JP2006161691A - Internal combustion engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To suppress combustion differences between cylinders by uniformizing recirculation amount of exhaust gas in the respective cylinders, in an internal combustion engine. <P>SOLUTION: In the V 8 engine, cavity shapes of respective cylinders are changed such that a compression ratio in the first cylinder #1, third cylinder #3, second cylinder #2 and sixth cylinder #6 in which an overlap period overlaps an exhaust stroke of the specific cylinder is set larger than that in the forth cylinder #4, fifth cylinder #5, seventh cylinder #7 and eighth cylinder #8 in which the overlap period does not overlap the exhaust stroke. <P>COPYRIGHT: (C)2006,JPO&NCIPI

Description

  The present invention relates to an unequal interval ignition type V-type multi-cylinder internal combustion engine in which a plurality of cylinders are divided into left and right banks and the cylinders of each bank are ignited and exploded at unequal intervals.

  In a general V-type multi-cylinder engine, the cylinder block has two banks inclined at a predetermined angle at the top, and a piston is movably fitted to each cylinder provided in each bank. It is connected to a crankshaft rotatably supported at the bottom. In addition, each combustion chamber is configured by fastening a cylinder head to the upper part of each bank of the cylinder block, and each combustion chamber is formed with an intake port and an exhaust port, and can be opened and closed by an intake valve and an exhaust valve. It has become. The downstream part of the intake pipe is branched into two and connected to the intake port of each bank, while the upstream part of the exhaust pipe is branched into two and connected to the exhaust port of each bank. A catalytic device is mounted in the downstream portion.

  For example, in a V-type 8-cylinder engine, the left bank is provided with a first cylinder, a third cylinder, a fifth cylinder, and a seventh cylinder, and the right bank is provided with a second cylinder, a fourth cylinder, a sixth cylinder, Eight cylinders are provided, and an exhaust pipe is connected to each bang through an exhaust manifold. In order to optimize the dynamic balance of the valve train, the firing order of each cylinder is as follows: first cylinder, eighth cylinder, seventh cylinder, third cylinder, sixth cylinder, fifth cylinder, fourth cylinder, It has 2 cylinders. Further, in the V-type 8-cylinder engine, the exhaust valve closing timing is retarded and the intake valve opening timing is advanced for the purpose of reducing the pumping loss due to the piston and reducing the generated NOx. The latter half of the valve opening period overlaps the first half of the intake valve opening period.

Japanese Patent Laid-Open No. 3-070810 Special table 2003-515025 gazette

  However, in the V-type 8-cylinder engine described above, since the ignition order is not the cylinder number order, the left bank is ignited at unequal intervals in the order of the first cylinder, the seventh cylinder, the third cylinder, and the fifth cylinder. In the right bank, the eighth cylinder, the sixth cylinder, the fourth cylinder, and the second cylinder are ignited in this order, and the ignition (explosion) intervals in each bank are unequal. Therefore, in the left bank, the overlap period of the first cylinder and the exhaust stroke of the seventh cylinder overlap, and the overlap period of the third cylinder and the exhaust stroke of the fifth cylinder overlap. In the right bank, the overlap period of the second cylinder and the exhaust stroke of the eighth cylinder overlap, and the overlap period of the sixth cylinder and the exhaust stroke of the fourth cylinder overlap.

  When the overlap period of one cylinder overlaps with the exhaust stroke of the other cylinder, the exhaust valve of the other cylinder opens while the intake valve and the exhaust valve of one cylinder are open, and this other cylinder Exhaust gas discharged from the exhaust port of the engine passes through the exhaust manifold and adversely affects one cylinder as exhaust pulsation. That is, the exhaust pulsation of the other cylinder acts on one cylinder, and in this one cylinder, the exhaust gas circulation amount that returns from the exhaust manifold to the cylinder (combustion chamber) through the exhaust port, that is, the internal EGR amount. Will increase. As a result, the amount of intake air also varies due to the difference in the internal EGR amount among the cylinders, the combustion varies and becomes unstable, the output torque fluctuates, the fuel consumption is improved by the overlap, and the exhaust gas performance is improved. The effects such as cannot be obtained.

  In addition, in the engine which performs non-uniform interval ignition, there exist the technique described in the said patent document 1, 2 as what suppresses the dispersion | variation in the filling efficiency and knock characteristic by exhaust interference in a specific cylinder, for example. However, the exhaust system of the internal combustion engine disclosed in Patent Document 1 has an ejector shape in which a collection portion of the exhaust manifold of the cylinder having a small ignition interval with the preceding cylinder in the ignition order is used. It is not possible to suppress the combustion fluctuation that occurs due to the increase in the internal EGR amount. Further, the multi-cylinder internal combustion engine of Patent Document 2 has a reduced exhaust impact by reducing the overlap period, and cannot sufficiently suppress an increase in the internal EGR amount of a specific cylinder due to exhaust pulsation. In addition, it is impossible to improve fuel consumption and exhaust gas performance due to overlap.

  An object of the present invention is to solve such a problem, and an object of the present invention is to provide an internal combustion engine in which the combustion state in each cylinder is made uniform to suppress variation in combustion among the cylinders.

  In order to solve the above-described problems and achieve the object, the internal combustion engine of the present invention has a plurality of cylinders divided into left and right banks, and the cylinders of each bank are ignited at unequal intervals, In the internal combustion engine having a period in which the opening period of the exhaust valve and the opening period of the intake valve overlap in the cylinder, in each bank, a first cylinder in which the overlapping period overlaps an exhaust stroke of a specific cylinder; Combustion parameter changing means is provided for changing a combustion parameter caused by a combustion state with the second cylinder in which the overlap period does not overlap the exhaust stroke of a specific cylinder.

  In the internal combustion engine of the present invention, the combustion parameter changing means changes the compression ratio of the first cylinder or the second cylinder as the combustion parameter.

  In the internal combustion engine of the present invention, the combustion parameter changing means changes a squish rate of the first cylinder or the second cylinder as the combustion parameter.

  In the internal combustion engine of the present invention, the combustion parameter changing means changes a tumble ratio of the first cylinder or the second cylinder as the combustion parameter.

  In the internal combustion engine of the present invention, the combustion parameter changing means changes the ignition energy of the first cylinder or the second cylinder as the combustion parameter.

  In the internal combustion engine of the present invention, the combustion parameter changing means does not change the ignition energy in an idle state of the internal combustion engine.

  In the internal combustion engine of the present invention, the combustion parameter changing means changes the compression ratio, the squish rate, the tumble ratio, or the ignition energy in accordance with the overlap period.

  According to the internal combustion engine of the present invention, a plurality of cylinders are divided into left and right banks, the cylinders in each bank are ignited at unequal intervals, and the exhaust valve opening period and the intake valve opening period in each cylinder In the internal combustion engine having a period in which the engine overlaps with each other, a first cylinder in which the overlap period overlaps with an exhaust stroke of the specific cylinder, and a second cylinder in which the overlap period does not overlap with the exhaust stroke of the specific cylinder Since the combustion parameter changing means for changing the combustion parameter caused by the combustion state is provided, in the first cylinder where the overlap period overlaps, the internal exhaust gas is regenerated by the exhaust pulsation compared to the second cylinder where the overlap period does not overlap. Although the circulation amount increases, the combustion parameter changing means changes the combustion parameter of the first cylinder or the second cylinder to thereby change the first cylinder or the second cylinder. Combustion state of the cylinder is made uniform, it is possible to suppress variations in combustion between the cylinders.

  Embodiments of an internal combustion engine according to the present invention will be described below in detail with reference to the drawings. Note that the present invention is not limited to the embodiments.

  FIG. 1 is a schematic plan view of a V-type 8-cylinder engine representing an internal combustion engine according to Embodiment 1 of the present invention, FIG. 2 is a schematic configuration diagram of the V-type 8-cylinder engine of Embodiment 1, and FIG. FIG. 3-2 is a cross-sectional view of a combustion chamber of a cylinder with a large internal EGR in the V-type 8-cylinder engine of Embodiment 1, and FIG. 3-2 is a cross-sectional view of a combustion chamber of a cylinder with a low internal EGR in the V-type 8-cylinder engine of Embodiment 1. FIG. 4 is a time chart showing the opening timing of the intake valve and the exhaust valve in the V-type 8-cylinder engine of the first embodiment.

