center.

  The outer peripheral surface of the piston 5 is provided with three ring grooves 55 to 57 for accommodating the piston ring. The ring grooves 55 are formed so as to extend over the entire circumference in the circumferential direction of the piston 5 on a plane perpendicular to the axis X of the piston 5. Of the three ring grooves 55 to 57, a top ring (not shown) is disposed in the first ring groove 55 provided at the uppermost position, and a second ring (not shown) is disposed in the second ring groove 56 provided in the center. (Not shown). An oil ring (not shown) is arranged in the third ring groove 57 provided at the lowest position among the three ring grooves 55 to 57. A region from the top surface of the piston 5 to the first ring groove 55 in the outer peripheral surface of the piston 5 is referred to as a top land 58.

  Here, the maximum value din of the depth of the intake valve recess 53 from the plane of the flat portion 51 of the piston 5 is referred to as the maximum depth of the intake valve recess 53, and is used for the exhaust valve from the plane of the flat portion 51 of the piston 5. The maximum value dex of the depth of the recess 54 is referred to as the maximum depth of the exhaust valve recess 54. In the present embodiment, the exhaust valve recess 54 is configured such that the maximum depth dex of the exhaust valve recess 54 is equal to or less than the maximum depth din of the intake valve recess 53. In particular, the exhaust valve recess 54 is preferably configured such that the maximum depth dex of the exhaust valve recess 54 is smaller than the maximum depth din of the intake valve recess 53.

  The maximum lift that the intake valve 9 and the exhaust valve 10 can take when the piston 5 is at the intake top dead center changes according to the depth of the intake valve recess 53 and the exhaust valve recess 54. The deeper the recesses 53 and 54 are, the greater the maximum lift that the intake valve 9 and the exhaust valve 10 can take. Therefore, in the present embodiment, the maximum lift that the exhaust valve 10 can take when the piston 5 is at the intake top dead center is equal to or less than the maximum lift that the intake valve 9 can take when the piston 5 is at the intake top dead center. Or smaller.

  The region where the depth of the intake valve recess 53 and the exhaust valve recess 54 is maximum is located close to the outer peripheral surface of the piston 5. For this reason, the length of the piston 5 from the plane of the flat portion 51 of the piston 5 to the first ring groove 55 in the direction of the axis X (hereinafter, referred to as “height of the top land 58”) is defined as the intake valve recess 53 and the exhaust. When the depth is approximately the same as the maximum depth of the valve recess 54, the strength of the piston 5 in the vicinity of the top land 58 decreases. Accordingly, the first ring groove 55 is such that the height of the top land 58 is larger than the maximum depth of the deeper one of the intake valve recess 53 and the exhaust valve recess 54 (that is, the intake valve recess 53 in the present embodiment). It is arranged to be somewhat larger. In particular, in the present embodiment, the first ring groove 55 is arranged such that the height of the top land 58 is as small as possible within a range where the strength of the piston 5 can be sufficiently ensured.

<Control of intake and exhaust valves>
Next, control of the phase angle (valve timing) of the intake valve 9 and the exhaust valve 10 will be described with reference to FIG. FIG. 5 is a diagram illustrating the phase angles of the intake valve 9 and the exhaust valve 10. 5A shows the phase angle of the intake valve 9 and the exhaust valve 10 when the engine load is high, and FIG. 5B shows the phase angle of the intake valve 9 and the exhaust valve 10 when the engine load is low. The corner is shown.

  In FIG. 5, IVO indicates the opening timing of the intake valve 9, IVC indicates the closing timing of the intake valve 9, EVO indicates the opening timing of the exhaust valve 10, and EVC indicates the closing timing of the exhaust valve 10, respectively. . A period indicated by an arrow IN in the drawing indicates a valve opening period of the intake valve 9, and a period indicated by an arrow EX in the drawing indicates a valve opening period of the exhaust valve 10.

