JP2013036353A - Exhaust device of multicylinder engine - Google Patents

Abstract

PROBLEM TO BE SOLVED: To combine using an ejector effect in a low-speed area and reducing exhaust resistance in a high-speed area.SOLUTION: The exhaust device includes: independent exhaust passages 16A-16D connected to cylinders 2A-2D; a collection part 17 where downstream ends of the respective independent exhaust passages 16A-16D are collected; a merging part 18 connected to the downstream side of the collection part 17; communication passages 25A, 25B communicating the respective independent exhaust passages of a plurality of cylinders, exhaust order of which is not in series, (between the exhaust passages 16A and 16D, and between the exhaust passages 16B and 16C); and variable valves 26 respectively provided for the communication passages 25A, 25B. Valve timing of each cylinder is set so that an overlapping period (O/L) of one cylinder in cylinders exhaust order of which is in series overlaps a valve-opening period of an exhaust valve 6 of the other cylinder. Opening of the variable valve 26 is kept at 100% (full throttle) when engine rotation speed is equal to or higher than a predetermined value Nex, and is reduced to lower than 100% when the engine rotation speed is lower than the predetermined value Nex.

Description

  The present invention relates to an exhaust device provided in a multi-cylinder engine having four or more cylinders and having an intake port, an exhaust port, an intake valve, and an exhaust valve provided in each cylinder.

  Conventionally, in an exhaust system of a multi-cylinder engine, attempts have been made to increase exhaust efficiency by utilizing a so-called ejector effect. The ejector effect is an action of sucking out another fluid by a negative pressure generated around the fluid ejected at a high speed.

  For example, in Patent Document 1 below, as an exhaust device for an in-line four-cylinder engine, four independent exhaust passages connected to the exhaust ports of each cylinder, a collective portion in which each independent exhaust passage gathers in one place, There is disclosed a variable exhaust valve provided on the upstream side of the collecting portion and having a variable exhaust valve capable of changing the flow area of each independent exhaust passage.

  In the exhaust device described in Patent Document 1, the variable exhaust valve is operated to reduce the flow area of each independent exhaust passage, so that exhaust gas discharged from a certain cylinder is transferred from the independent exhaust passage to the collecting portion. It flows in at high speed. Then, a negative pressure is generated around the exhaust gas flowing in at a high speed, and this negative pressure acts on another independent exhaust passage (ejector effect), so that the exhaust gas in the other independent exhaust passage flows downstream. The air is sucked out and the scavenging is promoted.

JP 2009-97335 A

  By the way, even when a plurality of independent exhaust passages are gathered in the collecting portion as described above to promote scavenging by the ejector effect, the exhaust gas flow resistance increases at a high speed region of the engine where the exhaust gas flow rate is high. As a result, the pump loss increases, and the advantages of the ejector effect are diminished, which may cause a decrease in engine output.

  The present invention has been made in view of the circumstances as described above, and is an exhaust system for a multi-cylinder engine capable of achieving both utilization of the ejector effect in the low speed range and reduction of exhaust resistance in the high speed range. The purpose is to provide.

  In order to solve the above-described problems, the present invention is an exhaust system provided in a multi-cylinder engine having four or more cylinders, and an intake port, an exhaust port, an intake valve, and an exhaust valve provided in each cylinder. And four or more independent exhaust passages individually connected to the exhaust ports of the cylinders, an aggregating portion aggregated so that downstream end portions of the independent exhaust passages are close to each other, Two or more communication passages that are connected to the downstream side and that have a common space that communicates with all of the independent exhaust passages formed therein and that communicate with the independent exhaust passages of a plurality of cylinders that are not in the exhaust order. And a variable valve that can be opened and closed respectively provided in each of the communication passages, and a control means that controls each of the variable valves. The cross-sectional area of the passage It is formed so as to be small, and at least in the high load region of the engine, the opening period of the intake valve and the exhaust valve of each cylinder overlaps with a predetermined overlap period, and one of the cylinders in which the exhaust sequence continues is continued. The operation timing of the intake valve and the exhaust valve of each cylinder is set so that the overlap period of the cylinder overlaps with the opening period of the exhaust valve of the other cylinder. The opening of each variable valve is maintained at a predetermined high opening when the engine speed is equal to or higher than a predetermined value, and the opening of each variable valve is set to the above when the engine speed is lower than the predetermined value. A value smaller than a predetermined high opening is set (claim 1).