  In the first embodiment, a V-type 8-cylinder engine is applied as the internal combustion engine. In this V-type 8-cylinder engine, as shown in FIGS. 1 and 2, the cylinder block 11 has left and right banks 12 and 13 inclined at a predetermined angle at the upper portion, and each bank 12 and 13 has four banks 12 and 13 respectively. Cylinder bores 14 and 15 are formed, and pistons 16 and 17 are fitted to the respective cylinder bores 14 and 15 so as to be vertically movable. A crankshaft (not shown) is rotatably supported at the lower part of the cylinder block 11, and the pistons 16 and 17 are connected to the crankshaft via connecting rods 18 and 19, respectively.

  On the other hand, cylinder heads 20 and 21 are fastened to the upper portions of the banks 12 and 13 of the cylinder block 11, and the combustion chambers 22 and 23 are constituted by the cylinder block 11, the pistons 16 and 17, and the cylinder heads 20 and 21. ing. The intake ports 24 and 25 and the exhaust ports 26 and 27 are formed on the upper portions of the combustion chambers 22 and 23, that is, the lower surfaces of the cylinder heads 20 and 21 so as to face each other. 27, the lower end portions of the intake valves 28 and 29 and the exhaust valves 30 and 31 are located. The intake valves 28 and 29 and the exhaust valves 30 and 31 are supported by the cylinder heads 20 and 21 so as to be movable in the axial direction, and are attached in a direction to close the intake ports 24 and 25 and the exhaust ports 26 and 27. It is supported. Further, intake camshafts 32 and 33 and exhaust camshafts 34 and 35 are rotatably supported on the cylinder heads 20 and 21, and the intake cams 36 and 37 and the exhaust cams 38 and 39 are interposed via a roller rocker arm (not shown). Are in contact with the upper ends of the intake valves 28 and 29 and the exhaust valves 30 and 31.

  Accordingly, when the intake camshafts 32 and 33 and the exhaust camshafts 34 and 35 rotate in synchronization with the engine, the intake cams 36 and 37 and the exhaust cams 38 and 39 operate the roller rocker arm, and the intake valves 28 and 29 and the exhaust valve 30 and 31 move up and down at a predetermined timing to open and close intake ports 24 and 25 and exhaust ports 26 and 27, intake ports 24 and 25 and combustion chambers 22 and 23, combustion chambers 22 and 23, and exhaust port 26. , 27 can communicate with each other.

  In addition, the valve mechanism of this engine is a variable intake valve timing mechanism (VVT) 40 that controls the intake valves 28 and 29 and the exhaust valves 30 and 31 at an optimal opening / closing timing according to the operating state. 41 and an exhaust variable valve mechanism 42, 43. The intake variable valve operating mechanisms 40 and 41 and the exhaust variable valve operating mechanisms 42 and 43 are configured, for example, by providing VVT controllers at the shaft end portions of the intake camshafts 32 and 33 and the exhaust camshafts 34 and 35. The opening / closing timing of the intake valves 28, 29 and the exhaust valves 30, 31 is advanced or retarded by changing the phase of each camshaft 32, 33, 34, 35 with respect to the cam sprocket by a pump (or electric motor). It is something that can be done. In this case, each variable valve mechanism 40, 41, 42, 43 advances or retards the opening / closing timing with the operating angle (opening period) of intake valves 28, 29 and exhaust valves 30, 31 being constant. The intake camshafts 32, 33 and the exhaust camshafts 34, 35 are provided with cam position sensors 44, 45, 46, 47 for detecting their rotational phases.

  A surge tank 50 is connected to the intake ports 24 and 25 of the cylinder heads 20 and 21 via intake manifolds 48 and 49, and an intake pipe 51 is connected to the surge tank 50. An air cleaner 52 is attached to the intake port. The intake pipe 51 is provided with an electronic throttle device 53 having a throttle valve located on the downstream side of the air cleaner 52. On the other hand, a connecting pipe 58 is connected to the exhaust ports 26 and 27 via exhaust manifolds 54 and 55 and catalyst devices 56 and 57, and an exhaust pipe 59 is connected to the connecting pipe 58. A catalyst device 60 is mounted.

  The cylinder heads 20 and 21 are respectively provided with injectors 61 and 62 for injecting fuel (gasoline) directly into the combustion chambers 22 and 23. Delivery pipes 63 and 64 are connected to the injectors 61 and 62, respectively. The delivery pipes 63 and 64 can be supplied with fuel at a predetermined pressure from the high-pressure fuel pump 65. The cylinder heads 20 and 21 are equipped with spark plugs 66 and 67 that are located above the combustion chambers 22 and 23 and ignite the air-fuel mixture.

  By the way, an electronic control unit (ECU) 68 is mounted on the vehicle, and this ECU 68 can control and detect the fuel injection timing of the injectors 61 and 62, the ignition timing of the spark plugs 66 and 67, and the like. The fuel injection amount, the injection timing, the ignition timing, and the like are determined based on the engine operating state such as the intake air amount, the intake air temperature, the throttle opening, the accelerator opening, the engine speed, and the cooling water temperature. That is, an air flow sensor 69 and an intake air temperature sensor 70 are mounted on the upstream side of the intake pipe 51, and the measured intake air amount and intake air temperature are output to the ECU 68. The electronic throttle device 53 is provided with a throttle position sensor 71, and the accelerator pedal is provided with an accelerator position sensor 72, which outputs the current throttle opening and accelerator opening to the ECU 68. Further, a crank angle sensor 73 is provided on the crankshaft, and the detected crank angle is output to the ECU 68. The ECU 68 calculates the engine speed based on the crank angle. Further, the cylinder block 11 is provided with a water temperature sensor 74 and outputs the detected engine cooling water temperature to the ECU 68.

  Further, the ECU 68 can control the intake variable valve mechanisms 40 and 41 and the exhaust variable valve mechanisms 42 and 43 based on the engine operating state. That is, when the temperature is low, the engine is started, the engine is idling, or the load is light, the exhaust gas is discharged from the intake port 24, 25 or the amount of air blown back to the combustion chambers 22 and 23 is reduced, and combustion stability and fuel efficiency can be improved. Further, at the time of medium load, by increasing the overlap, the internal EGR rate is increased to improve the exhaust gas purification efficiency, and the pumping loss is reduced to improve the fuel consumption. Further, at the time of high-load low-medium rotation, the closing timing of the intake valves 28 and 29 is advanced to reduce the amount of intake air that blows back to the intake ports 24 and 25, thereby improving volumetric efficiency. At the time of high load and high rotation, the closing timing of the intake valves 28 and 29 is retarded according to the rotational speed, thereby improving the volume efficiency as the timing according to the inertial force of the intake air.

  Incidentally, in the V-type 8-cylinder engine of this embodiment, as shown in FIG. 1, the first bank # 1, the third cylinder # 3, the fifth cylinder # 5, and the seventh cylinder # 7 are connected in series to the left bank 12. In the right bank 13, the second cylinder # 2, the fourth cylinder # 4, the sixth cylinder # 6, and the eighth cylinder # 8 are provided in series, and the exhaust manifolds 54 and 55 are provided in the bangs 12 and 13, respectively. The connecting pipe 58 and the exhaust pipe 59 are connected to each other. In order to optimize the dynamic balance of the valve train, the firing order of each cylinder is as follows: first cylinder # 1, eighth cylinder # 8, seventh cylinder # 7, third cylinder # 3, sixth cylinder # 6, fifth cylinder # 5, fourth cylinder # 4, and second cylinder # 2.

  Therefore, in this V-type 8-cylinder engine, since the ignition order is not the order of the cylinder number, in the left bank, the first cylinder # 1- (180 ° CA) —the seventh cylinder # 7— (90 ° CA) —the third cylinder— (180 ° CA) -Fifth cylinder # 5- (270 ° CA) -First cylinder # 1 are ignited at unequal intervals, and in the right bank, the eighth cylinder # 8- (270 ° CA) -No. The six cylinders # 6- (180 ° CA) -fourth cylinder # 4- (90 ° CA) -second cylinder # 2- (180 ° CA) -eighth cylinder # 8 will be ignited in this order. Ignition (explosion) intervals in are uneven. Therefore, in the left bank 12, the overlap period of a specific cylinder overlaps with the exhaust stroke of another cylinder, and the exhaust gas discharged from the cylinder in the exhaust stroke passes through the exhaust manifold to the cylinder in the overlap period. Acting as exhaust pulsation, only the internal EGR amount of this cylinder increases.