  As shown in FIG. 5A, when the engine load is high, the phase angle of the intake valve 9 is relatively advanced. In the example shown in FIG. 5A, the operating angle of the intake valve 9 is set to 235 °, the opening timing IVO of the intake valve 9 is set to 25 ° BTDC, and the intake valve 9 is closed. Timing IVC is set to 30 ° ABDC. By setting the valve opening timing IVO of the intake valve 9 to be more advanced than the top dead center, a valve overlap period (VO in the figure) in which both the intake valve 9 and the exhaust valve 10 are simultaneously opened is created. Thereby, the charging efficiency of the intake gas can be increased. Further, by slightly retarding the valve closing timing IVC of the intake valve 9 from the bottom dead center, the charging efficiency of the intake gas can be increased by the inertial force of the intake gas.

  Thus, by setting the valve opening timing IVO of the intake valve 9 to a timing advanced from the intake top dead center and setting the valve closing timing IVC to a timing retarded from the intake bottom dead center, The charging efficiency of the intake gas can be increased. In particular, the phase angle of the intake valve 9 is set so that the valve opening timing IVO of the intake valve 9 is advanced from the intake top dead center and the valve closing timing IVC is retarded from the intake bottom dead center. When the phase angle is set, the charging efficiency of the intake gas is maximized at the intake valve of that working angle. In the present embodiment, the intake variable valve timing mechanism 28 controls the intake valve 9 within a range including a timing at which the intake valve 9 has a phase angle at which the intake gas filling efficiency of the intake valve at its working angle is maximized. 9 can be controlled. As a result, a sufficient output can be ensured at the output point in the operating state where the output of the internal combustion engine 100 is the highest.

  On the other hand, as shown in FIG. 5A, when the engine load is high, the phase angle of the exhaust valve 10 is set to be relatively advanced. In the example shown in FIG. 5A, the operating angle of the exhaust valve 10 is set to 230 °, the opening timing EVO of the exhaust valve 10 is set to 45 ° BBDC, and the closing timing of the exhaust valve 10 is set. EVC is set to 5 ° ATDC. By setting the valve opening timing EVO of the exhaust valve 10 to be more advanced than the bottom dead center of the exhaust gas, the exhaust gas can be forcibly discharged using the pressure of the expansion of the air-fuel mixture. Further, by delaying the valve closing timing EVC of the exhaust valve 10 from the intake top dead center, a valve overlap period is created, so that the charging efficiency of the intake gas is increased.

  Further, the phase angles of the intake valve 9 and the exhaust valve 10 are changed according to the engine speed. When the engine load is high, the phase angles of the intake valve 9 and the exhaust valve 10 are basically advanced as the engine speed increases. Thereby, the charging efficiency of the intake gas can be increased.

  On the other hand, as shown in FIG. 5B, when the engine load is low, the phase angle of the intake valve 9 is relatively retarded. In the example shown in FIG. 5B, at this time, the valve opening timing IVO of the intake valve 9 is set to 0 ° BTDC, and the valve closing timing IVC of the intake valve 9 is set to 55 ° ABDC. By delaying the valve closing timing IVC of the intake valve 9 in this manner, pumping loss is reduced, so that thermal efficiency can be increased.

  In addition, in the present embodiment, as shown in FIG. 5B, even when the engine load is low, the phase angle of the exhaust valve 10 is set to be relatively advanced. In the example shown in FIG. 5B, the phase angle of the exhaust valve 10 is set to be the same as when the engine load is high, so that the valve opening timing EVO of the exhaust valve 10 is set to 45 ° BBDC, The valve closing timing EVC of No. 10 is set to 5 ° ATDC.

  In addition, even when the engine load is low, the phase angles of the intake valve 9 and the exhaust valve 10 can be changed according to the engine speed. When the engine load is low, the phase angle of the intake valve 9 is advanced as the engine speed increases when the engine speed is low (for example, less than 2,000 rpm), and when the engine speed is high (for example, (2000 rpm or more), it is retarded as the engine speed increases.

  In the present embodiment, the relationship between the engine load and the engine rotation speed and the phase angle of the intake valve 9 and the phase angle of the exhaust valve 10 is calculated in advance experimentally or by a test, and stored in the ROM 34 of the ECU 31 as a map. Then, based on the engine load detected by the load sensor 43 and the engine rotation speed calculated based on the output of the crank angle sensor 44, the target phase angle of the intake valve 9 and the target phase angle of the intake valve 9 are determined using a map stored in the ROM 34. A target phase angle of the exhaust valve 10 is calculated. The variable intake valve timing mechanism 28 and the variable exhaust valve timing mechanism 29 control the phase angles of the intake valve 9 and the exhaust valve 10 so as to achieve the target phase angle.