  In the present invention, the opening degree of each variable valve provided in the plurality of communication passages is set to a predetermined high opening degree in a high speed range where the engine rotation speed is equal to or higher than a predetermined value, and a plurality of cylinders whose exhaust order is not continuous are independent. By making the exhaust passages communicate with each other, the flow of exhaust gas discharged from a certain cylinder can be branched through the communication passages, and the exhaust gas can be passed through a plurality of independent exhaust passages. Thus, the exhaust resistance can be effectively suppressed and the pump loss can be reduced in the high speed region of the engine where the exhaust gas flow rate is particularly fast and the exhaust resistance tends to increase.

  On the other hand, when the engine speed is lower than the predetermined value, the opening degree of the variable valve is reduced, so that the branch flow of the exhaust gas is suppressed or stopped. Then, when an exhaust valve starts to open in a certain cylinder, exhaust gas (blowdown gas) exhausted vigorously from the cylinder is jetted to a joining portion mainly through an independent exhaust passage connected to the cylinder. As a result, a large negative pressure is generated around the jet flow. At this time, in the cylinder in which the exhaust order is one before the cylinder in which the exhaust valve has started to open, there is an overlap period in which both the intake valve and the exhaust valve are open. The exhaust gas is sucked out from the air and the inflow of air from the intake port is performed at the same time, thereby promoting scavenging and increasing the intake amount.

  Thus, according to the present invention, by utilizing the ejector effect in the low speed region of the engine to promote scavenging and the like, the engine output can be effectively improved, and the exhaust gas flow rate is particularly high. In the engine high speed range, the pump loss can be effectively reduced by branching the flow of the exhaust gas and suppressing the exhaust resistance.

  In the present invention, preferably, the control means reduces the opening of each of the variable valves as the rotation speed is lower, at least when the engine rotation speed is lower than the predetermined value, at least in a high engine load range. Is set to 0% (claim 2).

  According to this configuration, when the engine speed is the lowest, the opening of the variable valve is set to 0% and the communication path is completely shut off, thereby maximizing the ejector effect and effective scavenging. Can be promoted. On the other hand, in an intermediate speed range between the speed range in which the opening degree of the variable valve is set to 0% and the speed range in which the opening degree of the variable valve is set to the predetermined high opening degree, the engine speed is reduced. Accordingly, by linearly controlling the opening of the variable valve, it is possible to effectively prevent a sudden change in the engine output accompanying a change in the opening of the variable valve, unlike when the variable valve is suddenly opened and closed.

  In the present invention, preferably, each of the variable valves comprises a butterfly valve having a common rotating shaft, and the control means drives the opening and closing of the variable valves by rotating the common rotating shaft. 3).

  According to this configuration, since a plurality of variable valves can be driven simultaneously by rotating only one rotating shaft, each variable valve can be driven to open and close easily and reliably while matching the opening degrees of the variable valves. .

  In the present invention, preferably, the connection angle between each of the independent exhaust passages and the communication passage connected thereto is set to be the same for all the independent exhaust passages (claim 4).

  According to this configuration, when the variable valve is opened, the ratio of the exhaust gas that branches from the independent exhaust passage of one cylinder to the other independent exhaust passage through the communication passage is exhausted from any cylinder. Therefore, the exhaust resistance reduction effect can be effectively prevented from varying from cylinder to cylinder.

  In the present invention, preferably, the angle formed by the axis of the downstream end portion of each independent exhaust passage and the axis of the merging portion is set to be the same for all the independent exhaust passages (Claim 5).

  According to this configuration, the inflow angle of the exhaust gas flowing from each cylinder through the independent exhaust passage into the joining portion becomes the same regardless of whether the exhaust gas is discharged from any cylinder. It is possible to effectively prevent the effect of the above from varying from cylinder to cylinder.