  That is, as shown in FIG. 4, in the left bank 12, in the first cylinder # 1, the intake timing by the intake valve 28 is advanced by the intake variable valve mechanism 40 in the vicinity of the crank angle 360 ° CA, The exhaust timing by the exhaust valve 30 is retarded by the exhaust variable valve mechanism 42, so that an overlap period OL is provided here. On the other hand, in the seventh cylinder # 7, the exhaust stroke starts when the exhaust valve 30 starts to open near the crank angle of 360 ° CA. Therefore, in the vicinity of the crank angle 360 ° CA, the overlap period of the first cylinder # 1 overlaps the initial exhaust stroke of the seventh cylinder # 7.

  Then, when the intake valve 28 and the exhaust valve 30 of the first cylinder # 1 are opened, the exhaust valve 30 of the seventh cylinder # 7 is opened, and the exhaust gas discharged from the exhaust port 26 of the seventh cylinder # 7. Acts as exhaust pulsation on the first cylinder # 1 through the exhaust manifold 54. That is, when exhaust pulsation acts on the first cylinder # 1 from the seventh cylinder # 7, in this first cylinder # 1, the internal EGR that returns from the exhaust manifold 54 to the combustion chamber 22 through the exhaust port 26 by this exhaust pulsation. The amount increases, and the amount of air drawn from the intake port 24 decreases.

  This phenomenon is not limited to the first cylinder # 1, but the exhaust stroke of the eighth cylinder # 8 in the third cylinder # 3 in which the overlap period OL overlaps with the exhaust stroke of the fifth cylinder # 5 or in the right bank 13. This occurs in the sixth cylinder # 6 where the overlap period OL overlaps the exhaust stroke of the second cylinder # 2 and the fourth cylinder # 4 where the overlap period OL overlaps. When the internal EGR amount increases in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the fourth cylinder # 4, the fifth cylinder # 5, This is different from the internal EGR amount and the air amount of the seventh cylinder # 7 and the eighth cylinder # 8, the combustion varies and becomes unstable, and the output torque fluctuates.

  Therefore, in the first embodiment, as shown in FIGS. 3A and 3B, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8, and the fifth cylinder # as specific cylinders. 5. Compression ratios of the first cylinder # 1, the third cylinder # 3, the second cylinder # 2, and the sixth cylinder # 6 as the first cylinder in which the overlap period overlaps with the exhaust stroke of the fourth cylinder # 4 ( Combustion parameters) and the compression ratios of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinders in which the overlap period does not overlap the exhaust stroke. Changes (combustion parameter changing means) can be made to be different. In the first embodiment, as the combustion parameter changing means, the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 are provided by the pistons 16 and 17 of the banks 12 and 13, respectively. The cavity shape of each cylinder is changed so that the compression ratio becomes higher than the compression ratio of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8.

More specifically, for example, in the second bank # 2 having a large internal EGR amount due to overlapping overlap periods in the right bank 13, the cavity 80a of the piston 17 is set small as shown in FIG. In the fourth cylinder # 4 where the overlap period does not overlap and the internal EGR amount is small, the cavity 80b of the piston 17 is set small as shown in FIG. 3-2. Therefore, since the compression ratio is calculated by the following equation, the compression ratio of the second cylinder # 2 where the overlap period overlaps is set higher than the compression ratio of the fourth cylinder # 4 where the overlap period does not overlap. Will be.
Compression ratio = (compression stroke volume + combustion chamber volume) / combustion chamber volume The compression stroke volume + combustion chamber volume is the volume when the piston 17 is located at the bottom dead center, and the combustion chamber volume This is the volume at the top dead center.

  Therefore, in the first bank # 1 and the third cylinder # 3 in the left bank 12, the internal EGR amount is affected by the exhaust pulsation of the seventh cylinder # 7 and the fifth cylinder # 5 during the overlap period. To increase. On the other hand, in the fifth cylinder # 5 and the seventh cylinder # 7, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, the first cylinder # 1 and the third cylinder # 3 have different sizes from the cavities 80a of the first cylinder # 1 and the third cylinder # 3 and the cavities 80b of the fifth cylinder # 5 and the seventh cylinder # 7. 3 is set higher than the compression ratio of the fifth cylinder # 5 and the seventh cylinder # 7, so the first cylinder # 1 and the third cylinder # 3 having a large internal EGR amount have a high compression ratio. The combustion speeds of the cylinders # 1 and # 3 are increased to the same level as the combustion speeds of the fifth cylinder # 5 and the seventh cylinder # 7 having a small internal EGR amount. As a result, the combustion states of the first cylinder # 1 and the third cylinder # 3 having a large internal EGR amount and the combustion states of the fifth cylinder # 5 and the seventh cylinder # 7 having a small internal EGR amount are approximated. The combustion state in each of the 12 cylinders is almost uniform.

  Similarly, in the right bank 13, in the second cylinder # 2 and the sixth cylinder # 6, the internal EGR amount is affected by the exhaust pulsation of the eighth cylinder # 8 and the fourth cylinder # 4 during the overlap period. Quantitative increase. On the other hand, in the fourth cylinder # 4 and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, the second cylinder # 2 and the sixth cylinder # 6 are different in size from the cavity 80a of the second cylinder # 2 and the sixth cylinder # 6 and the size of the cavity 80b of the fourth cylinder # 4 and the eighth cylinder # 8. 6 is set to be higher than the compression ratio of the fourth cylinder # 4 and the eighth cylinder # 8, the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount have a high compression ratio. The combustion speeds of the cylinders # 2 and # 6 are increased to the same level as the combustion speeds of the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the combustion states of the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated. The combustion state in each of the 12 cylinders is almost uniform.

  Thus, in the internal combustion engine of the first embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. The compression ratio in the sixth cylinder # 6 is higher than the compression ratio of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set to be.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 where the cylinders do not overlap, the internal EGR amount does not increase during this period without being affected by exhaust pulsation, but the cavities of cylinders # 1, # 2, # 3, and # 6 By making the cavity 80a of 80a smaller than the cavity 80b of cylinders # 4, # 5, # 7, and # 8, the compression ratio of cylinders # 1, # 2, # 3, and # 6 is changed to cylinders # 4 and # 4. The cylinders # 1, # 2, # 3, # 6 with a large internal EGR amount are set to a high compression ratio and the combustion speed is small with a low internal EGR amount. It can be increased to the same level as the combustion speeds of # 4, # 5, # 7, and # 8. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  FIG. 5 is a plan view of a combustion chamber in a V-type 8-cylinder engine representing an internal combustion engine according to Embodiment 2 of the present invention, and FIG. 6-1 is a VI-VI cross section of FIG. FIG. 6B is a cross-sectional view illustrating an operating state of the compression ratio variable mechanism. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  In the V-type 8-cylinder engine of the second embodiment, as shown in FIGS. 5 and 6-1, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8, the fifth cylinder as specific cylinders. Compression of the first cylinder # 1, the third cylinder # 3, the second cylinder # 2, and the sixth cylinder # 6 as the first cylinder in which the overlap period overlaps with the exhaust strokes of the cylinder # 5 and the fourth cylinder # 4 Ratio (combustion parameter) and the compression ratio of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinder in which the overlap period does not overlap the exhaust stroke Can be changed (combustion parameter changing means). In the second embodiment, as the combustion parameter changing means, the compression ratio variable mechanism 81 is provided in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6. , # 2, # 3, # 6 can be changed to be higher than the compression ratios of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, the eighth cylinder # 8, And the amount of change is set according to the length of the overlap period.

  More specifically, in this embodiment, the compression ratio variable mechanism 81 is provided in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps with the exhaust stroke of a specific cylinder. For example, in the compression ratio variable mechanism 81 provided in the second cylinder # 2 of the right bank 13, as shown in FIGS. 5, 6-1, and 6-2, the lower surface of the cylinder head 21 that constitutes the combustion chamber 23. A slide ring 82 having a ring shape located around the spark plug 67 is supported at the center so as to be movable up and down. A moving oil passage 84 with a switching valve 83 is provided from above on the slide ring 82, and the hydraulic pressure of the moving oil passage 84 acts on the slide ring 82 by opening the switching valve 83. Thus, while the slide ring 82 can be moved downward, the in-cylinder pressure acts on the slide ring 82 by closing the switching valve 83, so that the slide ring 82 can be moved upward. Further, a control oil passage 85 is provided from the side of the slide ring 82, and this control oil passage 85 branches into two branch passages 85 a and 85 b via a switching valve 86, and a stopper portion of the slide ring 82. 82a or the leading end portion 84a of the moving oil passage 84 can be communicated.