<Action / Effect>
The operation and effect of the internal combustion engine 100 according to the embodiment will be described with reference to FIGS. FIG. 6 is a diagram illustrating the phase angles of the intake valve 9 and the exhaust valve 10. FIG. 6 shows a case where the phase angle of the exhaust valve 10 is set to a relatively retarded side for comparison with the case where the phase angle of the exhaust valve 10 shown in FIG. 5 is set to a relatively advanced side. ing. FIG. 6A shows a case where the phase angle of the intake valve 9 is set to a relatively advanced side similarly to FIG. 5A, and FIG. FIG. 9B shows a case where the angle is set relatively to the retard side similarly to FIG.

  In the example shown in FIGS. 6A and 6B, the operating angle of the exhaust valve 10 is set to 230 °, the valve opening timing EVO of the exhaust valve 10 is set to 15 ° BBDC, and the exhaust The valve closing timing EVC of the valve 10 is set to 35 ° ATDC.

  In the example shown in FIG. 6A, similarly to FIG. 5A, the valve opening timing IVO of the intake valve 9 is set to 25 ° BTDC, and the valve closing timing IVC of the intake valve 9 is set to 30 ° ABDC. Is set to On the other hand, in the example shown in FIG. 6B, the valve opening timing IVO of the intake valve 9 is set to 5 ° ATDC, and the valve closing timing IVC of the intake valve 9 is set to 60 ° ABDC.

  When the phase angle of the exhaust valve 10 is set in this manner, the exhaust valve 10 is more open than the intake valve 9 when the piston 5 is at the intake top dead center. Therefore, the exhaust valve recess 54 is formed so that the maximum depth thereof is larger than the maximum depth of the intake valve recess 53. Then, as described above, the height of the top land 58 is set according to the deepest maximum depth of the intake valve recess 53 and the exhaust valve recess 54. Therefore, when the phase angle of the exhaust valve 10 is set in this manner, the piston 5 is configured such that the height of the top land 58 is relatively high.

  7 and 8 are diagrams showing the relationship between the mechanical compression ratio and the thermal efficiency. A broken line A in the figure indicates a case where the phase angle of the intake valve 9 is set on the advance side and the phase angle of the exhaust valve 10 is set on the advance side, that is, these phase angles are shown in FIG. Is set as follows. A dashed line B in the figure indicates a case where the phase angle of the intake valve 9 is set to the retard side and the phase angle of the exhaust valve 10 is set to the advance side, that is, these phase angles are shown in FIG. FIG.

  On the other hand, a broken line C in the figure indicates that the phase angle of the intake valve 9 is set to the advanced side and the phase angle of the exhaust valve 10 is set to the retard side, that is, these phase angles are changed to those in FIG. It shows a case where the settings are made as shown. A dashed line D in the figure indicates a case where the phase angle of the intake valve 9 is set to the retard side and the phase angle of the exhaust valve 10 is set to the retard side, that is, these phase angles are shown in FIG. FIG.

  In addition, these lines A to D correspond to the maximum of the exhaust valve recess 54 so that the exhaust valve 10 does not interfere with the piston 5 even if the phase angle of the exhaust valve 10 is retarded as shown in FIG. The case where the piston 5 is formed so that the depth is larger than the maximum depth of the intake valve recess 53 is shown. Therefore, the lines A to D show the case where the height of the top land 58 of the piston 5 is relatively high.

  On the other hand, the solid line E in the figure indicates that the phase angle of the intake valve 9 is set to the retard side and the phase angle of the exhaust valve 10 is set to the advance side, that is, these phase angles are as shown in FIG. It shows a case where the settings are made as shown. In addition, a solid line E in the figure shows a case where the piston 5 is formed such that the maximum depth of the exhaust valve recess 54 is equal to or less than the maximum depth of the intake valve recess 53. Therefore, the solid line E indicates a case where the height of the top land 58 of the piston 5 is relatively low. In the case indicated by the solid line E, since the phase angle of the exhaust valve 10 is set on the advance side, even if the maximum depth of the exhaust valve recess 54 is reduced, the exhaust valve 10 and the piston 5 Does not interfere even at the intake top dead center.

