JP5673214B2 - Intake and exhaust system for multi-cylinder engine - Google Patents

Description

  The present invention relates to an intake / exhaust device for a multi-cylinder engine provided in an automobile or the like.

  2. Description of the Related Art Conventionally, intake and exhaust devices have been developed for the purpose of increasing engine output in engines such as automobiles.

  For example, Patent Document 1 is a device having a turbocharger, which is connected to an exhaust port of each cylinder and independent from each other, and an exhaust passage provided upstream of the turbocharger. Are provided, and a valve provided in the collecting portion and capable of changing the flow area of each exhaust passage is disclosed. In this apparatus, by reducing the flow passage area of the exhaust passage by the valve, the exhaust of the cylinder in the exhaust stroke flows from the predetermined exhaust passage into the collecting portion at a relatively high speed, and the periphery of the high-speed exhaust is The amount of gas supplied to the turbocharger is reduced by causing the negative pressure generated in the above-mentioned collecting portion to act on the other exhaust passage and sucking the exhaust in the other exhaust passage downstream by the so-called ejector effect. The engine output is increased to increase the engine output.

JP 2009-97335 A

  In the apparatus in which the flow passage area of the exhaust passage is reduced so as to obtain the intake amount increasing effect by the ejector effect as in the prior art, the exhaust speed can be increased while the exhaust resistance is increased. For this reason, in such a device, the effect of increasing the intake air amount due to the ejector effect can be obtained in a region where the exhaust flow rate is low and the engine speed is low, while in the region where the exhaust flow rate is high and the engine speed is high, However, there is a problem that the engine output is reduced with the increase of the engine.

  In view of such circumstances, an object of the present invention is to provide an intake / exhaust device for a multi-cylinder engine that can increase engine output in the entire rotation range with a simple configuration.

  In order to solve the above problems, the present invention provides a plurality of cylinders each having an intake port and an exhaust port and provided with an intake valve capable of opening and closing the intake port and an exhaust valve capable of opening and closing the exhaust port. An intake / exhaust device for a multi-cylinder engine having an intake passage connected to an intake port of each of the cylinders and an independent connection connected to an exhaust port of one cylinder or a plurality of cylinders whose exhaust sequences are not consecutive to each other An exhaust passage, a collecting portion connected to a downstream end of each independent exhaust passage and collecting gas passing through each independent exhaust passage, and valve driving means capable of driving the intake valve and the exhaust valve of each cylinder The independent exhaust passages connected to the cylinders in which the exhaust order is continuous among the independent exhaust passages are connected to the collecting portion at positions adjacent to each other. And downstream end portions of the independent exhaust passages are connected to other independent exhaust passages adjacent to each other by an ejector effect as exhaust is discharged from the cylinders through the exhaust ports and the independent exhaust passages to the collective portion. In order to generate a negative pressure in the exhaust port connected to the independent exhaust passage, the downstream side has a shape with a smaller flow path area than the upstream side. In a low speed region where the number is lower than a preset reference rotational speed, the opening period of the intake valve and the opening period of the exhaust valve of each cylinder overlap with each other by a predetermined overlap period, and the exhaust sequence is continuous. Driving the intake valve and the exhaust valve of each cylinder so that the overlap period of one cylinder overlaps with the timing when the exhaust valve of the other cylinder is opened between the cylinders, When the engine speed is higher than the reference speed, the quasi-rotational speed is such that the engine output from the multi-cylinder engine is smaller than the engine power when the flow area of each independent exhaust passage is constant. Each intake passage is provided in the upstream portion thereof, and the negative pressure wave of the intake pulsation generated in the vicinity of the intake valve as the intake valve is opened is reflected. The reversal part that becomes a positive pressure wave has a length and a cross-sectional area in the intake passage from the intake valve to the reversal part, and the positive pressure wave generated when the negative pressure wave is reflected once at the reversal part. The engine rotational speed at which the pressure wave reaches the vicinity of the intake valve and the closing timing of the intake valve are substantially the same, and the primary synchronized rotational speed of the intake pulsation that provides the inertial supercharging effect of the intake air Above The timing at which the positive pressure wave generated when the negative pressure wave is reversed twice at the reversal part and near the intake valve is substantially the same as the intake valve closing timing. And the secondary tuned rotational speed of the intake pulsation for obtaining the inertial supercharging effect of the intake air becomes lower than the reference rotational speed, and the primary tuned rotational speed and the secondary tuned speed Among the engine speeds between the engine speeds, the negative pressure amount of the negative pressure wave of the intake pulsation that reaches the vicinity of the intake valve at the closing timing of the intake valve is the largest. Provided is an intake / exhaust device for a multi-cylinder engine, wherein the non-synchronized rotational speed at which the volumetric efficiency drop caused by the negative pressure wave is maximized is set to a dimension lower than the reference rotational speed. To do.

  According to this apparatus, the engine output can be increased by effectively utilizing the ejector effect and the inertial supercharging effect of the intake air.

  Specifically, in the present apparatus, the flow passage area of each independent exhaust passage is configured to be reduced on the downstream side to obtain an ejector effect, and during a predetermined cylinder overlap period in the low speed region. The exhaust valves of other cylinders are opened. Therefore, in the low speed region, a negative pressure can be generated in the exhaust port of the cylinder during the overlap period by the ejector effect, and scavenging during the overlap period can be promoted.

  Here, exhaust resistance increases because the flow area of the independent exhaust passage is reduced. On the other hand, in the present apparatus, the primary synchronized rotational speed at which the inertial supercharging effect of the intake air can be obtained by appropriately adjusting the length and the cross-sectional area of the intake air passage is set in the high speed region. Therefore, even in the high-speed region where the effect of increased exhaust resistance is greater than the effect of improving the scavenging performance due to the ejector effect due to the large exhaust flow rate, the scavenging performance in the cylinder is increased due to the inertial supercharging effect of the intake air, thereby reducing the pump loss. The increase can be suppressed, and the engine output can be increased in the entire rotation region.