  Accordingly, as shown in FIG. 6B, when the switching valve 83 is opened and the switching valve 86 is switched to the branch path 85a from the state where the slide ring 82 is stopped at the upper position, the hydraulic pressure of the moving oil path 84 is increased. While acting on the slide ring 82, the hydraulic pressure of the control oil passage 85 acts on the stopper portion 82a, the stopper portion 82a comes off from the branch passage 85a, and the slide ring 82 is moved to the lower position as shown in FIG. At this time, the stopper portion 82a is fitted into the branch passage 85b, and the branch passage 85a is connected to the distal end portion 84a of the moving oil passage 84. On the other hand, when the switching valve 83 is closed and the switching valve 86 is switched to the branch path 85b from the state where the slide ring 82 is stopped at the lower position, the in-cylinder pressure acts on the slide ring 82 and the control oil path 85. Acts on the stopper portion 82a, the stopper portion 82a is disengaged from the branch path 85b, and the slide ring 82 can move to the upper position as shown in FIG. 6-1, and at this time, the stopper portion 82a The branch path 85 b is connected to the distal end portion 84 a of the moving oil path 84 while being fitted to the branch path 85 a.

  Then, in the second cylinder # 2 of the right bank 13, the overlap period is set according to the operating state by the intake variable valve mechanism 41 and the exhaust variable valve mechanism 43. When there is almost no overlap period, the compression is performed. The ratio ring mechanism 81 moves the slide ring 82 to the upper position, and the compression ratio of the second cylinder # 2 in which the overlap period overlaps with the exhaust stroke of the specific cylinder does not overlap the overlap period with the exhaust stroke of the specific cylinder. The compression ratio is the same as that of cylinders # 4, # 5, # 7, and # 8. Further, when the overlap period is set according to the operating state by the intake variable valve mechanism 41 and the exhaust variable valve mechanism 43, the slide ring 82 is moved to the lower position by the compression ratio variable mechanism 81, so that The compression ratio of the second cylinder # 2 where the overlap period overlaps with the exhaust stroke is set higher than the compression ratio of the cylinders # 4, # 5, # 7, and # 8 where the overlap period does not overlap with the exhaust stroke of the specific cylinder. .

  Accordingly, in each of the banks 12 and 13, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the seventh cylinder # 7 and the eighth cylinder # 8 in the overlap period. The internal EGR amount increases by a predetermined amount under the influence of the exhaust pulsation of the fifth cylinder # 5 and the fourth cylinder # 4. On the other hand, in the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, when the overlap period occurs in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the compression ring variable mechanism 81 moves the slide ring 82 to the lower position. Since the compression ratio is increased by reducing the volume of the combustion chamber 23, the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount have a high compression ratio, and the cylinders # 1, # 2, and # 6 have a high compression ratio. The combustion speeds of # 3 and # 6 are increased to the same level as the combustion speeds of cylinders # 4, # 5, # 7 and # 8 having a small internal EGR amount. As a result, the combustion states of cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount and the combustion states of cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount are approximated, The combustion state in each cylinder of each bank 12 and 13 is almost uniform.

  Thus, in the internal combustion engine of the second embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. When the compression ratio variable mechanism 81 is provided in the sixth cylinder # 6 and the overlap period is set by the variable intake valve mechanisms 40, 41 and the variable exhaust valve mechanisms 42, 43, the cylinders # 1, # 2 , # 3, # 6, the compression ratio becomes higher than the compression ratio of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set as follows.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 that do not overlap, the internal EGR amount does not increase during this period and is not affected by exhaust pulsation, but overshoots cylinders # 1, # 2, # 3, and # 6. When the lap period occurs, the compression ratio variable mechanism 81 moves the slide ring 82 to the lower position to reduce the volume of the combustion chamber 23, thereby reducing the compression ratio of the cylinders # 1, # 2, # 3, and # 6. The cylinders # 4, # 5, # 7, and # 8 have a higher compression ratio, and the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount have a high compression ratio, and the combustion speed is internal. The same as the combustion speed of cylinders # 4, # 5, # 7, and # 8 with a small EGR amount It can be increased. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  In the second embodiment, the variable compression ratio mechanism 81 can be moved to two positions, an upper position and a lower position, by hydraulic pressure or in-cylinder pressure. However, by using an electric motor or the like, an intake variable valve mechanism can be used. 40 and 41 and the variable exhaust valve mechanisms 42 and 43 may be adjusted to be continuously open according to the overlap period set.

  FIG. 7-1 is a plan view of a combustion chamber of a cylinder having a large internal EGR in a V-type 8-cylinder engine representing an internal combustion engine according to Embodiment 3 of the present invention, and FIG. 7-2 is a V-type 8-cylinder of Embodiment 3. It is a top view of the combustion chamber of a cylinder with few internal EGR in an engine. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  In the V-type 8-cylinder engine of the third embodiment, as shown in FIGS. 7A and 7B, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8 as specific cylinders, 1st cylinder # 1, 3rd cylinder # 3, 2nd cylinder # 2, 6th cylinder # 6 as the 1st cylinder where the overlap period overlaps with the exhaust stroke of 5th cylinder # 5 and 4th cylinder # 4 Of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 as the second cylinder in which the overlap period does not overlap with this exhaust stroke (combustion parameter). It can be changed (combustion parameter changing means) so that the squish rate is different. In the first embodiment, as the combustion parameter changing means, the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 are provided by the pistons 16 and 17 of the banks 12 and 13, respectively. The cavity shape of each cylinder is changed so that the squish rate is higher than the squish rate of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8.

  More specifically, for example, in the second bank # 2 having a large internal EGR amount due to overlapping overlap periods in the right bank 13, as shown in FIG. A cavity 91a is set to a predetermined size, two intake side valve recesses 92a are formed on one side of the cavity 91a, and an exhaust side valve recess 93a is formed on the other side. The cavity 91a, the intake side valve recess 92a, and the exhaust side By connecting the valve recess 93a with connecting surfaces 94a and 95a, a squish portion 96a inclined along the ceiling surface (cylinder head 21) of the combustion chamber 23 is set around the valve recess 93a. On the other hand, in the fourth cylinder # 4 in which the overlap period does not overlap and the amount of internal EGR is small, as shown in FIG. 7-2, the cavity 91b is set to a predetermined size at the approximate center of the top surface of the piston 17. Two intake side valve recesses 92b are formed on one side of the periphery, and an exhaust side valve recess 93b is formed on the other side. The cavity 91b, the intake side valve recess 92b, and the exhaust side valve recess 93b are connected by connecting surfaces 94b and 95b. Thus, a squish portion 96b that is inclined along the ceiling surface (cylinder head 21) of the combustion chamber 23 is set around the squish portion 96b.

In this case, the intake side valve recess 92a, the exhaust side valve recess 93a, and the connection surfaces 94a, 95a in the second cylinder # 2 are set smaller than the intake side valve recess 92b, the exhaust side valve recess 93b, and the connection surfaces 94b, 95b in the fourth cylinder # 4. Thus, the squish part 96a of the second cylinder # 2 is formed larger than the squish part 96b of the fourth cylinder # 4. Therefore, since the squish rate is calculated by the following formula, the squish rate of the second cylinder # 2 where the overlap period overlaps is set higher than that of the fourth cylinder # 4 where the overlap period does not overlap. Will be.
Squish ratio = surface area of squish portion / surface area of piston top surface In the squish portions 96a and 96b, when the piston 17 is located at the top dead center, the clearance between the top surface of the piston 17 and the bottom surface of the cylinder head 21 is remarkably large. Although it is a small region (for example, about 1 mm), although not shown, a squish portion inclined to face the squish portions 96a and 96b is also provided on the cylinder head 21 side. Further, the valve recess is a recessed portion that is punched out so as not to contact the valve when the piston is located at the top dead center.

