JP5531923B2 - 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 independent 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 independent passage is disclosed. In this apparatus, the flow area of the independent exhaust passage is reduced by the valve, so that the exhaust of the cylinder in the exhaust stroke flows from the predetermined independent passage into the collecting portion at a relatively high speed. The amount of gas supplied to the turbocharger by causing the negative pressure generated in the surrounding area to act on another independent passage in the collecting portion and sucking the exhaust gas in the other independent passage downstream by the so-called ejector effect. To increase the engine output.

JP 2009-97335 A

  In an engine such as an automobile, the demand for improving the engine output is still high, and it is required to further increase the engine output with a simple configuration.

  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 the intake air amount with a simple configuration and increase the engine output.

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 exhaust passing through the independent exhaust passage, and valve driving means capable of driving an intake valve and an 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 independent exhaust passages and the collecting portions 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 collecting portions. A shape in which a negative pressure is generated in an exhaust port connected to the independent exhaust passage, and each independent exhaust passage has a diameter of a perfect circle having the same area as the flow path area of the downstream end portion thereof, and The relationship between the diameter D of a perfect circle having the same area as the minimum flow path area of the collecting portion is a / D ≧ 0.5, and at least the engine speed is set in advance in the valve driving means. In a low speed region lower than the reference rotational speed, the opening period of the intake valve and the opening period of the exhaust valve of each cylinder overlap each other with a predetermined overlap period, and one of the cylinders in which the exhaust sequence continues Cylinder The intake valve and the exhaust valve of each cylinder are driven so that the overlap period overlaps with the intake top dead center and the time when the exhaust valve of the other cylinder is open, and at least in the low speed region, the other The opening timing of the exhaust valve is determined so that the time when the negative pressure generated in the exhaust port of the one cylinder becomes maximum due to the exhaust discharged from the cylinder overlaps the overlap period of the one cylinder. Provided is an intake / exhaust device for a multi-cylinder engine, which is advanced with an increase in rotational speed.

  According to the present apparatus, the ejector effect can be used more effectively, the scavenging of the cylinder can be promoted, and the intake efficiency and thus the engine output can be increased.

Specifically, in the present apparatus, the exhaust valves of the other cylinders are opened during the overlap period of the predetermined cylinders at least in the low speed region. As the high-speed exhaust gas is ejected to the collecting portion, a negative pressure can be generated in the exhaust port of the cylinder during the overlap period by the ejector effect, and the gas in the cylinder during the overlap period can be discharged to the exhaust port side. It can be sucked out to promote scavenging. In addition, since the negative pressure is maximized during the overlap period, the negative pressure can be effectively applied to the cylinders during the intake stroke to effectively increase the intake efficiency.
In particular, the relationship between the diameter a of the perfect circle having the same area as the flow area of the downstream end portion of the independent exhaust passage and the diameter D of the perfect circle having the same area as the minimum flow area of the gathering portion is a. Since it has a shape satisfying /D≧0.5, the ejector effect can be effectively obtained.

  Here, during the valve opening period of the exhaust valve, the time when the pressure and speed of the exhaust discharged to the exhaust port become maximum is some time after the start of the opening of the exhaust valve. The maximum negative pressure is generated at the exhaust port. The delay time from the start of opening of the exhaust valve until the high-speed exhaust is discharged, and the time from the start of opening of the exhaust valve to the time when the negative pressure becomes maximum, is the crank angle and the engine speed is The higher the value, the longer. Therefore, for example, when the exhaust valve opening start timing is the same, the higher the engine speed, the slower the timing when the negative pressure becomes maximum, and the timing when the negative pressure becomes maximum may shift from the overlap period. There is.

  On the other hand, in this apparatus, the opening timing of the exhaust valve is advanced as the engine speed increases, and the time when the negative pressure becomes maximum is earlier. And the overlap period can be more reliably overlapped.