  In particular, the engine speed at which the positive pressure wave is generated in the vicinity of the intake valve and the valve closing timing of the intake valve out of the engine speed between the primary synchronized speed and the secondary synchronized rotational speed are greatly different. As the negative pressure wave with a large negative pressure reaches the vicinity of the intake valve at the closing timing of the intake valve, the amount of deterioration in volumetric efficiency due to the negative pressure wave is the largest, and the unsynchronized rotational speed at which the volumetric efficiency is lowered is reduced. It is comprised so that it may become lower than the said reference rotation speed. Therefore, a decrease in volume efficiency due to the influence of the negative pressure wave can be compensated by the ejector effect, and a decrease in volume efficiency can be suppressed small. Therefore, the scavenging performance can be more reliably improved from the low speed region to the entire high speed region.

  Here, as a specific configuration for effectively obtaining the ejector effect, the independent exhaust passage has a diameter a of a perfect circle having the same area as the flow path area at the downstream end thereof, and a minimum of the collecting portion. Examples include a shape in which the relationship with the diameter D of a perfect circle having the same area as the flow path area is a / D ≧ 0.5 (Claim 2).

  Further, in order to further increase the flow rate of the exhaust gas, it is preferable to form a throttle portion having a reduced flow area at the downstream end portion of the independent exhaust passage. In this case, the diameter a is further set to the downstream end portion. The relationship between the diameter of a perfect circle having the same area as the flow path area of the throttle part and the diameter D of a perfect circle having the same area as the minimum flow path area of the aggregate part satisfies a / D ≧ 0.5. (Claim 3).

  Further, in the present invention, a bypass passage connected to a portion on the upstream side of the downstream end portion of each independent exhaust passage, and a gas connected to the downstream end of the bypass passage and passing through each bypass passage Is connected to the bypass passage side collecting portion and the downstream end of the collecting portion, the final collecting portion where the gas passing through the bypass passage side collecting portion and the collecting portion gathers, Provided in each bypass passage, comprising a variable flow area valve that can change the flow area of each bypass passage, and a variable flow area valve driving means that can open and close the variable flow area valve. The flow passage area of each bypass passage is the sum of the flow passage area of each bypass passage and the flow passage area at the downstream end of each independent exhaust passage. The flow path area variable valve driving means is set to have a dimension that is equal to or larger than the flow path area of the portion upstream of the portion to which the road is connected, and the required torque for at least the multi-cylinder engine in the low speed region is set. In a high load region higher than a predetermined value, the flow passage area variable valve is made smaller than fully opened to reduce the flow rate of exhaust gas passing through each bypass passage, while the engine speed is higher than the reference speed. In the region, it is preferable that the flow passage area variable valve is fully opened to open the bypass passages.

  In this way, in a high speed region where the exhaust flow rate is large, the exhaust discharged from the cylinder can be added to the independent exhaust passage and pass through the bypass passage to be discharged downstream, and the exhaust resistance can be suppressed to a low level. The output can be further increased.

  Examples of each bypass passage include those having a constant flow path area or a shape having a flow path area larger on the downstream side than on the upstream side (Claim 5).

  As described above, according to the present invention, the engine output can be increased in the entire rotation region.

1 is a schematic configuration diagram of an engine system including an intake / exhaust device for a multi-cylinder engine according to an embodiment of the present invention. It is the figure which abbreviate | omitted the high speed side channel | path, the high speed side gathering part, etc. in FIG. It is the figure which abbreviate | omitted the low speed side channel | path, the low speed side gathering part, etc. in FIG. It is a schematic longitudinal cross-sectional view of the engine system corresponding to FIG. It is the VV sectional view taken on the line of FIG. It is sectional drawing which shows the other structure of a low speed side channel | path. It is a figure for demonstrating the valve timing of an intake valve and an exhaust valve. It is the figure which showed the mode of the pressure pulsation near an intake valve. It is a figure for demonstrating the valve opening timing and valve closing timing of an intake valve and an exhaust valve in the intake / exhaust apparatus of the multicylinder engine which concerns on embodiment of this invention. It is a graph for demonstrating the setting procedure of a reference | standard rotation speed. It is a graph for demonstrating the effect of this invention.

  An embodiment of an intake / exhaust device for a multi-cylinder engine according to the present invention will be described with reference to the drawings.

  FIG. 1 is a schematic configuration diagram of an engine system 100 including an intake / exhaust device for the multi-cylinder engine. The engine system 100 includes an engine main body 1 having a cylinder head 9 and a cylinder block, an ECU 2 for engine control, a plurality of intake pipes 3 connected to the engine main body 1 and constituting a part of an intake passage, and the intake air. A surge tank 4 connected to the pipe 3 and constituting a part of the intake passage, an exhaust manifold 5 connected to the engine body 1, and a catalyst device 6 connected to the exhaust manifold 5 are provided.

  A plurality of cylinders 12 into which pistons are respectively inserted are formed in the cylinder head 9 and the cylinder block. In the present embodiment, the engine body 1 is an in-line four-cylinder engine, and four cylinders 12 are formed in series in the cylinder head 9 and the cylinder block. Specifically, a first cylinder 12a, a second cylinder 12b, a third cylinder 12c, and a fourth cylinder 12d are formed in order from the right in FIG. Each cylinder head 9 is provided with a spark plug 15 so as to face a combustion chamber partitioned above the piston.

  The engine body 1 is a four-cycle engine. As shown in FIG. 7, the cylinders 12a to 12d are ignited by the spark plug 15 at a timing shifted by 180 ° C. A, and the intake stroke and the compression stroke are performed. The expansion stroke and the exhaust stroke are each shifted by 180 ° C. A. In the present embodiment, ignition is performed in the order of the first cylinder 12a → the third cylinder 12c → the fourth cylinder 12d → the second cylinder 12b, and the exhaust stroke and the like are performed in this order.