  Therefore, in the first bank # 1 and the third cylinder # 3 in the left bank 12, the internal EGR amount is affected by the exhaust pulsation of the seventh cylinder # 7 and the fifth cylinder # 5 during the overlap period. To increase. On the other hand, in the fifth cylinder # 5 and the seventh cylinder # 7, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, since the squish rates of the first cylinder # 1 and the third cylinder # 3 are set higher than the squish rates of the fifth cylinder # 5 and the seventh cylinder # 7, the first cylinder # having a large internal EGR amount. In the first and third cylinders # 3, the turbulence of the air-fuel mixture in the combustion chamber 23 is increased by the squish, and the combustion speeds of the cylinders # 1 and # 3 are the fifth cylinder # 5 and the small internal EGR amount The combustion speed increases to the same level as the combustion speed of the seventh cylinder # 7. As a result, the combustion states of the first cylinder # 1 and the third cylinder # 3 having a large internal EGR amount and the combustion states of the fifth cylinder # 5 and the seventh cylinder # 7 having a small internal EGR amount are approximated. The combustion state in each of the 12 cylinders is almost uniform.

  Similarly, in the right bank 13, in the second cylinder # 2 and the sixth cylinder # 6, the internal EGR amount is affected by the exhaust pulsation of the eighth cylinder # 8 and the fourth cylinder # 4 during the overlap period. Quantitative increase. On the other hand, in the fourth cylinder # 4 and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, since the squish rates of the second cylinder # 2 and the sixth cylinder # 6 are set higher than the squish rates of the fourth cylinder # 4 and the eighth cylinder # 8, the second cylinder # having a large internal EGR amount. In the second and sixth cylinders # 6, the turbulence of the air-fuel mixture in the combustion chamber 23 is increased by the squish, and the combustion speeds of the cylinders # 2 and # 6 have a small internal EGR amount and the fourth cylinder # 4 and The combustion speed increases to the same level as the combustion speed of the eighth cylinder # 8. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the combustion states of the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated. The combustion state in each of the 12 cylinders is almost uniform.

  Thus, in the internal combustion engine of the third embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. The squish rate in the sixth cylinder # 6 is higher than the squish rate in the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set to be.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 that do not overlap, the internal EGR amount does not increase during this period without being affected by exhaust pulsation, but the squish of cylinders # 1, # 2, # 3, and # 6 The ratio is set to be higher than the squish ratio of cylinders # 4, # 5, # 7, and # 8, and mixing in the combustion chamber 23 of cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount is performed. The turbulence is enhanced, and the combustion speed can be increased to the same level as that of the cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  FIG. 8 is a plan view of a combustion chamber in a V-type 8-cylinder engine representing an internal combustion engine according to Embodiment 4 of the present invention, and FIG. 9 is a cross-sectional view showing a squish rate variable mechanism taken along the line IX-IX in FIG. It is. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  In the V-type 8-cylinder engine of the fourth embodiment, as shown in FIGS. 8 and 9, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8, and the fifth cylinder # as specific cylinders. 5. The squish rate of the first cylinder # 1, the third cylinder # 3, the second cylinder # 2, and the sixth cylinder # 6 as the first cylinders whose overlap period overlaps with the exhaust stroke of the fourth cylinder # 4 ( Combustion parameters) and squish ratios of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinders in which the overlap period does not overlap with the exhaust stroke. Changes (combustion parameter changing means) can be made to be different. In the fourth embodiment, as the combustion parameter changing means, the squish rate variable mechanism 101 is provided in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6. , # 2, # 3, # 6 can be changed to be higher than the squish ratio of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, eighth cylinder # 8, And the amount of change is set according to the length of the overlap period.

  Specifically, in this embodiment, the squish rate variable mechanism 101 is provided in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps with the exhaust stroke of a specific cylinder. For example, in the squish rate variable mechanism 101 provided in the second cylinder # 2 of the right bank 13, the squish portion 102 is provided along the circumferential direction on the outer peripheral portion of the lower surface of the cylinder head 21 that constitutes the ceiling portion of the combustion chamber 23. Is formed. A slide ring 103 having a ring shape located inside the squish portion 102 is supported on the lower surface of the cylinder head 21 so as to be movable up and down, and can be moved up and down by a driving device 104. A squish portion 105 is formed on the outer peripheral portion of the top surface of the piston 17 so as to face the squish portion 102 and the slide ring 103.

  In this case, in the second cylinder # 2 of the right bank 13, the overlap period is set according to the operation state by the intake variable valve mechanism 41 and the exhaust variable valve mechanism 43, but when there is almost no overlap period, The slide ring 103 is moved upward by the driving device 104 of the squish rate variable mechanism 101 to move away from the squish portion 105, and the squish rate of the second cylinder # 2 in which the overlap period overlaps with the exhaust stroke of the specific cylinder is The squish rate is the same as that of cylinders # 4, # 5, # 7, and # 8 where the overlap period does not overlap with the exhaust stroke of the cylinder. Further, when the overlap period is set according to the operating state by the intake variable valve mechanism 41 and the exhaust variable valve mechanism 42, the slide ring 103 is moved downward by the drive device 104 of the squish rate variable mechanism 101 and squished. The squish rate of the second cylinder # 2, which overlaps with the exhaust stroke of the specific cylinder and overlaps with the exhaust stroke of the specific cylinder, is set to the cylinder # 4, # 5, # 7 where the overlap period does not overlap with the exhaust stroke of the specific cylinder. , Set higher than the squish rate of # 8.

  Accordingly, in each of the banks 12 and 13, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the seventh cylinder # 7 and the eighth cylinder # 8 in the overlap period. The internal EGR amount increases by a predetermined amount under the influence of the exhaust pulsation of the fifth cylinder # 5 and the fourth cylinder # 4. On the other hand, in the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, when the overlap period occurs in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the slide ring 103 is moved downward by the squish rate variable mechanism 101 and squished. By making it approach the part 105, a squish area is expanded and a squish rate becomes high. For this reason, in the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount, the disturbance of the air-fuel mixture in the combustion chambers 22 and 23 is promoted, and the cylinders # 1, # 2, # 3, and # 6 are increased. The combustion speed of No. 6 increases to the same level as the combustion speed of cylinders # 4, # 5, # 7, and # 8 with a small amount of internal EGR. As a result, the combustion states of cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount and the combustion states of cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount are approximated, The combustion state in each cylinder of each bank 12 and 13 is almost uniform.

  Thus, in the internal combustion engine of the fourth embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. When the squish rate variable mechanism 101 is provided in the sixth cylinder # 6 and the overlap period is set by the variable intake valve mechanisms 40, 41 and the variable exhaust valve mechanisms 42, 43, the cylinders # 1, # 2 , # 3 and # 4 are higher than the squish rates of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7 and the eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set as follows.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 that do not overlap, the internal EGR amount does not increase during this period and is not affected by exhaust pulsation, but overshoots cylinders # 1, # 2, # 3, and # 6. When the lap period occurs, the squish rate variable mechanism 101 moves the slide ring 103 to the lower position to enlarge the squish area, thereby changing the squish rate of the cylinders # 1, # 2, # 3, and # 6 to the cylinder #. In cylinders # 1, # 2, # 3, and # 6, which have higher squish ratios of 4, # 5, # 7, and # 8 and have a large internal EGR amount, turbulence of the air-fuel mixture in the combustion chambers 22 and 23 The combustion speed is EGR amount less cylinder # 4, # 5, # 7, can be increased to the same extent as the burn rate of the # 8. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  In the fourth embodiment, the squish rate variable mechanism 101 can be moved up and down by the drive device 104. However, like the compression ratio variable mechanism 81 in the second embodiment, the squish rate variable mechanism 101 can be moved up and down by hydraulic pressure or in-cylinder pressure. good. In addition, by using an electric motor or the like as the driving device 104 of the squish rate variable mechanism 101, it is continuously opened according to the overlap period set by the variable intake valve mechanisms 40, 41 and the variable exhaust valve mechanisms 42, 43. You may make it adjust to.

  Furthermore, in the third embodiment, the squish portion on the pistons 16 and 17 side is changed, and in the fourth embodiment, the squish portion on the cylinder heads 20 and 21 side is changed. Both may be changed.

  10-1 is a cross-sectional view of a combustion chamber of a cylinder having a large internal EGR in a V-type 8-cylinder engine representing an internal combustion engine according to Embodiment 5 of the present invention, and FIG. 10-2 is a V-type 8-cylinder of Embodiment 3. It is sectional drawing of the combustion chamber of a cylinder with few internal EGR in an engine. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  In the V-type 8-cylinder engine of the fifth embodiment, as shown in FIGS. 10-1 and 10-2, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8 as specific cylinders, 1st cylinder # 1, 3rd cylinder # 3, 2nd cylinder # 2, 6th cylinder # 6 as the 1st cylinder where the overlap period overlaps with the exhaust stroke of 5th cylinder # 5 and 4th cylinder # 4 Of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinder in which the overlap period does not overlap the exhaust stroke. It can be changed (combustion parameter changing means) so that the tumble ratio is different. In the first embodiment, as the combustion parameter changing means, the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 are provided by the pistons 16 and 17 of the banks 12 and 13, respectively. The intake port shape of each cylinder is changed so that the tumble ratio becomes higher than the tumble ratios of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8.