  As described above, according to the exhaust system for a multi-cylinder engine of the present invention, it is possible to achieve both the utilization of the ejector effect in the low speed range and the reduction of the exhaust resistance in the high speed range.

It is a figure showing the whole engine composition to which the exhaust system of the present invention is applied. It is the front view illustrated centering on the some independent exhaust passage with which the said exhaust apparatus is equipped. It is the perspective view illustrated centering on the said independent exhaust passage. FIG. 4 is a partially cutaway perspective view of FIG. 3. (A) (b) is explanatory drawing regarding the angle which the axis line of the downstream end part of each said independent exhaust passage and the axis line of a confluence | merging part make, and the connection angle of each said independent exhaust passage and a communicating path. It is a figure for demonstrating the valve timing in each cylinder of the said engine. It is a figure for demonstrating how the opening degree of the variable valve which opens and closes the said communicating path is controlled according to driving | running conditions. It is a schematic diagram for demonstrating the flow of exhaust gas when the said variable valve is fully closed. It is a schematic diagram for demonstrating the flow of exhaust gas when the said variable valve is opened.

(1) Overall Configuration of Engine FIG. 1 is a diagram showing an overall configuration of an engine to which an exhaust device of the present invention is applied. The engine shown in the figure is a four-cycle spark ignition type multi-cylinder engine mounted on a vehicle as a power source for driving driving. Specifically, the engine of the present embodiment includes an in-line four-cylinder engine body 1 having four cylinders 2A to 2D arranged in a row, and an intake device 9 for introducing combustion air into the engine body 1. An exhaust device 15 for exhausting exhaust gas generated by combustion of the air-fuel mixture performed in the engine body 1 and an ECU 30 (control means) for controlling each part of the engine are provided.

  A piston (not shown) is inserted into each of the cylinders 2A to 2D of the engine body 1 so as to be slidable back and forth, and a combustion chamber is defined above each piston. This combustion chamber is a space for burning an air-fuel mixture of fuel and air injected from an injector (fuel injection valve) (not shown). That is, in the combustion chambers of the cylinders 2A to 2D, the air-fuel mixture is burned by spark ignition (ignition by a spark plug 7 described later), and high-temperature exhaust gas is generated. The generated exhaust gas is discharged from the combustion chamber to the exhaust device 15 in the exhaust stroke of each of the cylinders 2A to 2D.

  In the upper part (cylinder head) of the engine body 1, an intake port 3 for introducing air supplied from the intake device 9 into the combustion chamber of each cylinder 2A to 2D, and an intake valve 5 for opening and closing the intake port 3 are provided. And an exhaust port 4 for leading the exhaust gas generated in the combustion chambers of the cylinders 2A to 2D to the exhaust device 15 and an exhaust valve 6 for opening and closing the exhaust port 4. The illustrated engine is a so-called double overhead camshaft (DOHC) engine, and two intake valves 5 and two exhaust valves 6 are provided for each cylinder.

  Four spark plugs 7 are provided on the upper portion (cylinder head) of the engine body 1 so as to face the combustion chambers of the cylinders 2A to 2D from above. Each spark plug 7 supplies ignition energy to the air-fuel mixture of each of the cylinders 2A to 2D in response to power supply from an ignition circuit (not shown). In the in-line four-cylinder engine as in the present embodiment, ignition is performed at a timing shifted by 180 ° CA in the order of the first cylinder 2A → the third cylinder 2C → the fourth cylinder 2D → the second cylinder 2B. An exhaust stroke or the like is performed. Note that “° CA” represents a rotation angle (crank angle) of a crank shaft that is an output shaft of the engine.

  The intake device 9 includes a single intake pipe 10 equipped with components (not shown) such as a throttle valve, and a surge tank 11 connected to a downstream end (end on the engine body 1 side) of the intake pipe 10. And four independent intake passages 12 communicating the surge tank 11 and the intake ports 3 of the cylinders 2A to 2D.