  Two intake ports 17 and two exhaust ports 18 that open toward the combustion chamber are provided at the top of each cylinder 12. The intake port 17 is for introducing intake air into each cylinder 12. The exhaust port 18 is for exhausting the exhaust from each cylinder 12. Each intake port 17 is provided with an intake valve 19 for opening and closing the intake port 17 to communicate or block the intake port 17 and the inside of the cylinder 12. Each exhaust port 18 is provided with an exhaust valve 20 for opening and closing the exhaust port 18 to communicate or block the exhaust port 18 and the inside of the cylinder 12. The intake valve 19 is driven by an intake valve drive mechanism (valve drive means) 30 to open and close the intake port 17 at a predetermined timing. The exhaust valve 20 is driven by an exhaust valve drive mechanism (valve drive means) 40 to open and close the exhaust port 18 at a predetermined timing.

  The intake valve drive mechanism 30 has an intake camshaft 31 and an intake VVT 32 connected to the intake valve 19. The intake camshaft 31 is connected to the crankshaft via a known power transmission mechanism such as a chain / sprocket mechanism, and rotates with the rotation of the crankshaft to open and close the intake valve 19.

  The intake VVT 32 is for changing the valve timing of the intake valve 19. The intake VVT 32 is arranged coaxially with the intake camshaft 31 and changes the phase difference between a predetermined driven shaft that is directly driven by the crankshaft and the intake camshaft 31, thereby the crankshaft and the intake air By changing the phase difference from the camshaft 31, the valve timing of the intake valve 19 is changed. As a specific configuration of the intake VVT 32, for example, a plurality of liquid chambers arranged in the circumferential direction are provided between the driven shaft and the intake camshaft 31, and a pressure difference is provided between the liquid chambers to thereby change the position. A hydraulic mechanism that changes the phase difference, an electromagnetic mechanism that has an electromagnet provided between the driven shaft and the intake camshaft 31, and changes the phase difference by applying electric power to the electromagnet, etc. Is mentioned. The intake VVT 32 changes the phase difference based on the target valve timing of the intake valve 19 calculated by the ECU 2.

  The exhaust valve drive mechanism 40 has the same structure as the intake valve drive mechanism 30. That is, the exhaust valve drive mechanism 40 changes the valve timing of the exhaust valve 20 by changing the phase difference between the exhaust camshaft 41 and the crankshaft, and the exhaust camshaft 41 connected to the exhaust valve 20 and the crankshaft. And an exhaust VVT 42 to be used. The exhaust VVT 42 changes the phase difference based on the target valve timing of the exhaust valve 20 calculated by the ECU 2. The exhaust camshaft 41 rotates with the rotation of the crankshaft under this phase difference to drive the exhaust valve 20 to open and close at the target valve timing.

  In the present embodiment, the intake VVT 32 and the exhaust VVT 42 open the intake valve 19 and the exhaust valve 20 while keeping the valve opening period and the lift amount, that is, the valve profile, of the intake valve 19 and the exhaust valve 20 constant. The valve timing and the valve closing timing are each changed.

  The intake port 17 of each cylinder 12 is connected to the intake pipe 3 on the upstream side. Specifically, four intake pipes 3 are provided corresponding to the number of cylinders, and two intake ports 17 provided in each cylinder 12 are connected to one intake pipe 3.

  Each intake pipe 3 is connected to the surge tank 4 on the upstream side, and air stored in the surge tank is distributed to each intake pipe 3. The surge tank 4 extends in the direction in which the intake pipes 3 are arranged, and has a sufficient volume capable of storing air for distribution to the intake pipes 3. Accordingly, when the intake valve 19 is opened, the negative pressure wave generated in the vicinity of the intake valve 19 propagates through the intake pipe 3 and reaches the surge tank 4 and is inverted and reflected by the surge tank 4 to become a positive pressure wave. . This positive pressure wave returns to the intake valve 19 side, reverses and reflects near the inlet of the cylinder 12 to become a negative pressure wave, and propagates again to the sirgin tank 4 side. Thus, in this embodiment, the surge tank 4 functions as a reversal part where a negative pressure wave is reflected and becomes a positive pressure wave. With the reflection of the pressure wave at the surge tank 4, a pressure pulsation as shown in FIG. FIG. 8 is a diagram schematically showing a pressure change due to pressure pulsation in the vicinity of the intake valve 19 with the horizontal axis as time. As shown in FIG. 8, in the vicinity of the intake valve 19, after the opening of the intake valve 19 (time t0), peaks (high pressure state) and valleys (low pressure state) appear alternately while being attenuated. .

  The intake air inertia supercharging effect is that immediately before the intake valve 19 is closed, the vicinity of the intake valve 19 is in the state of the mountain and the pressure on the intake port 17 side is increased, and intake to the cylinder 12 is promoted. Volume efficiency is improved. This inertia supercharging effect becomes smaller as the number of times the negative pressure wave is reversed and reflected by the surge tank 4 increases. Specifically, the first peak due to the positive pressure wave generated by reversing and reflecting the negative pressure wave once by the surge tank 4 has the highest pressure, and the highest inertial supercharging effect can be obtained. . Then, the pressure of the secondary peak due to the positive pressure wave generated by reversing and reflecting the negative pressure wave twice by the surge tank 4 becomes the next higher, and the inertial supercharging effect next to the primary peak is obtained. be able to. Therefore, if the vicinity of the intake valve 19 is made the primary peak just before the intake valve 19 is closed, the volumetric efficiency can be sufficiently increased.

  On the other hand, when the vicinity of the intake valve 19 is in the valley state just before the intake valve 19 is closed, and the negative pressure wave of the intake pulsation reaches the vicinity of the intake valve 19, the pressure on the intake port 17 side becomes the negative pressure wave. Therefore, the amount of intake air introduced into the cylinder 12 can be kept small. That is, introduction of intake air into the cylinder 12 is hindered by the negative pressure wave. In particular, in the primary valley between the primary peak and the secondary peak, the amount of negative pressure that reaches the vicinity of the intake valve 19 immediately before the intake valve 19 is closed becomes the largest, so the intake intake amount deteriorates. That is, the volumetric efficiency is greatly reduced.