  In the above configuration, the valve driving means is configured so that at least in the low speed region, the time when the negative pressure generated in the exhaust port of the one cylinder by the exhaust exhausted from the other cylinder becomes the maximum is the engine speed. Regardless of this, it is preferable to advance the opening timing of the exhaust valve as the engine speed increases so that the crank angle timing is substantially the same (claim 2).

  In this way, regardless of the engine speed, the time when the negative pressure becomes maximum is the time when exhaust gas in the cylinder can be more effectively sucked out by this negative pressure during the overlap period, for example, near TDC. can do.

  In the above configuration, the valve driving means has a variable lift mechanism capable of changing the opening / closing timing and lift amount of the exhaust valve, and at least in the low speed region, the exhaust valve is increased as the engine speed increases by the variable lift mechanism. It is preferable to advance the valve opening start timing of the exhaust valve while increasing the lift amount.

  In this way, the opening timing of the exhaust valve can be advanced as the engine speed increases, so that the time when the negative pressure becomes maximum can be set to an appropriate time, and the engine speed can be increased. Accordingly, the scavenging performance can be ensured corresponding to the exhaust gas flow rate that increases with this.

  In the above configuration, the valve driving means increases the lift amount of the exhaust valve with the increase in engine speed by the variable lift mechanism at least in the low speed region, and starts opening the exhaust valve. It is preferable to retard the closing timing of the exhaust valve while advancing the timing.

  In this way, if the exhaust valve closing timing is retarded and thereby the end timing of the overlap period is retarded, the exhaust for overlapping the timing when the negative pressure becomes maximum and the overlap period are overlapped. Since the advance amount of the valve opening start timing can be kept small, the reduction of expansion work due to the large advance amount of the exhaust valve opening start timing can be suppressed, and the engine output is more reliably ensured. Can be increased.

In the present invention, at least the downstream portion of each independent exhaust passage has a shape in which the flow passage area is smaller on the downstream side than on the upstream side. The length and the cross-sectional area of the portion until the pressure wave of the intake pulsation generated by opening the valve is reversed are the time when the primary peak of the intake pulsation occurs in the vicinity of the intake valve and the closing of the intake valve. The primary tuning rotational speed at which the valve timing is substantially the same and the inertial supercharging effect of the intake is obtained is higher than a preset tuning reference rotational speed, and the primary valley of the intake pulsation is in the vicinity of the intake valve Is set to a dimension in which the non-synchronized rotation speed at which the intake valve closing timing and the intake valve closing timing are substantially the same is lower than the tuning reference rotation speed. Engine output The preferably set in the vicinity of the engine rotational speed is smaller than the engine output in case of a constant flow area of the independent exhaust passages (claim 5).

  In this way, the engine output can be increased in a wider engine speed range by effectively utilizing the ejector effect and the inertial supercharging effect of the intake air.

  Specifically, since the flow area of each independent exhaust passage is narrowed on the downstream side, exhaust can be ejected at a high speed from the independent exhaust passage to the collecting portion, and a high ejector effect can be obtained. Here, exhaust resistance increases because the flow area of the independent exhaust passage is reduced. On the other hand, in this configuration, the primary synchronized rotational speed at which the inertial supercharging effect of the intake air is obtained is a speed at which the engine output becomes smaller than the engine output when the flow passage area of each independent exhaust passage is constant. Because it is set in the high speed region where the engine speed is higher than the tuning reference speed set near the number, the effect of the increase in exhaust resistance is better than the scavenging performance improvement effect due to the ejector effect due to the large exhaust flow rate. Even in an increasing high speed region, the scavenging performance in the cylinder can be enhanced by the inertia supercharging effect of the intake air to suppress an increase in pump loss, and the engine output can be increased in a wider engine speed region. In particular, the non-synchronized rotational speed at which the primary valley of the intake pulsation in which the inertia supercharging effect cannot be obtained and the closing timing of the intake valve are substantially the same is lower than the tuning reference rotational speed. Therefore, the scavenging performance can be more reliably improved over the entire high speed region where the engine speed is higher than the tuning reference speed.