  The exhaust device 15 is a so-called 4-1 type, and includes four independent exhaust passages 16A to 16D individually connected to the exhaust ports 4 of the respective cylinders 2A to 2D, and downstream of the independent exhaust passages 16A to 16D. An end portion (an end portion in a direction away from the engine main body 1) is aggregated (bundled) so as to be close to each other, and is connected to a downstream side of the aggregation portion 17, and is connected to the independent exhaust passages 16A to 16D. A common space communicating with all of them has a merging portion 18 formed therein, and one exhaust pipe 19 connected to the downstream side of the merging portion 18. Although not shown in FIG. 1, the exhaust pipe 19 on the downstream side is equipped with a catalyst for purifying exhaust gas, a silencer, and the like.

(2) Specific Configuration of Exhaust Device FIGS. 2 and 3 are a front view and a perspective view illustrating the independent exhaust passages 16 </ b> A to 16 </ b> D of the exhaust device 15. FIG. 4 is a partially cutaway perspective view of FIG. As shown in FIGS. 2 to 4 and FIG. 1, a flange 13 is fixed to a side surface of the engine body 1 on the exhaust side via bolts or the like. Four holes 14 corresponding to the exhaust ports 4 of 2A to 2D are provided. The upstream end portions (end portions on the engine body 1 side) of the four independent exhaust passages 16A to 16D are connected to the positions of the holes 14 of the flange 13, whereby the independent exhaust passages 16A to 16D are connected. Is attached in communication with the exhaust ports 4 of the cylinders 2A to 2D.

  The four independent exhaust passages 16 </ b> A to 16 </ b> D are curved so that the downstream side away from the engine body 1 is directed toward the inner side (the center side in the cylinder row direction of the engine body 1). It is collected in one place at a position corresponding to 2C. And the said aggregation part 17 is comprised by the downstream end part of this aggregated each independent exhaust passage 16A-16D, the holding member etc. which hold | maintain these in an aggregation state. As shown in FIG. 2, the downstream end portions of the independent exhaust passages 16 </ b> A to 16 </ b> D each have a fan-shaped cross section obtained by dividing a circle into four parts, and each downstream end portion having such a cross section has By gathering four, the substantially circular aggregation part 17 is formed as a whole.

  The downstream end portions of the independent exhaust passages 16A to 16D arranged close to each other in the aggregation portion 17 are formed in a nozzle shape whose passage cross-sectional area decreases toward the downstream side (see FIG. 5). For this reason, the exhaust gas that has passed through the downstream end of each of the independent exhaust passages 16A to 16D is accelerated there (after increasing the flow velocity) and then ejected to the junction 18.

  FIGS. 5A and 5B are diagrams for explaining at what angle the exhaust gas that has passed through the downstream end of each of the independent exhaust passages 16 </ b> A to 16 </ b> D flows into the junction 18. . As shown in this figure, the angle formed by the axis of the downstream end of the independent exhaust passage 16A connected to the first cylinder 2A and the axis of the merging portion 18 is θa, and the independent exhaust passage connected to the second cylinder 2B. The angle between the axis of the downstream end of 16B and the axis of the merging portion 18 is θb, and the angle of the axis of the downstream end of the independent exhaust passage 16C connected to the third cylinder 2C and the axis of the merging portion 18 Is θc, and θd is an angle formed by the axis of the downstream end of the independent exhaust passage 16D connected to the fourth cylinder 2D and the axis of the merging portion 18, all four angles are set to be the same. (Θa = θb = θc = θd). As a result, the inflow angles of the exhaust gas flowing from the cylinders 2A to 2D through the independent exhaust passages 16A to 16D into the merging portion 18 are equalized for all the independent exhaust passages. In addition, from the viewpoint of further enhancing the ejector effect described later, it is desirable that the angles θa to θd be as small as possible.

  As shown in FIG. 1, the nozzle portion 20 formed so that the passage cross-sectional area becomes smaller toward the downstream side and the straight formed so as to have a substantially uniform passage cross-sectional area in the merging portion 18. The part 21 and the diffuser part 22 formed so that the passage cross-sectional area becomes larger toward the downstream side are formed in order from the upstream side. For this reason, the exhaust gas ejected from any one of the downstream ends of the independent exhaust passages 16A to 16D first flows into the nozzle portion 20, where it further accelerates (the pressure decreases). And the exhaust gas accelerated by the nozzle part 20 is decelerated when passing the diffuser part 22, and the pressure of exhaust gas recovers in connection with this.