  Note that the pressure in the vicinity of the intake valve 19 is positive immediately before the intake valve 19 is closed, regardless of whether the pressure wave of the intake pulsation or the negative pressure wave reaches the vicinity of the intake valve 19. In such a case, the amount of positive pressure can only be kept small.

  In the present embodiment, the length and cross-sectional area of the intake pipe 3 are set so that a high inertial supercharging effect due to the primary peak can be obtained in a high speed region where the engine speed is higher than a reference speed N1 described later. Is set. Further, in this high-speed region, the intake air valve 19 is prevented from becoming a valley near the intake valve 19 immediately before the intake valve 19 is closed, so that the inertia supercharging effect can be obtained in the entire high-speed region. The length and cross-sectional area of the tube 3 are set.

  That is, the primary synchronized rotation speed Nin_1a at which the time t1a (see FIG. 8) at which the primary peak of the intake pulsation occurs in the vicinity of the intake valve 19 and the closing timing of the intake valve 19 is substantially the same is the reference rotation. The length and cross-sectional area of the intake pipe 3 are set so as to be higher than the number N1. The time t2a (see FIG. 8) at which the secondary peak of the intake pulsation occurs in the vicinity of the intake valve 19 and the secondary synchronized rotational speed Nin_2a at which the valve closing timing of the intake valve 19 becomes substantially the same is the reference rotation. The time t1b (see FIG. 8) at which the primary valley of the intake pulsation occurs in the vicinity of the intake valve 19 and the timing at which the intake valve 19 is closed is substantially the same as the first non-synchronization. The length and the cross-sectional area of the intake pipe 3 are set so that the rotational speed Nin_1b is lower than the reference rotational speed N1. Here, the primary non-synchronization rotational speed Nin_1b is specifically a rotational speed between the primary synchronous rotational speed Nin_1a and the secondary synchronous rotational speed Nin_2a, and between these synchronous rotational speeds. The engine speed at which the volume efficiency deterioration due to the influence of the negative pressure wave of the intake pulsation is the largest. For example, the non-tuned rotation speed Nin_1b has the lowest volumetric efficiency among the engine speeds between the primary tuned rotation speed Nin_1a and the secondary tuned rotation speed Nin_2a in an engine where the ejector effect cannot be obtained. The engine speed is set.

  In order to obtain a high inertial supercharging effect due to the primary peak in a higher speed region, the length of the intake pipe 3 is made shorter or the cross-sectional area of the intake pipe 3 is made larger. Although the change in which the cross-sectional area of the intake pipe 3 is increased is greatly limited in layout, in this embodiment, a high inertial supercharging effect can be obtained in a high speed region by shortening the length of the intake pipe 3. I am doing so.

  The exhaust manifold 5 includes three independent exhaust passages 52, three high-speed side passages (bypass passages) 53, three flow passage area variable valves 58, and a high-speed side collecting portion 57 (bypass passage-side collecting portion, see FIG. 3). ), A mixing tube (collecting portion) 56a, a straight tube 56b, and a diffuser 56c (see FIG. 2).

  Each independent exhaust passage 52 is connected to the exhaust port 18 of each cylinder 12. Specifically, in the cylinder 12, the exhaust port 18 of the first cylinder 12a and the exhaust port 18 of the fourth cylinder 12d are individually connected to the independent exhaust passages 52a and 52d, respectively. On the other hand, the exhaust ports 18 of the second cylinder 12b and the third cylinder 12c in which the exhaust strokes are not adjacent to each other and the exhaust order is not continuous are not exhausted simultaneously from these cylinders, so that the structure is simplified. It is connected to one independent exhaust passage 52b. More specifically, the independent exhaust passage 52b connected to the exhaust port 18 of the second cylinder 12b and the third cylinder 12c is separated into two passages on the upstream side thereof, and the second cylinder is provided in one of the two passages. The exhaust port 18 of 12b is connected, and the exhaust port 18 of the third cylinder 12c is connected to the other.

  These independent exhaust passages 52 are independent from each other, and are exhausted from the second cylinder 12b or the third cylinder 12c, exhaust exhausted from the first cylinder 12a, and exhaust exhausted from the fourth cylinder 12d. Are discharged downstream through the independent exhaust passages 52 independently of each other.

  A high speed side passage (bypass passage) 53 that bypasses the downstream side of each independent exhaust passage 52 is connected to the downstream side of each independent exhaust passage 52. That is, the downstream side of each independent exhaust passage 52 is separated into a low speed side passage 54 and a high speed side passage 53 constituting the independent exhaust passage 52. In the present embodiment, as shown in FIG. 4, the low speed side passage 54 extends downstream along the independent exhaust passage 52 upstream from the separation point from the high speed side passage 53, and the high speed side passage is provided. 53, after curving upward from the low speed side passage 54, extends to the downstream side substantially parallel to the low speed side passage 54. Further, the high speed side passage 53 and the low speed side passage 54 corresponding to the exhaust ports 18 of the second cylinder 12b and the third cylinder 12c are linearly opposed to the central portion of these cylinders, that is, the substantially central portion of the engine body 1. The high speed side passage 53 and the low speed side passage 54 corresponding to the exhaust ports 18 of the other cylinders extend from the positions facing the corresponding exhaust ports 18 to the second cylinder 12b and the third cylinder 12c. It curves and extends toward the passages 52 and 54. The cross-sectional areas, that is, the flow areas of the respective high-speed passages 53 are set to be the same, and the cross-sectional areas, that is, the flow areas of the respective low-speed passages 54 are set to be the same.

  The mixing pipe 56a is connected to the downstream side of each low speed side passage 54, and the gas that has passed through each low speed side passage 54 gathers in the mixing pipe 56a. In the mixing pipe 56a, the three low speed side passages 54 are connected at positions where their downstream ends are adjacent to each other.