  As described above, according to the present invention, the ejector effect can be used more effectively, the scavenging of the cylinder can be promoted, and the intake efficiency and thus the engine output can be increased.

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 elements on larger scale of FIG. It is a schematic longitudinal cross-sectional view of the engine system corresponding to FIG. It is the IV-IV sectional view taken on the line of FIG. It is a figure for demonstrating the valve timing of an intake valve and an exhaust valve. It is the figure which showed the pressure change in an exhaust port together with the valve lift curve. It is the figure which showed the pressure change in an exhaust port in each engine speed together with the valve lift curve. It is the figure which showed the pressure change in an exhaust port in each engine speed together with the valve lift curve. It is a figure for demonstrating the valve-opening timing and valve-closing timing of the intake valve and exhaust valve which concern on embodiment of this invention. It is the figure which showed typically the mode of the intake air pulsation near an intake valve. It is a graph for demonstrating the setting procedure of a tuning reference | standard rotation speed. It is a figure for demonstrating the valve opening time and valve closing time of the exhaust valve 20 which concerns on other embodiment of this invention.

  An embodiment of an 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 constituting an intake passage, and these intake pipes 3. A surge tank 4, an exhaust manifold 5 connected to the engine body 1, and a catalyst device 6 connected to the exhaust manifold 5.

  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. 5, 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 an exhaust camshaft 41 and an exhaust CVVL 42 connected to the exhaust valve 20. The exhaust camshaft 41 is connected to the crankshaft similarly to the intake camshaft 31 and rotates with the rotation of the crankshaft to drive the exhaust valve 20 to open and close.

  The exhaust CVVL 42 is for changing the valve timing of the exhaust valve 20. The exhaust CVVL 42 is a so-called continuously variable valve lift mechanism capable of continuously (steplessly) changing the lift amount of the exhaust valve 20.

  In the present embodiment, the exhaust CVVL 42 changes the lift amount of the exhaust valve 20 while variably changing the lift amount of the exhaust valve 20 while fixing the valve closing timing of the exhaust valve 20 at a predetermined crank angle position. change.

  A CVVL having such a configuration is already known, and a specific example thereof is disclosed in, for example, Japanese Patent Application Laid-Open No. 2007-85241 (referred to as VVE in the same document).

  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. In this way, pressure pulsation (intake pulsation) as shown in FIG. 10 occurs in the intake pipe 3. FIG. 10 is a diagram schematically showing a pressure change due to the intake pulsation in the vicinity of the intake valve 19 with the horizontal axis as time. As shown in FIG. 10, in the vicinity of the intake valve 19, after the start of opening of the intake valve 19 (time t0), peaks (high pressure state) and valleys (low pressure state) appear alternately while being attenuated. .

  Inertial supercharging effect of intake air means that immediately before the intake valve 19 is closed, the vicinity of the intake valve 19 is in a mountain state and the pressure on the intake port 17 side is increased, and intake to the cylinder 12 is promoted. That's it. In particular, 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 a large pressure, and a high inertial supercharging effect can be obtained.

  In the present embodiment, the length L_in of the intake pipe 3 (see FIG. 5) is obtained so that a high inertial supercharging effect due to this primary peak is obtained in a high speed region where the engine speed is higher than the tuning reference speed N1 described later. 1) and a cross-sectional area, that is, an area in a direction orthogonal to the flow direction of the exhaust gas or a flow path area 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 L_in and the cross-sectional area of the tube 3 are set.

  That is, the primary tuning rotational speed Nin_1a at which the time t1a (see FIG. 10) 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 tuning reference. The time t1b (see FIG. 10) at which the primary trough of the intake pulsation occurs in the vicinity of the intake valve 19 and the closing timing of the intake valve 19 so as to be higher than the rotational speed N1 and the closing timing of the intake valve 19 are substantially the same. The length L_in and the cross-sectional area of the intake pipe 3 are set so that the non-tuned rotation speed Nin_1b is lower than the tuning reference rotation speed N1.