  Here, FIG. 7 shows the relationship between the mechanical compression ratio and the thermal efficiency in an internal combustion engine that is not a lean burn internal combustion engine and cannot supply a large amount of EGR gas to the combustion chamber 6. Therefore, FIG. 7 shows that the pressure difference obtained by subtracting the pressure of the exhaust gas exhausted from the combustion chamber 6 from the pressure of the intake gas supplied to the combustion chamber 6 (hereinafter, also referred to as the “intake and exhaust pressure difference”) is −10 kPa. Also shows the relationship in an internal combustion engine operating in a low state.

  As can be seen from FIG. 7, in the internal combustion engine that is operated in a state where the intake / exhaust pressure difference is lower than −10 kPa, the dashed line D has the highest thermal efficiency at any mechanical compression ratio. That is, in the internal combustion engine that is operated in a state where the intake / exhaust pressure difference is lower than −10 kPa, the phase angle of the intake valve 9 is set to the retard side and the phase angle of the exhaust valve 10 is set to the retard side. It can be seen that the thermal efficiency is high regardless of the mechanical compression ratio.

  On the other hand, FIG. 8 shows the relationship between the mechanical compression ratio and the thermal efficiency in a lean burn internal combustion engine or an internal combustion engine capable of supplying a large amount of EGR gas to the combustion chamber 6. Therefore, FIG. 8 shows a relationship in the internal combustion engine 100 that is operated in a state where the intake / exhaust pressure difference is −10 kPa or more.

  As can be seen from FIG. 8, in the region where the mechanical compression ratio is less than 15, the thermal efficiency indicated by the dashed line D is the highest. Therefore, even in an internal combustion engine that is operated in a state where the intake / exhaust pressure difference is lower than −10 kPa, in a region where the mechanical compression ratio is less than 15, the phase angle of the intake valve 9 is set to the retard side and the exhaust valve 10 When the phase angle is set to the retard side, it can be said that the thermal efficiency is high.

  However, it can be seen that the solid line E has the highest thermal efficiency in the region where the mechanical compression ratio is 15 or more. Therefore, in an internal combustion engine that is operated in a state where the intake / exhaust pressure difference is lower than −10 kPa, in the region where the mechanical compression ratio is 15 or more, the maximum depth of the exhaust valve recess 54 is reduced and the top land of the piston 5 is reduced. When the height of the valve 58 is reduced, and the phase angle of the intake valve 9 is set to the retard side and the phase angle of the exhaust valve 10 is set to the advance side, the thermal efficiency increases.

  As described above, in the internal combustion engine 100 operated with the intake / exhaust pressure difference of −10 kPa or more, the solid line E has the highest thermal efficiency in the region where the mechanical compression ratio is 15 or more for the following reason. It is believed that there is.

  That is, as the mechanical compression ratio increases, the ratio of the clevis region (region between the top land 58 of the piston 5 and the inner surface of the cylinder 4) to the volume of the combustion chamber 6 when the piston 5 is at the top dead center increases. . Since the air-fuel mixture present in the clevis region remains unburned even when ignited by the air-fuel mixture in the combustion chamber 6, unburned loss increases as the ratio of the clevis region increases. Therefore, the unburned loss increases as the mechanical compression ratio increases.

  Here, as described above, when the height of the top land 58 decreases, the clevis region decreases. Therefore, when the height of the top land 58 is reduced, the unburned loss can be reduced. Therefore, the unburned loss can be reduced by reducing the maximum depth of the exhaust valve recess 54.