  Each of the low speed side passages 54 and the mixing pipes 56a has a negative pressure generated around the high speed exhaust gas as the exhaust gas is ejected from the low speed side passages 54 at a high speed and the exhaust gas passes through the mixing pipe 56a at a high speed. Due to the pressure action, that is, the ejector effect, the adjacent low speed side passage 54 and the exhaust port 18 corresponding to the low speed side passage 54 have a negative pressure, and the gas in the exhaust port 18 is sucked downstream. doing.

  Specifically, each of the low-speed passages 54 has a shape in which the flow passage area decreases toward the downstream, and exhaust gas is ejected from each low-speed passage 54 to the downstream side at a high speed. ing. In the present embodiment, as shown in FIG. 5, each low-speed passage 54 has a cross-sectional area that is reduced from the upstream portion having a substantially circular cross section toward the downstream, and the upstream portion has a circular shape at the downstream end. It has a sector shape that is approximately 1/3 of the cross-sectional area. These low speed side passages 54 are aggregated and connected to the mixing pipe 56a so that each downstream end forming a fan shape forms a substantially circular cross section as a whole.

  The mixing pipe 56a is a true circle having the same area as the cross-sectional area of the downstream end of the low speed side passage 54, where D1 is a diameter of a perfect circle having the same area as the flow path area at the downstream end thereof (see FIG. 4). When the diameter of the circle is a (see FIG. 4), the shape is a / D = 0.65.

  Here, the specific structure of the mixing pipe 56a is not limited to the above, but the mixing pipe 56a has a shape in which the flow area of at least one of the upstream end and the downstream end is the smallest flow area, If a / D is set in the range of a / D ≧ 0.5, it is known that the exhaust passes through the mixing pipe 56a at a sufficiently high speed and the ejector effect is sufficiently obtained. What has the above shapes is preferable. In order to further increase the inflow speed of the exhaust gas to the mixing pipe 56a, as shown in FIG. 6, a portion having a reduced flow area, that is, a throttle portion 54a is provided at the downstream end of the low speed side passage 54. In this case, it is preferable that the diameter of the flow passage area of the throttle portion 54a be a, and the mixing tube 56a be shaped so that a / D ≧ 0.5. FIG. 6 is an enlarged view of a portion corresponding to the region X in FIG.

  The exhaust gas flowing into the mixing pipe 56a passes through the straight pipe 56b and the diffuser 56c and flows out downstream. The straight pipe 56c has the same cross-sectional shape as the downstream end of the mixing pipe 56a, that is, the shape extending to the downstream side with the same flow area, continuously from the mixing pipe 56a. The diffuser 56c continuously extends from the straight pipe 56b to the downstream side, and has a shape in which the diameter of the diffuser 56c increases toward the downstream and the flow passage area increases.

  On the other hand, the high-speed side collecting portion 57 is connected to the downstream side of each high-speed side passage 53, and the exhaust gas that has passed through each high-speed side passage 53 is collected in the high-speed side collecting portion 57. The high-speed side passage 53 and the high-speed side collecting portion 57 are different from the low-speed side passage 54 and the mixing pipe 56a in such a shape that the exhaust easily expands in the high-speed side passage 53 and the high-speed side collecting portion 57, that is, the exhaust It has a shape that reduces the resistance.

  Specifically, the high-speed side passage 53 has a shape that extends downstream with a substantially constant flow area, and the flow area is larger than the flow area at the downstream end of the low-speed passage 54. It has a set shape. In the present embodiment, the flow area of one high speed side passage 53 is set to substantially match the total flow area of the downstream ends of the three low speed side passages 54. The high speed side passage 53 has a total area of the flow passage area and the flow passage area at the downstream end of each low speed side passage 54 so that the exhaust gas flows more smoothly. 53 is preferably set to a dimension that is equal to or larger than the flow area of the upstream portion of the portion to which 53 is connected, but the specific shape is not limited to the above, and the flow passage area is larger than that of the upstream side. The downstream side may have a larger shape.

  The high-speed side collecting portion 57 has a substantially cylindrical shape extending along the flow direction of the exhaust gas, and has a shape in which the flow path area is substantially constant throughout.

  The flow passage area variable valve 58 is for changing the flow passage area of each high speed side passage 53 and thereby changing the flow passage area of each high speed side passage 53. Each of these flow path area variable valves 58 is provided in each high speed side passage 53.

  The flow path area variable valve 58 rotates about the rotation shaft 58a as the rotation shaft 58a provided at the center thereof is driven to rotate. In the present embodiment, a common rotation shaft 58a is fixed to each flow path area variable valve 58, and the three flow path area variable valves 58 rotate integrally. Each flow path area variable valve 58 extends in a direction substantially parallel to the exhaust flow direction and opens to the high speed side passage 53 (broken line in FIG. 4), and extends in a direction substantially perpendicular to the exhaust flow direction. It rotates between the fully closed position (solid line in FIG. 4) that blocks the passage 53, and opens and closes the high speed side passage 53 to change the flow area of the high speed side passage 53.

  The pivot shaft 58a is rotationally driven by a valve actuator (flow path area variable valve driving means) 58b provided at the end thereof. The valve actuator 58b rotates the pivot shaft 58a in accordance with the target opening degree of the flow path area variable valve 58 calculated by the ECU 2, and drives the flow path area variable valve 58 to the fully closed or fully opened position. . The valve actuator 58b may be any one as long as it can rotate the flow path area variable valve 58 by rotationally driving the rotation shaft 58a.

  The downstream end of the diffuser 56c and the downstream end of the high speed side collecting portion 57 are connected to a casing 62, which will be described later, of the catalyst device 6, and the exhaust gas that has passed through the diffuser 56c and the high speed side collecting portion 57 is placed in the casing 62. Inflow.