  Here, in order to obtain a high inertial supercharging effect due to the primary mountain in a higher speed region, the length L_in of the intake pipe 3 is made shorter or the cross-sectional area of the intake pipe 3 is made larger. However, since a change in the cross-sectional area of the intake pipe 3 is greatly limited in layout, in this embodiment, by shortening the length L_in of the intake pipe 3, a high inertial supercharging effect can be obtained in a high speed region. I try to get it.

  The exhaust manifold 5 includes three independent exhaust passages 52, a mixing pipe (aggregation part) 56a, a straight pipe 56b, and a diffuser 56c.

  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.

  The independent exhaust passage 52 corresponding to the exhaust port 18 of the second cylinder 12b and the third cylinder 12c extends linearly facing the central portion of the cylinders 12b and 12c, that is, the substantially central portion of the engine body 1, The independent exhaust passages 52 corresponding to the exhaust ports 18 of the other cylinders 12a and 12d are directed from the positions facing the corresponding exhaust ports 18 to the independent exhaust passages 52 corresponding to the second cylinder 12b and the third cylinder 12c. Curved and extended.

  The mixing pipe 56a is connected to the downstream side of each independent exhaust passage 52, and the exhaust gas that has passed through each independent exhaust passage 52 gathers in the mixing pipe 56a. In the mixing pipe 56a, the three independent exhaust passages 52 are connected to the mixing pipe 56a at positions where their downstream ends are adjacent to each other.

  Each of the independent exhaust passages 52 and the mixing pipes 56a is negatively generated around the high-speed exhaust gas as the exhaust gas is ejected from the independent exhaust passages 52 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 independent exhaust passage 52 and the gas in the exhaust port 18 communicating with the independent exhaust passage 52 are sucked downstream.

  Specifically, each of the independent exhaust passages 52 has a shape in which the flow passage area decreases toward the downstream, and the exhaust is ejected from each independent exhaust passage 52 to the downstream side at a high speed. ing. More specifically, as shown in FIGS. 2 and 4, each of the independent exhaust passages 52 is reduced in cross-sectional area from the upstream portion having a substantially elliptical cross section toward the downstream, and at the downstream end, It has a sector shape that is approximately 1/3 of the elliptical cross-sectional area of the side portion. These independent exhaust passages 52 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 has a cross-sectional area at the upstream end larger than the total area of the downstream ends of the independent exhaust passages 52, the diameter of the mixing pipe 56a decreases toward the downstream, and the flow path area decreases. The diameter of a perfect circle having the same area as the flow path area of D is D (see FIG. 3), and the diameter of a perfect circle having the same area as the cross-sectional area of the downstream end of the independent exhaust passage 52 is a (see FIG. 3). In this case, 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 flow rate of the exhaust gas into the mixing pipe 56a, when a portion having a reduced flow area, that is, a throttle portion is provided at the downstream end of the independent exhaust passage 52, the throttle portion The diameter of the channel area is preferably a, and the mixing tube 56a is preferably shaped so that a / D ≧ 0.5.

  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.

  A casing 62 (to be described later) of the catalyst device 6 is connected to the downstream side of the diffuser 56c, and the exhaust gas that has passed through the diffuser 56c flows into the casing 62.

  The catalyst device 6 is a device for purifying the exhaust discharged from the engine body 1. The catalyst device 6 includes a catalyst main body 64 such as a three-way catalyst and a casing 62 that accommodates the catalyst main body 64. The casing 62 has a substantially cylindrical shape extending in parallel with the exhaust flow direction. The catalyst body 64 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 end of the diffuser 56c is connected to the upstream end 62a of the casing 62, and the exhaust discharged from the diffuser 56c flows into the upstream end 62a of the casing 62 and then proceeds to the catalyst body 64 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, performs various calculations based on the signals, and transmits drive signals to various actuators based on the calculation results. The intake VVT 32 and the exhaust CVVL 42 receive the signal from the ECU 2 and drive the intake valve 19 and the exhaust valve 20.