  As described above, when the exhaust gas is ejected at a high speed from the downstream end portion of any one of the independent exhaust passages 16A to 16D toward the nozzle portion 20 of the merging portion 18, A low negative pressure is generated. Therefore, as the exhaust gas is ejected from the independent exhaust passage (any one of 16A to 16D) of a certain cylinder to the merging portion 18, negative pressure acts on the independent exhaust passage and the like of the other cylinders, and the exhaust gas is Sucked downstream (ejector effect).

  Since exhaust gas is sucked out by the ejector effect, scavenging can be promoted. In addition, since exhaust gas discharged from an independent exhaust passage (any one of 16A to 16D) of a certain cylinder is prevented from flowing back to the independent exhaust passage of another cylinder, there is an advantage that exhaust interference is suppressed. .

  Here, in order to ensure scavenging by the ejector effect, when the exhaust valve 6 of a certain cylinder starts to open, the intake valve 5 and the exhaust valve in the cylinder whose exhaust order is one before that cylinder. Both 6 need to be open.

  FIG. 6 is a diagram showing operation timings of the intake and exhaust valves 5 and 6 of the first cylinder 2A to the fourth cylinder 2D set for the purpose of promoting scavenging as described above. As shown in this figure, in any of the first cylinder 2A to the fourth cylinder 2D, the intake valve 5 and the exhaust valve 6 are changed from the end of the exhaust stroke to the start of the intake stroke (from the period O / L in the drawing). Both are open. Hereinafter, a period during which both the intake and exhaust valves 5 and 6 are open, that is, a period O / L in which the valve opening period of the intake valve 5 and the valve opening period of the exhaust valve 6 overlap, This is called an overlap period. On the other hand, when a certain cylinder is in the overlap period (O / L), the exhaust valve 6 starts to open in the cylinder that reaches the exhaust stroke next to the cylinder.

  Specifically, when the first cylinder 2A is in the overlap period, the exhaust valve 6 of the third cylinder 2C that reaches the exhaust stroke next to the first cylinder 2A starts to open. Similarly, during the overlap period of the third cylinder 2C, the exhaust valve 6 of the fourth cylinder 2D that reaches the next exhaust stroke starts to open, and during the overlap period of the fourth cylinder 2D, the next exhaust stroke reaches the next. The exhaust valve 6 of the 2-cylinder 2B starts to open.

  If the operation timing of the intake / exhaust valves 5 and 6 is set as described above, the exhaust gas having the fastest flow rate (so-called blowdown gas) discharged immediately after the exhaust valve 6 of a certain cylinder is opened is the junction 18. The scavenging performance of the cylinder whose exhaust order is one before that of the cylinder is enhanced by the action of the large negative pressure (ejector effect) generated around the jet flow. For example, when the exhaust valve 6 of the first cylinder 2A starts to open and high-speed exhaust gas (blow-down gas) is ejected to the merging portion 18 through the independent exhaust passage 16A, a negative pressure associated therewith is generated from the first cylinder 2A. Also acts on the second cylinder 2B in the previous exhaust order. At this time, since the second cylinder 2B is in the overlap period, the suction of exhaust gas from the exhaust port 4 and the inflow of air from the intake port 3 into the cylinder are performed simultaneously. Thereby, scavenging is sufficiently achieved and the intake amount is increased.

  1 to 4 again, the exhaust device 15 includes first and second communication passages 25A and 25B that communicate the independent exhaust passages of a pair of cylinders whose exhaust order is not continuous. In other words, the first communication passage 25A communicates with the independent exhaust passage 16A of the first cylinder 2A and the independent exhaust passage 16D of the fourth cylinder 2D, which has a relationship in which the exhaust order jumps one with respect to the first cylinder 2A. Is provided. Further, the second communication passage 25B communicates with the independent exhaust passage 16B of the second cylinder 2B and the independent exhaust passage 16C of the third cylinder 2C, which has a relationship in which the exhaust order jumps by one with respect to the second cylinder 2B. Is provided.