  The catalyst device 6 is a device for purifying the exhaust discharged from the engine body 1. The catalyst device 6 includes a catalyst body such as a three-way catalyst and a casing 62 that houses the catalyst body. The casing 62 has a substantially cylindrical shape extending in parallel with the exhaust flow direction. The catalyst body is accommodated in a central portion in the upstream and downstream direction of the casing 62, and a predetermined space is formed at the upstream end 62 a of the casing 62. The downstream ends of the diffuser 5c and the high speed side collecting portion 57 are connected to the upstream end 62a of the casing 62, and the exhaust discharged from these downstream ends is at the upstream end 62a of the casing 62 functioning as the final collecting portion. After gathering, it proceeds to the catalyst body side.

  The ECU 2 is a controller based on a known microcomputer, and includes a CPU for executing a program, a memory including a RAM and a ROM for storing a program and data, and an I / O for inputting and outputting various signals. It has a bus. The ECU 2 receives signals from various sensors via the I / O bus and performs various calculations based on the signals.

  The ECU 2 stores the target valve timings of the intake valve 19 and the exhaust valve 20 and the target opening of the flow path area variable valve 58 set in advance according to the operating conditions. The ECU 2 receives signals from various sensors. Based on this, the current operating condition is calculated and a target value corresponding to this operating condition is extracted, and the valve timing of the intake valve 19 and the exhaust valve 20 and the opening of the flow path area variable valve become this target value. The intake VVT 32, the exhaust VVT 42, and the valve actuator 58b are driven.

  Next, the target valve timing of the intake valve 19 and the exhaust valve 20 and the target opening of the flow path area variable valve 58 will be described.

  The target valve timings of the intake valve 19 and the exhaust valve 20 are the intake top dead center between the valve opening period of the exhaust valve 20 and the valve opening period of the intake valve 19 in a low speed region where the engine speed is lower than the reference speed N1. (TDC) is overlapped, and the exhaust valve 20 is set to start opening during the overlap period T_O / L of the other cylinders 12. Specifically, as shown in FIG. 7, the exhaust valve 20 of the third cylinder 12c is opened during the period in which the intake valve 19 and the exhaust valve 20 of the first cylinder 12a overlap, and the third cylinder The exhaust valve 20 of the fourth cylinder 12d opens while the intake valve 19 and exhaust valve 20 of 12c overlap, and the intake valve 19 and exhaust valve 20 of the fourth cylinder 12d overlap. The exhaust valve 20 of the second cylinder 12b is opened during the period in which the exhaust valve 20 is open, and the exhaust valve 20 of the first cylinder 12a is opened during the period in which the intake valve 19 and the exhaust valve 20 of the second cylinder 12b are overlapping. It is set to do.

  Further, the target valve timings of the intake valve 19 and the exhaust valve 20 are set such that the exhaust valve 20 is opened and the intake valve 19 is opened in the high speed region where the engine speed is higher than the reference speed N1, as in the low speed region. While the valve period is set to overlap, the overlap period T_L / O is set to be smaller than the overlap set in the low speed region. For example, the overlap period T_O / L in the low speed region is set to 60 ° C. A or more and set to 80 ° C. A, while the overlap period T_O / L in the high speed region is set to 40 ° C. or less, for example. Has been.

  The target opening degree of the flow path area variable valve 58 is set to be fully closed in the low speed load region, and is set to be fully open in the high speed region.

  In the engine system 100, the valve opening timing and the valve closing timing of the intake valve 19 and the exhaust valve 20, respectively, are as follows. As shown in FIG. It is a time of falling, for example, a time of 0.4 mm lift.

  Next, the intake performance in the engine system 100 configured as described above will be described.

  When an exhaust valve 20 of a predetermined cylinder 12 (hereinafter referred to as an exhaust stroke cylinder 12 as appropriate) is opened, exhaust is discharged from the cylinder 12 to the corresponding exhaust port 18 and the independent exhaust passage 52 at a high speed. In particular, a very high-pressure, high-speed gas (so-called blowdown gas) is discharged from the cylinder 12 immediately after the exhaust valve 20 starts to open.

  In the low speed region, the flow path area variable valve 58 is closed, and the exhaust discharged from the cylinder 12 flows only into the low speed side passage 54. As described above, the low speed side passage 54 and the mixing pipe 56a are configured so that the exhaust gas is ejected from the predetermined low speed side passage 54 to the mixing pipe 56a at a high speed due to the ejector effect. It is configured to be sucked out downstream. In the low speed region, when the exhaust valve 20 of the exhaust stroke cylinder 12 starts to be opened, another cylinder 12 whose exhaust sequence is set immediately before the exhaust stroke cylinder 12 (hereinafter referred to as an intake stroke cylinder as appropriate). 12) is set to be during the overlap period.

  Accordingly, in the low speed region, the exhaust gas in the exhaust stroke cylinder 12 flows into the low speed side passage 54 and is ejected from the low speed side passage 54 to the mixing pipe 56a at a high speed. Residual gas in the stroke cylinder 12 is sucked out to the exhaust port 18 side, scavenging of the intake stroke cylinder 12 is promoted, and intake efficiency and thus engine output is increased.

  In particular, the downstream end of each low speed side passage 54 is disposed adjacent to the mixing pipe 56a. Therefore, the suction force by the low speed side passage 54 connected to the exhaust stroke cylinder 12 effectively acts on the low speed side passage 54 connected to the intake stroke cylinder 12.

  Here, the flow passage area of the low speed side passage 54 is narrowed down on the downstream side so that the ejector effect can be effectively obtained. Therefore, when the exhaust gas passes only through the low speed side passage 54, if the engine speed increases and the exhaust gas flow rate increases, the exhaust resistance increases and the pump loss increases. As a result, the engine output deteriorates. .