  The intake VVT 32 and the exhaust CVVL 42 receive a drive signal from the ECU 2, and the intake valve 19 and the exhaust valve 20 are set to have an opening period of the exhaust valve 20 and an opening period of the intake valve 19 in the entire operation region. The engine is driven so that it overlaps across the intake top dead center (TDC) and the exhaust valve 20 starts to open during the overlap period T_O / L of the other cylinders 12. Specifically, as shown in FIG. 5, the exhaust valve 20 of the third cylinder 12c opens 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. Thus, the valves 19 and 20 are driven.

  The exhaust CVVL 32 advances the valve opening start timing of the exhaust valve 20 while increasing the lift amount of the exhaust valve 20 as the engine speed increases. More specifically, the exhaust CVVL 32 is configured so that the time when the pressure in the exhaust port 18 becomes the maximum negative pressure generated along with the discharge of blowdown gas described later is substantially constant regardless of the engine speed. As the engine speed increases, the valve opening start timing of the exhaust valve 20 is advanced. In the present embodiment, the angle is advanced so that the time when the maximum negative pressure is reached substantially coincides with TDC.

  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.

  In this apparatus, the independent exhaust passage 52 and the mixing pipe 56a are negatively pressured in the other independent exhaust passage 52 as the exhaust gas is ejected from the predetermined independent exhaust passage 52 to the mixing pipe 56a at a high speed due to the ejector effect. And the gas in the exhaust port 18 communicating with the other independent exhaust passage 52 and the other independent exhaust passage 52 is sucked out to the downstream side. Then, during the overlap period of a predetermined cylinder 12 (hereinafter referred to as intake stroke cylinder 12 as appropriate), another cylinder 12 (hereinafter referred to as an exhaust stroke as appropriate) whose exhaust sequence is set to be one after this intake stroke cylinder 12. The exhaust valve 20 of the cylinder 12) is set to open.

  Accordingly, as the exhaust valve 20 of the exhaust stroke cylinder 12 is opened and the exhaust gas in the exhaust stroke cylinder 12 is ejected through the independent exhaust passage 52 to the mixing pipe 56a at a high speed, it is overlapped by the ejector effect. A negative pressure is generated in the exhaust port 18 of the intake stroke cylinder 12 during the period, thereby promoting scavenging in the intake stroke cylinder 12 during the overlap period. In particular, the downstream end of each independent exhaust passage 52 is disposed adjacent to the mixing pipe 56a. Therefore, the suction force by the independent exhaust passage 52 connected to the exhaust stroke cylinder 12 effectively acts on the independent exhaust passage 52 connected to the intake stroke cylinder 12.

  FIG. 6 shows the result of examining the effect of the negative pressure in the exhaust port 18 due to the ejector effect. FIG. 6 shows the result of examining the pressure change in the exhaust port 18 using the engine system 100 together with the lift curves of the exhaust valve 20 and the intake valve 19, and in FIG. Two pressure graphs P1 and P2 indicate pressure changes in the exhaust ports 18 of two cylinders in which the exhaust order is continuous. In FIG. 6, the pressure in each exhaust port 18 becomes the maximum pressure (Pmax) immediately after the opening of the exhaust valve 20 (when the crank angle = CA_1 at the pressure P1), and the exhaust order is one before. A short time after the pressure of the cylinder 12 reaches the maximum pressure (a time when the crank angle = CA_2 at the pressure P2), the maximum negative pressure (Pmin) is reached. The maximum negative pressure generated by the ejector effect is the exhaust valve 20. The high-pressure, high-speed gas (so-called blowdown gas) discharged from the cylinder 12 immediately after the start of the valve opening is generated as it is ejected through the independent exhaust passage 52 to the mixing pipe 56a at a high speed. I understand.