More specifically, for example, in the second bank # 2 having a large internal EGR amount due to overlapping overlap periods in the right bank 13, as shown in FIG. 10-1, the position of the intake port 25 is set low. In the fourth cylinder # 4 where the overlap period does not overlap and the amount of internal EGR is small, the position of the intake port 25 is set high as shown in FIG. 10-2. Therefore, in the second cylinder # 2, the intake path from the intake port 25 to the combustion chamber 23 is a gentle straight line, and the inclination angle of the intake air introduced into the combustion chamber 23 is small, while in the fourth cylinder # 4 The intake path from the intake port 25 to the combustion chamber 23 is gently bent, and the inclination angle of the intake air introduced into the combustion chamber 23 is large and almost upright. Therefore, since the tumble ratio is calculated by the following formula, the tumble ratio of the second cylinder # 2 where the overlap period overlaps is set higher than the tumble ratio of the fourth cylinder # 4 where the overlap period does not overlap. Will be.
Tumble ratio = Number of swirling of airflow in combustion chamber / one rotation of crank Note that the tumble is generated in the combustion chamber 23 by the air flow sucked into the combustion chamber 23 from the intake port 25 in the intake stroke and the compression stroke. Is a swirl flow.

  Therefore, in the first bank # 1 and the third cylinder # 3 in the left bank 12, the internal EGR amount is affected by the exhaust pulsation of the seventh cylinder # 7 and the fifth cylinder # 5 during the overlap period. To increase. On the other hand, in the fifth cylinder # 5 and the seventh cylinder # 7, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, since the tumble ratio of the first cylinder # 1 and the third cylinder # 3 is set higher than the tumble ratio of the fifth cylinder # 5 and the seventh cylinder # 7, the first cylinder # having a large internal EGR amount is set. In the first and third cylinders # 3, the turbulence of the air-fuel mixture in the combustion chamber 23 is promoted by tumble, and the combustion speeds of the cylinders # 1 and # 3 are the fifth cylinder # 5 and the smaller internal EGR amount. The combustion speed increases to the same level as the combustion speed of the seventh cylinder # 7. As a result, the combustion states of the first cylinder # 1 and the third cylinder # 3 having a large internal EGR amount and the combustion states of the fifth cylinder # 5 and the seventh cylinder # 7 having a small internal EGR amount are approximated. The combustion state in each of the 12 cylinders is almost uniform.

  Similarly, in the right bank 13, in the second cylinder # 2 and the sixth cylinder # 6, the internal EGR amount is affected by the exhaust pulsation of the eighth cylinder # 8 and the fourth cylinder # 4 during the overlap period. Quantitative increase. On the other hand, in the fourth cylinder # 4 and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, since the tumble ratio of the second cylinder # 2 and the sixth cylinder # 6 is set higher than the tumble ratio of the fourth cylinder # 4 and the eighth cylinder # 8, the second cylinder # having a large internal EGR amount. In the second and sixth cylinders # 6, the turbulence of the air-fuel mixture in the combustion chamber 23 is promoted by tumble, and the combustion speeds of the cylinders # 2 and # 6 have a small internal EGR amount. The combustion speed increases to the same level as the combustion speed of the eighth cylinder # 8. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the combustion states of the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated. The combustion state in each of the 12 cylinders is almost uniform.

  Thus, in the internal combustion engine of the fifth embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. The tumble ratio in the sixth cylinder # 6 is higher than the tumble ratio in the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set to be.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 where the cylinders do not overlap, the internal EGR amount does not increase during this period without being affected by exhaust pulsation, but the cylinders # 1, # 2, # 3, and # 6 are tumbled The ratio is set higher than the tumble ratio of the cylinders # 4, # 5, # 7, and # 8, and the mixing in the combustion chamber 23 of the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount is set. The turbulence is enhanced, and the combustion speed can be increased to the same level as that of the cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  FIG. 11 is a cross-sectional view of a combustion chamber in a V-type 8-cylinder engine that represents an internal combustion engine according to Embodiment 6 of the present invention. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  In the V-type 8-cylinder engine of the sixth embodiment, as shown in FIG. 11, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8, the fifth cylinder # 5, Tumble ratio (combustion parameter) of the first cylinder # 1, the third cylinder # 3, the second cylinder # 2, and the sixth cylinder # 6 as the first cylinder in which the overlap period overlaps with the exhaust stroke of the four cylinder # 4 And the tumble ratio of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinders whose overlap periods do not overlap with each other in the exhaust stroke. (Combustion parameter changing means) is possible. In the fifth embodiment, as the combustion parameter changing means, the tumble ratio variable mechanism 111 is provided in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6. , # 2, # 3, # 6 can be changed so that the tumble ratio of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, eighth cylinder # 8 is higher than the tumble ratio, And the amount of change is set according to the length of the overlap period.

  Specifically, in the present embodiment, the tumble ratio variable mechanism 111 is provided in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps with the exhaust stroke of a specific cylinder. For example, in the tumble ratio variable mechanism 111 provided in the second cylinder # 2 of the right bank 13, a partition wall 113 is provided in the intake port 25 along the longitudinal direction, and the inside is provided with the upper flow path 112a and the lower side. The flow path 112b is partitioned so that the flow area of the lower flow path 112b is larger than the flow area of the upper flow path 112a. Further, on the upstream side of the partition wall 113 in the intake port 25, a substantially central portion of the control plate 114 is rotatably supported by a support shaft 115, and can be rotated by a driving device (not shown). The control plate 114 has an opening 116 corresponding to the lower flow path 112b so that the upper flow path 112a can be opened and closed.

  In this case, in the second cylinder # 2 of the right bank 13, the overlap period is set according to the operation state by the intake variable valve mechanism 41 and the exhaust variable valve mechanism 43, but when there is almost no overlap period, The control plate 114 is rotated by the tumble ratio variable mechanism 111 to the position shown by the solid line in FIG. 11 to close only the upstream flow path 112a, and the intake air is introduced from the opening 116 into the combustion chamber 23 through the downstream flow path 112b. . Therefore, the cylinder # 4, # 5, # 7, # in which the overlap period does not overlap the exhaust stroke of the specific cylinder is lowered by lowering the tumble ratio of the second cylinder # 2 where the overlap period overlaps the exhaust stroke of the specific cylinder. Same as tumble ratio of 8. When the overlap period is set according to the operating state by the intake variable valve mechanism 41 and the exhaust variable valve mechanism 42, the control plate 114 is moved by the tumble ratio variable mechanism 111 to the position indicated by a two-dotted line in FIG. And the upstream flow path 112a and the downstream flow path 112b are opened, and the intake air is introduced from the opening 116 into the combustion chamber 23 through the respective side flow paths 112a and 112b. Therefore, the tumble ratio of the second cylinder # 2 in which the overlap period overlaps with the exhaust stroke of the specific cylinder is increased, and the tumble ratio of the second cylinder # 2 in which the overlap period overlaps with the exhaust stroke of the specific cylinder is specified It is set higher than the tumble ratio of cylinders # 4, # 5, # 7, and # 8 where the overlap period does not overlap with the exhaust stroke of the cylinder.

  Accordingly, in each of the banks 12 and 13, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the seventh cylinder # 7 and the eighth cylinder # 8 in the overlap period. The internal EGR amount increases by a predetermined amount under the influence of the exhaust pulsation of the fifth cylinder # 5 and the fourth cylinder # 4. On the other hand, in the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, when an overlap period occurs in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the control plate 114 is rotated by the tumble ratio variable mechanism 111 to the downstream side. By opening only the flow path 112b, the tumble in the combustion chambers 22 and 23 is easily generated, and the tumble ratio is increased. Therefore, in the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount, the turbulence of the air-fuel mixture in the combustion chambers 22 and 23 is increased by tumble, and the cylinders # 1, # 2, and # 3 The combustion speed of # 6 increases to the same level as the combustion speed of cylinders # 4, # 5, # 7, and # 8 with a small internal EGR amount. As a result, the combustion states of cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount and the combustion states of cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount are approximated, The combustion state in each cylinder of each bank 12 and 13 is almost uniform.