  On the other hand, in the engine system 100, the reference rotational speed N1 is set near the rotational speed at which the engine output decreases as compared to the case where the passage is not throttled as the low speed side passage 54 is throttled. Has been. In other words, the reference engine speed N1 is higher in the engine system 100 when the engine output when the exhaust gas is allowed to pass through only the low speed side passage 54 (hereinafter sometimes referred to as the low speed side passage specification) in the engine system 100. When the low-speed side passage 54 and the high-speed side passage 53 are abolished and an exhaust passage having a constant flow path area is used (hereinafter, sometimes referred to as a constant flow path area specification), the engine speed is smaller than the engine output. Is set. As described above, by adjusting the length and the cross-sectional area of each intake pipe 3, the primary tuning rotation speed Nin_1a is higher than the reference rotation speed N1, and the primary non-tuning rotation speed Nin_1b is set to the reference rotation. It is set lower than the number N1. Therefore, the scavenging performance can be enhanced by the inertial supercharging effect of the intake air in the entire high speed region where the engine speed is higher than the reference speed N1 where the engine output decreases as the low speed side passage 54 is throttled, Thereby, an increase in pump loss can be suppressed and a decrease in engine output can be suppressed to a small level.

  FIG. 10 shows an example (solid line) of the engine output (engine torque) of the low-speed side passage specification (broken line) and an example (broken line) of the engine output (engine torque) of the constant passage area specification. As shown in this figure, in the region where the engine speed is lower than the torque lowering speed N0, the engine torque is higher in the low speed side passage specification than in the constant flow path area specification. On the other hand, in the region where the engine speed is higher than the torque reduction speed N0, the engine torque is lower in the low speed side passage specification than in the constant flow path area specification.

  In the present embodiment, the primary tuning rotation speed Nin_1a is set to the torque reduction rotation speed N0 so that the decrease in the engine torque is reliably suppressed, and the reference rotation speed N1 is set to the torque reduction rotation speed. The number of revolutions is set slightly smaller than N0. For example, while the torque reduction rotational speed N0 is 5000 rpm, the reference rotational speed N1 is set to 4500 rpm. In the engine body 1 with a displacement of 2 liters, the intake pipe 3 is set to have a diameter of 45 mm and a length of 500 mm so that the primary synchronized rotation speed Nin_1a5000 rpm is obtained. The dimension of 500 mm is the length from the intake valve 19 to the surge tank 4 in the intake pipe 3, and the length from the engine head to the surge tank 4 is, for example, 400 mm. In this case, the secondary tuning rotation speed Nin_2a is 2500 rpm, and the primary non-tuning rotation speed Nin_1b is 3300 rpm. FIG. 10 schematically shows a change in volume efficiency due to the intake inertia supercharging effect in the vicinity of the intake valve 19 at the closing timing of the intake valve 19 in accordance with the engine torque.

  The reference rotational speed N1 is the case where the exhaust gas is allowed to pass through only the low-speed side passage 54 in the engine system 100 when the intake pulsation effect that can be obtained with the same length and cross-sectional area of the intake pipe 3 is equivalent. The engine output of the engine system 100 is set near a rotational speed that is smaller than the engine output when the low speed side passage 54 and the high speed side passage 53 are abolished and an exhaust passage having a constant flow passage area is used in the engine system 100. Has been. Further, the engine torque comparison results of the respective specifications shown in FIG. 10 are comparisons when the intake pulsation effect that can be obtained with the same length and the cross-sectional area of the intake pipe 3 is equivalent. Accordingly, the low-speed side passage 54 and the high-speed side passage 53 are abolished and an exhaust passage having a constant flow path area is used, and an inertial supercharging effect of intake air can be obtained in a region where the engine speed is relatively low as in the prior art. The engine output of the engine system configured as described above (the configuration without the variable device for changing the synchronized rotation speed of the intake passage) is lower than that of the engine system 100 in the high speed region as shown by the chain line in FIG. In the system 100, an engine output higher than that of a conventional engine system can be obtained due to the inertial supercharging effect of the intake air.

  Further, in the engine system 100, the flow path area variable valve 58 is opened in the high speed region, and the exhaust gas discharged from the cylinder 12 has a configuration in which the exhaust resistance is reduced in addition to the low speed side passage 54. It flows into the high speed side passage 53 made. Therefore, in this high speed region, the pump loss is kept small and the engine output is increased.

  As described above, in the present engine system 100, the intake inertia effect cannot be sufficiently obtained, and further, in the low speed region where the system efficiency is deteriorated due to the negative pressure wave of the intake pulsation, the ejector effect is effectively used to make the cylinder 12 can promote scavenging, increase the volume efficiency by utilizing the intake inertia effect in the high speed region, and further reduce the pump loss by passing the exhaust through the low speed side passage 54 and the high speed side passage 53. The engine output can be increased over the entire speed range.

  FIG. 11 is a graph showing the relationship between the engine speed and the engine torque in the engine system 100. In FIG. 11, the solid line is the engine torque torque in the engine system 100, and the broken line is the same graph as the broken line in FIG. As shown in FIG. 11, in the engine system 100, the engine output can be increased in the entire speed region.

  Here, the high speed side passage 53 and the variable flow area 58 can be omitted. That is, the exhaust gas may be allowed to pass only through the low speed side passage 54 in the entire rotational speed region. By doing so, the structure can be simplified and the cost can be kept small with the omission of the high speed side passage 53 and the flow path area variable valve 58. However, if the high-speed side passage 53 is provided as in the above-described embodiment and the exhaust gas is added to the low-speed side passage 54 and passes through the high-speed side passage 53 in the high-speed region, the exhaust resistance in the high-speed region is kept small. The engine output can be increased more reliably.

  Further, the flow path area variable valve 58 is fully opened only in a high load area where the required torque for the engine is higher than a predetermined value in the low speed area, and the flow path area variable valve 58 is opened in the other low load areas in the low speed area. May be opened at a full opening or an opening smaller than the full opening.

  Further, the position of the catalyst device 6 is not limited to the above. However, according to the engine system 100, the intake efficiency can be increased by reducing the ejector effect and the exhaust resistance, which is useful in an engine system that does not have a turbocharger. And when it does not have a turbocharger in this way, the catalyst device 6 can be directly connected to each independent exhaust passage 53 as in the above embodiment, and can be arranged at a more upstream position. The temperature of the exhaust gas flowing into the catalyst body 64 can be maintained high, and the catalyst body 64 can be activated early.