  Here, after the opening of the exhaust valve 20 is started, the crank angle period until the blowdown gas is discharged from the exhaust stroke cylinder 12, and thus from the start of the opening of the exhaust valve 20 to the intake stroke cylinder 12 by the blowdown gas. Until the high negative pressure is generated, the crank angle period becomes longer as the engine speed increases, as shown in FIG. FIG. 7 shows the opening start timing and the closing timing of the exhaust valve 20 at a plurality of different engine speeds using an apparatus having an engine body 1 and an exhaust manifold having the same structure as the engine system 100. It is the figure which showed the result of having investigated the pressure change in the exhaust port 18 as fixed irrespective of the rotation speed. In FIG. 7, lines L1, L2, L3, L4, L5, and L6 are the results when the engine speed is 1000 rpm, 2000 rpm, 3000 rpm, 4000 rpm, 5000 rpm, and 6000 rpm, respectively. As shown in FIG. 7, when the valve opening start timing and the valve closing timing of the exhaust valve 20 are the same, the blowdown gas is increased after the exhaust valve 20 starts to open as the engine speed increases. The crank angle period until the is discharged is long. Along with this, the time when the negative pressure in the exhaust port 18 becomes maximum is delayed by the crank angle. This is presumably because the absolute time from when the exhaust valve 20 starts to open until the blowdown gas is discharged into the exhaust port 18 is constant regardless of the engine speed. Therefore, for example, if the valve opening start timing and the valve closing timing of the exhaust valve 20 are constant regardless of the engine speed, the negative pressure in the exhaust port 18 cannot be maximized during the overlap period. The scavenging promotion effect due to negative pressure cannot be exhibited effectively.

  In contrast, in the engine system 100, as described above, the valve opening start timing of the exhaust valve 20 is advanced as the engine speed increases. Therefore, the negative pressure in the exhaust port 18 becomes maximum in the vicinity of TDC during the overlap period regardless of the engine speed, so that the scavenging promotion effect can be effectively obtained. In particular, since the negative pressure in the exhaust port 18 is maximized in the vicinity of the TDC, the residual gas in the cylinder 12 can be sucked out more effectively, and high intake efficiency can be obtained.

  FIG. 8 shows a change in pressure in the exhaust port 18 corresponding to FIG. 7 in the engine system 100 in which control is performed to advance the valve opening start timing of the exhaust valve 20 as the engine speed increases. As shown in FIG. 8, in the engine system 100, the pressure in the exhaust port 18 is a high negative pressure in the vicinity of a high TDC regardless of the engine speed.

  Further, in the engine system 100, the lift amount of the exhaust valve 20 is increased as the engine speed increases, and the exhaust gas whose flow rate increases as the engine speed increases can be efficiently discharged. Therefore, it is possible to increase the intake efficiency by reducing the exhaust resistance while ensuring the effect of improving the intake efficiency by the ejector effect.

  Further, since the flow area of the independent exhaust passage 52 is reduced on the downstream side so that the ejector effect can be effectively obtained, if the engine speed increases and the exhaust flow rate increases, the exhaust resistance increases. As a result of an increase in pump loss and an increase in pump loss, there arises a problem that the engine output deteriorates. However, the present engine system 100 is configured so that the inertial supercharging effect of intake air can be effectively obtained in a high speed region. It is possible to maintain a high earner output even in a high speed region.

  Specifically, in the engine system 100, the tuning reference rotational speed N1 is set such that the engine output is reduced by reducing the flow area of the exhaust passage by reducing the flow area of the independent exhaust passage 52 as described above. It is set near the rotational speed at which the engine output is lower than when it is not. That is, the tuning reference rotational speed N1 is determined when the engine output of the engine system 100 uses an exhaust passage having a constant flow area instead of the independent exhaust passage 52 in the engine system 100 (hereinafter referred to as flow rate). (It may be called a constant road area specification). In the engine system 100, as described above, by adjusting the length L_in and the cross-sectional area of each intake pipe 3, the primary tuning rotational speed Nin_1a is higher than the tuning reference rotational speed N1, and the primary non-rotating speed. The tuning rotation speed Nin_1b is set lower than the tuning reference rotation speed 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 tuning reference speed N1 in which the engine output decreases as the independent exhaust passage 52 is throttled. Thus, it is possible to suppress an increase in pump loss and to suppress a decrease in engine output.