  As shown in FIGS. 5A and 5B, the connection angles between the first communication passage 25A and the independent exhaust passages 16A and 16D are respectively θ1 and θ4, and the second communication passage 25A and the independent exhaust passages 16B and 16C. Are set to be the same (θ1 = θ2 = θ3 = θ4). Here, the “connection angle” refers to an angle formed by the axes of two passages connected to each other as illustrated.

  The first and second communication passages 25A and 25B are each provided with a variable valve 26 (FIGS. 3 and 4) that can be opened and closed. These two variable valves 26 comprise butterfly valves having a common (one) rotating shaft 27, and one end of the common rotating shaft 27 is connected to an electric motor 28. The electric motor 28 is electrically connected to the ECU 30 (FIG. 1). When the electric motor 28 is operated in response to a command from the ECU 30, the rotating shaft 27 is rotationally driven. Each variable valve 26 opens and closes the first and second communication passages 25A and 25B, respectively. The two variable valves 26 are both opened and closed via a common rotating shaft 27, so that the opening degrees of both variable valves 26 are always the same.

  FIG. 7 is a diagram for explaining how the opening degree of the variable valve 26 is controlled by the ECU 30. The upper part of FIG. 7 shows the rotational speed and load range in which the engine of this embodiment can be operated, and the lower part of FIG. 7 shows the opening of the variable valve 26 determined based on the rotational speed of the engine. Show. In this figure, Ne0 represents the idling speed, Ne1 represents the maximum speed, and WOT represents the full load line of the engine.

  The opening degree of the variable valve 26 is controlled based on the engine speed. Specifically, when the engine speed is in a relatively low speed range from the idling speed Ne0 to a higher speed Nex ′, the opening degree of the variable valve 26 is set to 0% (fully closed). . On the other hand, on the higher rotation side than the speed range of the opening degree 0%, the opening degree of the variable valve 26 is gradually increased as the rotational speed increases, and when the rotational speed reaches a predetermined value Nex, the opening degree is increased. 100% (fully open) is set. When the engine rotation speed becomes equal to or higher than the predetermined value Nex, the opening degree of the variable valve 26 is maintained at 100% regardless of the rotation speed.

  In other words, the opening degree of the variable valve 26 is maintained at 100% (fully opened) when the engine speed is equal to or higher than the predetermined value Nex, whereas it is higher than 100% when the engine speed is lower than the predetermined value Nex. Is set to a small value, and is set to 0% (fully closed) at the minimum. The predetermined value Nex, which is a threshold value for determining whether or not the opening degree of the variable valve 26 is set to 100%, is set to any value in the range from 4000 rpm to 4500 rpm, for example.

  Further, the opening degree control of the variable valve 26 according to the rotational speed as described above is executed regardless of the engine load. That is, regardless of whether the engine load is high or low, the opening degree control as shown in the lower part of FIG. 7 is executed based only on the engine speed.

  FIGS. 8 and 9 are schematic diagrams for easily explaining the operation of the variable valve 26 by setting the opening degree. Among these, FIG. 8 shows the flow of exhaust gas when the opening degree of the variable valve 26 is set to 0% (fully closed). When the opening degree of the variable valve 26 is 0%, the first and second communication passages 25A and 25B are both completely shut off, so that the exhaust gas discharged from any of the cylinders 2A to 2D Only through the independent exhaust passage (any one of 16A to 16D) connected to the cylinder, the air flows into the downstream junction 18.

  For example, the exhaust gas discharged from the first cylinder 2A flows into the merging portion 18 only through the independent exhaust passage 16A (the leftmost passage in the figure) connected to the first cylinder 2A, as indicated by an arrow X. The exhaust gas discharged from the second cylinder 2B flows into the merging section 18 only through the independent exhaust passage 16B (second passage from the left) connected to the second cylinder 2B, as indicated by the arrow Y. To do. The same applies to the exhaust gas discharged from the third cylinder 2C and the fourth cylinder 2D.