DESCRIPTION OF SYMBOLS 1 Engine main body 5 Exhaust manifold 17 Intake port 18 Exhaust port 19 Intake valve 20 Exhaust valve 30 Intake valve drive mechanism (valve drive means)
40 Exhaust valve drive mechanism (valve drive means)
52 Independent exhaust passage 53 High-speed side passage (bypass passage)
54 Low speed passage (exhaust passage)
56a Mixing tube (aggregation part)
57 High-speed side assembly (bypass passage side assembly)
58 Channel area variable valve 58b Valve actuator (Channel area variable valve driving means)
62a Upstream end of casing (final assembly)

Claims (5)

An intake / exhaust device for a multi-cylinder engine having a plurality of cylinders, each having an intake port and an exhaust port, and an intake valve capable of opening and closing the intake port and an exhaust valve capable of opening and closing the exhaust port. ,
An intake passage connected to each intake port of each cylinder;
An independent exhaust passage connected to exhaust ports of one cylinder or a plurality of cylinders whose exhaust sequences are not continuous with each other;
A collecting portion connected to the downstream end of each independent exhaust passage, where gas passing through each independent exhaust passage gathers;
Valve drive means capable of driving the intake valve and the exhaust valve of each cylinder,
The independent exhaust passages connected to the cylinders in which the exhaust order is continuous among the independent exhaust passages are connected to the collecting portion at positions adjacent to each other,
The downstream end of each independent exhaust passage is connected to another independent exhaust passage adjacent to each other by an ejector effect as exhaust is discharged from each cylinder through each exhaust port and each independent exhaust passage to the collective portion. In order to generate a negative pressure in the exhaust port connected to the exhaust passage, the flow path area is smaller on the downstream side than on the upstream side,
The valve driving means has a predetermined overlap period between the valve opening period of the intake valve and the valve opening period of the exhaust valve in each cylinder at least in a low speed region where the engine speed is lower than a preset reference speed. Between the cylinders that overlap and the exhaust sequence continues, the overlap period of one cylinder overlaps with the time when the exhaust valve of the other cylinder is open. Drive
When the engine speed is higher than the reference speed, the engine output from the multi-cylinder engine is smaller than the engine output when the flow area of each independent exhaust passage is constant. Is set near the number of revolutions,
Each of the intake passages is provided in an upstream portion thereof, and has an inversion portion that reflects a negative pressure wave of the intake pulsation generated in the vicinity of the intake valve as the intake valve is opened to become a positive pressure wave. ,
The length and cross-sectional area from the intake valve to the reversal part of the intake passage are determined when the positive pressure wave generated by the negative pressure wave reflecting once at the reversal part reaches the vicinity of the intake valve. And the intake valve closing timing are substantially the same, and the primary tuned rotational speed of the intake pulsation for obtaining the inertial supercharging effect of the intake air is higher than the reference rotational speed, and The time when the positive pressure wave generated when the negative pressure wave is inverted twice at the reversal part reaches the vicinity of the intake valve and the closing timing of the intake valve are the engine speed and the intake air The secondary tuned rotational speed of the intake pulsation that provides the inertial supercharging effect becomes lower than the reference rotational speed, and the engine rotational speed between the primary tuned rotational speed and the secondary tuned rotational speed And these synchronous rotation Among the engine speeds during the period, the negative pressure amount of the negative pressure wave of the intake pulsation that reaches the vicinity of the intake valve at the closing timing of the intake valve is the largest, and the amount of decrease in volume efficiency due to the negative pressure wave is the largest An intake / exhaust device for a multi-cylinder engine, wherein the synchronized rotational speed is set to a dimension that is lower than the reference rotational speed. The intake / exhaust device for a multi-cylinder engine according to claim 1,
The independent exhaust passage has a relationship between the diameter a of a perfect circle having the same area as the flow path area at the downstream end thereof and the diameter D of a perfect circle having the same area as the minimum flow path area of the collective part a / An intake / exhaust device for a multi-cylinder engine having a shape satisfying D ≧ 0.5. The intake / exhaust device for a multi-cylinder engine according to claim 2,
The downstream end portion of the independent exhaust passage is formed with a throttle portion with a reduced flow area,
In the independent exhaust passage, the diameter a is a diameter of a perfect circle having the same area as the flow area of the throttle portion at the downstream end, and a diameter D of a perfect circle having the same area as the minimum flow area of the gathering portion The intake / exhaust device for a multi-cylinder engine has a shape in which the relationship of a / D ≧ 0.5. The multi-cylinder engine intake / exhaust device according to any one of claims 1 to 3,
Bypass passages respectively connected to portions upstream of the downstream end of each independent exhaust passage;
Connected to the downstream end of the bypass passage, and a bypass passage-side assembly where gas passing through the bypass passages gathers;
Connected to the downstream end of the bypass passage side collecting portion and the collecting portion, and a final collecting portion in which gas passing through the bypass passage side collecting portion and the collecting portion gathers,
A variable flow area valve provided in each bypass passage and capable of changing the flow area of each bypass passage;
A channel area variable valve driving means capable of opening and closing the channel area variable valve;
The flow passage area of each bypass passage is the sum of the flow passage area of each bypass passage and the flow passage area at the downstream end of each independent exhaust passage. Is set to a dimension that is equal to or larger than the flow area of the upstream portion of the portion
The variable flow area variable valve driving means is configured to make the variable variable flow area valve smaller than fully open in at least a high load area where a required torque for the multi-cylinder engine is higher than a predetermined value in the low speed area. The multi-cylinder is characterized in that the flow rate of the exhaust gas passing through the cylinder is reduced, and in the high speed region where the engine speed is higher than the reference speed, the flow passage area variable valve is fully opened to open the bypass passages. Engine intake and exhaust system. The intake / exhaust device for a multi-cylinder engine according to claim 4,
Each bypass passage has a constant flow passage area or a shape having a flow passage area larger on the downstream side than on the upstream side, and an intake / exhaust device for a multi-cylinder engine.

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