  As described above, in the internal combustion engine of the sixth embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. When the tumble ratio variable mechanism 111 is provided in the sixth cylinder # 6 and the overlap period is set by the variable intake valve mechanisms 40, 41 and the variable exhaust valve mechanisms 42, 43, the cylinders # 1, # 2 , # 3 and # 4, the tumble ratio is higher than the tumble ratio of the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set as follows.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 that do not overlap, the internal EGR amount does not increase during this period and is not affected by exhaust pulsation, but overshoots cylinders # 1, # 2, # 3, and # 6. When the lap period occurs, the control plate 114 is rotated by the tumble ratio variable mechanism 111 to open only the downstream flow path 112b. In the cylinders # 1, # 2, # 3, and # 6, which are higher than the tumble ratios of the cylinders # 4, # 5, # 7, and # 8 and have a large internal EGR amount, the air-fuel mixture in the combustion chambers 22 and 23 Turbulence increases, and the combustion speed is the amount of internal EGR. No cylinder # 4, # 5, # 7, it can be increased to the same extent as the burn rate of the # 8. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  In the fifth and sixth embodiments, the tumble ratio is changed depending on the shape of the intake ports 24 and 25. However, the present invention is not limited to this structure. For example, a tumble flow introduction port is provided separately from the intake port. It may be possible to open and close. In addition, by using an electric motor or the like as a drive device for the control plate 114, the variable intake valve mechanisms 40, 41 and the variable exhaust valve mechanisms 42, 43 are adjusted to be continuously open according to the overlap period set. You may do it.

  FIG. 12 is a front view of a spark plug in a V-type 8-cylinder engine that represents an internal combustion engine according to a seventh embodiment of the present invention. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  Since the overall configuration of the V-type 8-cylinder engine of Embodiment 7 is the same as that of Embodiment 1 described above, it will be described with reference to FIGS. 1 and 2. In the V-type 8-cylinder engine of the seventh embodiment, as shown in FIGS. 1 and 2, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8, and the fifth cylinder # as specific cylinders. 5. Ignition energy of the first cylinder # 1, the third cylinder # 3, the second cylinder # 2, and the sixth cylinder # 6 as the first cylinder in which the overlap period overlaps the exhaust stroke of the fourth cylinder # 4 ( Combustion parameters) and ignition energy of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinder in which the overlap period does not overlap with the exhaust stroke. Changes (combustion parameter changing means) can be made to be different. In the first embodiment, ignition plug ignition in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 is performed in each of the banks 12 and 13 as the combustion parameter changing means. The energy is changed to be higher than the ignition energy of the spark plugs in the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8.

  More specifically, as shown in FIG. 12, the spark plugs 66 and 67 have a center electrode 121 mounted in the through hole at the tip, and the center electrode 121 and an insulator (not shown) at the tip. The ground electrode 122 is mounted at the opposite position, and a discharge gap P (Pa, Pb) is provided between the center electrode 121 and the ground electrode 122. Accordingly, by generating a spark discharge between the center electrode 121 and the ground electrode 122 by the ignition device, the air-fuel mixture flowing in the combustion chambers 22 and 23 can be ignited. In the present embodiment, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 that have a large internal EGR amount due to overlapping overlap periods, the center electrode 121 and the ground electrode 122 In the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 in which the discharge gap Pa is set large and the overlap period is not overlapped and the internal EGR amount is small, the center electrode 121 is used. The discharge gap Pb between the ground electrode 122 and the ground electrode 122 is set small. Therefore, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, since the discharge gap Pa is large, the ignition area is expanded and the flame temperature is increased. However, it is set higher than the ignition energy of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 where the overlap periods do not overlap.

  Accordingly, in each of the banks 12 and 13, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the seventh cylinder # 7 and the eighth cylinder # 8 in the overlap period. The internal EGR amount increases by a predetermined amount under the influence of the exhaust pulsation of the fifth cylinder # 5 and the fourth cylinder # 4. On the other hand, in the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, the ignition energy of the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 is the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder. Since it is set higher than the ignition energy of # 8, in the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount, the flame temperature and propagation speed in the combustion chambers 22 and 23 are increased by the large flame energy. As a result, the combustion speed of the cylinders # 1, # 2, # 3, and # 6 increases to the same level as the combustion speeds of the cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount. As a result, the combustion states of cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount and the combustion states of cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount are approximated, The combustion state in each cylinder of each bank 12 and 13 is almost uniform.

  Thus, in the internal combustion engine of the seventh embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. The ignition energy in the sixth cylinder # 6 is higher than the ignition energy in the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. It is set to be.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 that do not overlap, the internal EGR amount does not increase during this period without being affected by exhaust pulsation, but ignition of cylinders # 1, # 2, # 3, and # 6 The energy is set higher than the ignition energy of the cylinders # 4, # 5, # 7, and # 8, and the flames in the combustion chambers 23 of the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount The temperature and propagation speed are increased, and the combustion speed can be increased to the same level as the combustion speeds of cylinders # 4, # 5, # 7, and # 8 with a small amount of internal EGR. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  In the seventh embodiment, the magnitude of the discharge gap P of the spark plugs 66 and 67 is changed as the one that changes the ignition energy of the cylinder. However, the ignition coil may be changed to a different one. good.

  FIG. 13 is a schematic plan view of a V-type 8-cylinder engine representing an internal combustion engine according to Embodiment 8 of the present invention, and FIG. 14 is a flowchart of ignition control in the V-type 8-cylinder engine of Embodiment 8. In addition, the same code | symbol is attached | subjected to the member which has the same function as what was demonstrated in the Example mentioned above, and the overlapping description is abbreviate | omitted.

  In the V-type 8-cylinder engine of the eighth embodiment, as shown in FIG. 13, in each of the banks 12 and 13, the seventh cylinder # 7, the eighth cylinder # 8, the fifth cylinder # 5, Ignition energy (combustion parameter) of the first cylinder # 1, the third cylinder # 3, the second cylinder # 2, and the sixth cylinder # 6 as the first cylinder in which the overlap period overlaps with the exhaust stroke of the four cylinder # 4 And the ignition energy of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 as the second cylinders whose overlap periods do not overlap with each other in this exhaust stroke. (Combustion parameter changing means) is possible. In the eighth embodiment, the number of spark plugs in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 is set in each bank 12 and 13 as the combustion parameter changing means. However, the number of spark plugs is changed to be larger than the number of spark plugs of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8.

  More specifically, in each of the banks 12 and 13, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 that have a large internal EGR amount due to overlapping overlap periods. In the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8, in which the two spark plugs 66 and 67 are mounted and the overlap period is not overlapped and the internal EGR amount is small. One spark plug 66, 67 is attached. Therefore, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, since there are two ignition parts (ignition plugs 66 and 67), the ignition area is expanded and the flame temperature is increased. The ignition energy is set higher than the ignition energy of the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 where the overlap periods do not overlap. The Rukoto.

  Accordingly, in each of the banks 12 and 13, in the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6, the seventh cylinder # 7 and the eighth cylinder # 8 in the overlap period. The internal EGR amount increases by a predetermined amount under the influence of the exhaust pulsation of the fifth cylinder # 5 and the fourth cylinder # 4. On the other hand, in the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8, the internal EGR amount does not increase without being affected by the exhaust pulsation of other cylinders during the overlap period. However, the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 are ignited by the two spark plugs 66 and 67, so that the ignition energy is the fourth cylinder # 4, Since the ignition energy of the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8 is set higher than that of the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount, a large flame is generated. The energy increases the flame temperature and the propagation speed in the combustion chambers 22 and 23, and the combustion speeds of the cylinders # 1, # 2, # 3, and # 6 are the cylinders # 4, # 5, # having a small internal EGR amount. 7. Increases to the same rate as # 8. As a result, the combustion states of cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount and the combustion states of cylinders # 4, # 5, # 7, and # 8 having a small internal EGR amount are approximated, The combustion state in each cylinder of each bank 12 and 13 is almost uniform.