  FIG. 11 shows an example (solid line) of the engine output (engine torque) of the engine system 100 and an example (broken line) of the engine output (engine torque) with the constant flow path area specification. As shown in FIG. 11, in the region where the engine speed is lower than the torque reduction speed N0, the engine system 100 has higher engine torque than the constant flow path area specification. On the other hand, in the region where the engine speed is higher than the torque-decreasing speed N0, the engine system 100 has a lower engine torque than 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 tuning reference rotational speed N1 is set to 4500 rpm. The intake pipe 3 is set to have a diameter of 45 mm and a length L_in of 400 mm so that the primary synchronized rotation speed Nin_1a5000 rpm is obtained. Note that the displacement of the engine body 1 having four cylinders according to the present embodiment is 2 liters. In this case, the primary non-synchronization rotational speed Nin_1b is 3500 rpm. FIG. 11 schematically shows how the pressure changes due to the intake pulsation in the vicinity of the intake valve 19 at the closing timing of the intake valve 19 in accordance with the engine torque.

  The engine torque comparison result shown in FIG. 11 is a comparison in the case where the intake pulsation effect that can be obtained with the same length and cross-sectional area of the intake pipe 3 is the same. Accordingly, an engine system configured to use an exhaust passage having a constant flow area instead of the independent exhaust passage 52 and obtain an inertial supercharging effect of intake air in a region where the engine speed is relatively low as in the prior art. As shown by the chain line in FIG. 11, the engine output of the engine is lower than that of the engine system 100 in the high speed region. In the engine system 100, the engine is higher than the conventional engine system due to the inertial supercharging effect of the intake air. Output can be obtained.

  As described above, according to the present engine system 100, the ejector effect can be used more effectively, the scavenging of the cylinder 12 can be promoted, the intake efficiency can be increased, and the intake inertia effect can be effectively achieved in the high speed region. Therefore, the pump loss can be suppressed to a small value, and the engine output can be increased in the entire speed range.

  Here, in the above-described embodiment, the case where the closing timing of the exhaust valve is made constant is shown. However, if the opening timing of the exhaust valve 20 is excessively advanced, the expansion work in the engine body 1 may be reduced. There is. Therefore, in such a case, as shown in FIG. 12, the valve opening start timing of the exhaust valve 20 is advanced as the engine speed increases, and the exhaust valve 20 is closed as the engine speed increases. It is preferable to retard the timing. In this case, the delay amount of the closing timing of the exhaust valve 20 with respect to the increase amount of the engine speed is set smaller than the advance amount of the opening start timing of the exhaust valve 20 with respect to the increase amount of the engine speed. It is preferable to do this.

  In the above embodiment, the intake valve 19 and the exhaust valve 20 are overlapped and the opening of the exhaust valves 20 of the other cylinders 12 is started so that a high scavenging acceleration effect due to the ejector effect can be obtained in the entire operation region. The case where the control for overlapping the time period and the overlap period and the control for overlapping the time period when the maximum negative pressure is accompanied with the increase in the engine speed and the overlap period are described. You may perform only in the driving | running | working area | region lower than a predetermined reference | standard rotation speed. That is, in the operating region where the engine speed is high, the exhaust gas flow rate increases, so that the scavenging promotion effect obtained by reducing the pump loss may be higher than the scavenging promotion effect obtained by the ejector effect. Therefore, in such a case, it is preferable to control the intake valve 19 and the exhaust valve 20 so that the scavenging promotion effect can be further enhanced.

  Further, as described above, in order to reduce the pump loss in the operation region where the engine speed is high, the region from the region where the flow area becomes small in each of the independent exhaust passages 52 to the downstream portion of the diffuser 56c is bypassed. A passage is provided, the passage area is made constant and the exhaust resistance is not increased, and a valve for opening and closing the passage is attached to the passage, and the valve is operated in the operation region where the engine speed is low. It may be configured such that the exhaust passes only through the independent exhaust passage 52 and that the exhaust also passes through the bypass passage by opening the valve in an operation region where the engine speed is high.