  On the other hand, by retarding the phase angle of the exhaust valve 10, the valve opening timing IVO of the exhaust valve 10 is retarded. For this reason, the conversion ratio of the combustion energy of the exhaust gas to the kinetic energy of the piston 5 increases, so that the thermal efficiency of the internal combustion engine can be improved. When the mechanical compression ratio is less than 15, the improvement of the thermal efficiency by delaying the valve opening timing IVO of the exhaust valve 10 is more effective than the reduction of the unburned loss by reducing the clevis region, Is considered to have a large effect on the thermal efficiency. As a result, in the region where the mechanical compression ratio is less than 15, the thermal efficiency along the dashed line D is considered to be the highest. On the other hand, when the mechanical compression ratio is 15 or more, the reduction in the unburned loss due to the smaller clevis region is more effective than the improvement in the thermal efficiency due to the retardation of the next valve opening IVO of the exhaust valve 10 compared to the overall internal combustion engine. Is considered to have a large effect on the thermal efficiency. As a result, it is considered that the solid line E has the highest thermal efficiency in the region where the mechanical compression ratio is 15 or more.

  Further, in the internal combustion engine 100 that is operated in a state where the intake / exhaust pressure difference is lower than −10 kPa, the pumping loss is originally large. Therefore, by delaying the phase angle of the exhaust valve 10 to extend the valve overlap period, the pumping loss can be greatly reduced, and the thermal efficiency can be greatly improved. For this reason, in the internal combustion engine 100 in which the intake / exhaust pressure difference is smaller than −10 kPa, the thermal efficiency is improved by delaying the valve closing timing IVC of the exhaust valve 10, rather than by reducing the unburned loss due to the reduced clevis region. Is considered to have a greater effect on the thermal efficiency of the entire internal combustion engine. As a result, in FIG. 7, it is considered that the thermal efficiency along the dashed line D is the highest regardless of the mechanical compression ratio.

  On the other hand, in the internal combustion engine 100 that is operated with the intake / exhaust pressure difference of −10 kPa or more, the pumping loss is originally small. Therefore, even if the phase angle of the exhaust valve 10 is retarded to extend the valve overlap period, the effect of reducing the pumping loss is not so high. For this reason, in FIG. 8, the thermal efficiency of the solid line E is considered to be the highest depending on the mechanical compression ratio.

  As described above, in the present embodiment, even when the engine load is low, in the internal combustion engine having a mechanical compression ratio of 15 or more in which the pressure of the intake gas is controlled such that the intake and exhaust pressure difference is operated in a state of −10 kPa or more, By reducing the maximum depth of the exhaust valve recess 54, the height of the top land 58 of the piston 5 can be reduced. Therefore, the thermal efficiency can be made as high as possible.

DESCRIPTION OF SYMBOLS 1 Engine main body 2 Cylinder block 3 Cylinder head 4 Cylinder 5 Piston 6 Combustion chamber 9 Intake valve 10 Exhaust valve 28 Intake variable valve timing mechanism 29 Exhaust variable valve timing mechanism 51 Flat part 52 Ridge 53 Recess for intake valve 54 Recess for exhaust valve 55 First ring groove 58 Top land

Claims (1)

An internal combustion engine having a mechanical compression ratio of 15 or more,
A piston that reciprocates in a cylinder, an intake valve that opens and closes an intake port communicating with the cylinder and has an operating angle of 180 ° or more, and an exhaust port that opens and closes an exhaust port communicating with the cylinder and has an operating angle of 180 ° or more. An exhaust valve, comprising an intake variable valve timing mechanism for changing the phase angle of the intake valve,
Controlling the pressure of the intake gas so that the pressure difference obtained by subtracting the pressure of the exhaust gas discharged from the combustion chamber from the pressure of the intake gas supplied to the combustion chamber formed in the cylinder is -10 kPa or more; Is configured so that
The intake variable valve timing mechanism is configured such that the opening timing of the intake valve can be set to a timing at which the charging efficiency of intake gas on the advanced side of the intake top dead center is maximized,
The exhaust valve can be driven to close on a more retarded side than the intake top dead center,
The piston includes an intake valve recess provided at a top surface of the piston so that the piston does not interfere with the intake valve when the piston is at the intake top dead center. An internal combustion engine comprising: an exhaust valve recess provided so as not to interfere with the exhaust valve, wherein a maximum depth of the exhaust valve recess is less than or equal to a maximum depth of the intake valve recess.

Images