  Here, ignition control in the V-type 8-cylinder engine of this embodiment will be described. As shown in FIG. 14, in step S <b> 1, it is determined based on the engine speed, for example, whether the engine operating state is an idle operating state. In step S1, if the engine is not in an idling state, two spark plugs are provided in first cylinder # 1, second cylinder # 2, third cylinder # 3, and sixth cylinder # 6 in step S2. Ignition is performed by 66 and 67, and ignition is performed by one spark plug 66 and 67 in the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8. On the other hand, if the engine is in an idle operation state in step S1, one ignition is performed in first cylinder # 1, second cylinder # 2, third cylinder # 3, and sixth cylinder # 6 in step S3. Ignition is performed by the plugs 66 and 67.

  That is, in idle operation of the engine, exhaust gas is blown back to the intake ports 24 and 25 or the combustion chambers 22 and 23 by eliminating the overlap period by the variable intake valve mechanisms 40 and 41 and the variable exhaust valve mechanisms 42 and 43. The amount is reduced, combustion stability and fuel efficiency are improved, and the first cylinder # 1, the second cylinder # 2, the third cylinder # 3, and the sixth cylinder # 6 are not affected by the exhaust pulsation of other cylinders. . Therefore, at this time, ignition is performed with one spark plug 66, 67, similarly to the fourth cylinder # 4, the fifth cylinder # 5, the seventh cylinder # 7, and the eighth cylinder # 8.

  As described above, in the internal combustion engine of the eighth embodiment, in the V-type 8-cylinder engine, the first cylinder # 1, the second cylinder # 2, and the third cylinder # in which the overlap period overlaps with the exhaust stroke of a specific cylinder. 3. The number of spark plugs 66, 67 in the sixth cylinder # 6 is the same as that in the fourth cylinder # 4, fifth cylinder # 5, seventh cylinder # 7, and eighth cylinder # 8 where the overlap period does not overlap the exhaust stroke. The number is set to be larger than the number of spark plugs 66 and 67.

  Therefore, in the cylinders # 1, # 2, # 3, and # 6 where the overlap period overlaps the exhaust stroke, the internal EGR amount increases by a predetermined amount due to the influence of exhaust pulsation during this period, while the overlap period overlaps the exhaust stroke. In cylinders # 4, # 5, # 7, and # 8 that do not overlap, the internal EGR amount does not increase during this period without being affected by exhaust pulsation, but ignition of cylinders # 1, # 2, # 3, and # 6 The energy is set higher than the ignition energy of the cylinders # 4, # 5, # 7, and # 8, and the flames in the combustion chambers 23 of the cylinders # 1, # 2, # 3, and # 6 having a large internal EGR amount The temperature and propagation speed are increased, and the combustion speed can be increased to the same level as the combustion speeds of cylinders # 4, # 5, # 7, and # 8 with a small amount of internal EGR. As a result, the combustion states of the second cylinder # 2 and the sixth cylinder # 6 having a large internal EGR amount and the fourth cylinder # 4 and the eighth cylinder # 8 having a small internal EGR amount are approximated, and combustion between the cylinders is approximated. Variations can be suppressed, fluctuations in output torque can be suppressed, and fuel consumption improvement and exhaust gas performance improvement due to overlap can be appropriately achieved.

  In each of the above-described embodiments, the firing order of each cylinder is changed to the first cylinder # 1, the eighth cylinder # 8, the seventh cylinder # 7, the third cylinder # 3, the sixth cylinder # 6, and the fifth cylinder #. 5. The fourth cylinder # 4 and the second cylinder # 2 are not limited to this order. The cylinders in the banks 12 and 13 may ignite and explode at unequal intervals. It is. In each of the above-described embodiments, the cylinder injection internal combustion engine that directly injects fuel into the combustion chamber is used. However, a port injection internal combustion engine that injects fuel into the intake system may be used. Furthermore, although the engine of the present invention has been described as a V-type 8-cylinder engine, the number of cylinders is not limited to this.

  As described above, the internal combustion engine according to the present invention changes the combustion parameter resulting from the combustion state so that the combustion state of the cylinder where the exhaust pulsation acts is the same as the combustion state of the cylinder where the exhaust pulsation does not act. This is useful for unequal-interval ignition / explosion type internal combustion engines.

1 is a schematic plan view of a V-type 8-cylinder engine that represents an internal combustion engine according to Embodiment 1 of the present invention. 1 is a schematic configuration diagram of a V-type 8-cylinder engine according to Embodiment 1. FIG. 2 is a cross-sectional view of a combustion chamber of a cylinder having a large internal EGR in the V-type 8-cylinder engine of Embodiment 1. FIG. FIG. 3 is a cross-sectional view of a combustion chamber of a cylinder having a small internal EGR in the V-type 8-cylinder engine of the first embodiment. 3 is a time chart showing the opening timing of the intake valve and the exhaust valve in the V-type 8-cylinder engine of the first embodiment. It is a top view of the combustion chamber in the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 2 of this invention. FIG. 6 is a cross-sectional view taken along the line VI-VI in FIG. 5 and illustrating a variable compression ratio mechanism. It is sectional drawing showing the operation state of a compression ratio variable mechanism. It is a top view of the combustion chamber of a cylinder with much internal EGR in the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 3 of this invention. FIG. 6 is a plan view of a combustion chamber of a cylinder having a small internal EGR in the V-type 8-cylinder engine of Example 3. It is a top view of the combustion chamber in the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 4 of this invention. It is IX-IX section of Drawing 8, and is a sectional view showing a squish rate variable mechanism. It is sectional drawing of the combustion chamber of a cylinder with much internal EGR in the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 5 of this invention. FIG. 6 is a cross-sectional view of a combustion chamber of a cylinder having a low internal EGR in the V-type 8-cylinder engine of Example 3. It is sectional drawing of the combustion chamber in the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 6 of this invention. It is a front view of the spark plug in the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 7 of this invention. It is a schematic plan view of the V type 8 cylinder engine showing the internal combustion engine which concerns on Example 8 of this invention. 10 is a flowchart of ignition control in a V-type 8-cylinder engine according to an eighth embodiment.

Explanation of symbols

12, 13 Bank 16, 17 Piston 22, 23 Combustion chamber 24, 25 Intake port 26, 27 Exhaust port (combustion parameter changing means)
28, 29 Intake valve 30, 31 Exhaust valve 36, 37 Intake cam 38, 39 Exhaust cam 40, 41 Intake variable valve mechanism 42, 43 Exhaust variable valve mechanism 54, 55 Exhaust manifold 61, 62 Injector 66, 67 Ignition plug (Combustion parameter changing means)
68 Electronic control unit, ECU
80a, 80b cavity (combustion parameter changing means)
81 Compression ratio variable mechanism (combustion parameter changing means)
82 Slide ring 96a, 96b Squish part (combustion parameter changing means)
101 Squish rate variable mechanism (combustion parameter changing means)
103 Slide ring 111 Tumble ratio variable mechanism (combustion parameter changing means)
113 Partition wall 114 Control plate 121 Center electrode 122 Ground electrode Pa, Pb Discharge gap (combustion parameter changing means)
# 1, # 2, # 3, # 6 cylinder (first cylinder)
# 4, # 5, # 7, # 8 cylinder (second cylinder)

Claims (7)

  A plurality of cylinders are divided into left and right banks, and the cylinders in each bank are ignited at unequal intervals, and the exhaust valve opening period and the intake valve opening period in each cylinder overlap each other. In the internal combustion engine, in each bank, combustion between a first cylinder in which the overlap period overlaps with an exhaust stroke of a specific cylinder and a second cylinder in which the overlap period does not overlap with an exhaust stroke of the specific cylinder An internal combustion engine comprising combustion parameter changing means for changing a combustion parameter caused by a state.   2. The internal combustion engine according to claim 1, wherein the combustion parameter changing means changes a compression ratio of the first cylinder or the second cylinder as the combustion parameter.   2. The internal combustion engine according to claim 1, wherein the combustion parameter changing means changes a squish rate of the first cylinder or the second cylinder as the combustion parameter.   2. The internal combustion engine according to claim 1, wherein the combustion parameter changing means changes a tumble ratio of the first cylinder or the second cylinder as the combustion parameter.   2. The internal combustion engine according to claim 1, wherein the combustion parameter changing means changes ignition energy of the first cylinder or the second cylinder as the combustion parameter.   6. The internal combustion engine according to claim 5, wherein the combustion parameter changing means does not change the ignition energy in an idle state of the internal combustion engine.   6. The internal combustion engine according to claim 1, wherein the combustion parameter changing unit changes the compression ratio, the squish rate, the tumble ratio, or ignition energy in accordance with the overlap period. A characteristic internal combustion engine.

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