  Further, the position of the catalyst device 6 is not limited to the above. However, according to the present engine system 100, the intake efficiency can be increased by reducing the ejector effect and the back pressure, which is useful in an engine system that does not have a turbocharger. When the turbocharger is not provided as described above, the catalyst device 6 can be directly connected to each independent exhaust passage 52 as in the above-described embodiment, and can be disposed at a position on the upstream side. 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 56a Mixing pipe (collecting part)

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 the exhaust 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,
Each of the independent exhaust passages and the collective portion includes another independent exhaust passage and the independent exhaust adjacent thereto due to an ejector effect as exhaust gas is discharged from each cylinder through the exhaust port and the independent exhaust passage to the collective portion. A shape in which negative pressure is generated in the exhaust port connected to the passage;
Each of the independent exhaust passages has a relationship between a diameter a of a perfect circle having the same area as the flow path area of the downstream end portion thereof and a diameter D of a perfect circle having the same area as the minimum flow path area of the collecting portion. /D≧0.5 has a shape,
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. Intake of each cylinder so that the overlap period of one cylinder overlaps between the intake top dead center and the time when the exhaust valve of the other cylinder is opened between the cylinders that overlap and the exhaust sequence continues. The valve and the exhaust valve are driven, and at least in the low-speed region, the time when the negative pressure generated in the exhaust port of the one cylinder by the exhaust exhausted from the other cylinder becomes the maximum is the excess of the one cylinder The multi-cylinder engine is characterized in that the opening timing of the exhaust valve is advanced as the engine speed increases so as to overlap with a lap period. Jin intake and exhaust device. The intake / exhaust device for a multi-cylinder engine according to claim 1,
In the valve driving means, at least in the low speed region, the time when the negative pressure generated in the exhaust port of the one cylinder is maximized by the exhaust discharged from the other cylinder is substantially the same regardless of the engine speed. An intake / exhaust device for a multi-cylinder engine, wherein the valve opening start timing of the exhaust valve is advanced as the engine speed increases so that the crank angle timing is reached. An intake / exhaust device for a multi-cylinder engine according to claim 1 or 2,
The valve drive means has a variable lift mechanism capable of changing the opening / closing timing and lift amount of the exhaust valve, and at least in the low speed region, the lift variable amount of the exhaust valve is increased by the variable lift mechanism as the engine speed increases. An intake / exhaust device for a multi-cylinder engine, wherein the valve opening start timing of the exhaust valve is advanced while increasing. The intake / exhaust device for a multi-cylinder engine according to claim 3,
The valve driving means increases the lift amount of the exhaust valve as the engine speed increases by the variable lift mechanism at least in the low speed region, and advances the valve opening start timing of the exhaust valve. An intake / exhaust device for a multi-cylinder engine, wherein the closing timing of the exhaust valve is retarded. The intake / exhaust device for a multi-cylinder engine according to any one of claims 1 to 4,
  At least the downstream portion of each independent exhaust passage has a shape in which the flow area is smaller on the downstream side than on the upstream side,
  The length and cross-sectional area of the portion of each intake passage from the intake valve until the pressure wave of the intake pulsation generated by opening the intake valve is reversed is the first peak of the intake pulsation. The timing that occurs in the vicinity of the intake valve and the valve closing timing of the intake valve are substantially the same, and the primary tuning rotational speed at which the inertial supercharging effect of the intake is obtained is higher than a preset tuning reference rotational speed, and The non-tuned rotational speed at which the primary valley of the intake pulsation occurs in the vicinity of the intake valve and the closed timing of the intake valve is set to a dimension that is lower than the tuning reference rotational speed.
  The tuning reference rotational speed is set in the vicinity of an engine rotational speed at which 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. Multi-cylinder engine intake and exhaust system.

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