  As described above, when the opening of the variable valve 26 is set to 0% and the flow of exhaust gas is controlled as shown in FIG. 8 (that is, when the flow of exhaust gas does not branch at all), each independent exhaust Since the exhaust gas that has passed through the passages 16A to 16D flows into the merging portion 18 at a high speed, a large negative pressure is generated around the exhaust gas, thereby obtaining the maximum ejector effect.

  On the other hand, as shown in FIG. 9, when the variable valve 26 is opened, the first and second communication passages 25A and 25B are opened, so that the exhaust gas is exhausted through the communication passages 25A and 25B. The gas flow branches. For example, the exhaust gas discharged from the first cylinder 2A flows through the independent exhaust passage 16A (the leftmost passage in the figure) connected to the first cylinder 2A and the flow from the independent exhaust passage 16A. The flow branches to the flow of the arrow X2 flowing into the independent exhaust passage 16D (the rightmost passage) of the fourth cylinder 2D through the one communication passage 25A. Further, the exhaust gas discharged from the second cylinder 2B flows in the direction indicated by the arrow Y1 passing through the independent exhaust passage 16B (second passage from the left) connected to the second cylinder 2B and from the independent exhaust passage 16B. The flow branches to the flow indicated by the arrow Y2 flowing into the independent exhaust passage 16C (second passage from the right) of the third cylinder 2C through the two communication passages 25B. Note that the exhaust gas discharged from the fourth cylinder 2D and the exhaust gas discharged from the third cylinder 2C also have two flows (two independent passages) via the first communication path 25A or the second communication path 25B, respectively. Branches into the flow through the exhaust passage.

  If the exhaust gas discharged from each of the cylinders 2A to 2D branches so as to pass through the two independent exhaust passages (16A and 16D or 16B and 16C) as described above, two exhaust gases from one cylinder can be obtained. Since the gas flows into the merging portion 18 through the downstream end portion of the independent exhaust passage, the inflow speed of the exhaust gas to the merging portion 18 becomes slow, and the ejector effect is weakened. On the other hand, after the exhaust gas branches into two flows, the exhaust gas flow cross-sectional area increases, and the exhaust gas flow resistance (exhaust resistance) is reduced.

  As a matter of course, the above action is maximized when the opening degree of the variable valve 26 is 100% (fully open). That is, when the opening degree is 100%, the ejector effect is greatly weakened while the exhaust resistance becomes the smallest. On the other hand, when the opening degree of the variable valve 26 is lower than 100%, the ejector effect becomes stronger and the exhaust resistance increases as the opening degree becomes lower.

(3) Operational effects and the like As described above, the exhaust system of the in-line four-cylinder engine of the present embodiment includes independent exhaust passages 16A to 16D connected to the exhaust ports 4 of the four cylinders 2A to 2D, respectively. The concentrating portion 17 in which the downstream end portions of the independent exhaust passages 16A to 16D are aggregated, the merging portion 18 connected to the downstream side of the consolidating portion 17, and the independent exhaust passages (16A and 16A 16D, 16B and 16C) are provided with first and second communication passages 25A and 25B communicating with each other, and variable valves 26 respectively provided in the communication passages 25A and 25B. The downstream end portions of the independent exhaust passages 16A to 16D gathered by the gathering portion 17 are formed such that the passage cross-sectional area decreases toward the downstream side, and the intake and exhaust valves 5 of the cylinders 2A to 2D are formed. , 6 is set so that the overlap period (period O / L in FIG. 6) of one cylinder between the cylinders in which the exhaust order is continuous overlaps with the valve opening period of the exhaust valve 6 of the other cylinder. ing. Further, the opening degree of the variable valve 26 is maintained at 100% (fully open) when the engine speed is equal to or higher than the predetermined value Nex, and is reduced to less than 100% when the engine speed is lower than the predetermined value Nex. The According to such a configuration, there is an advantage that it is possible to achieve both the utilization of the ejector effect in the low speed region and the reduction of the exhaust resistance in the high speed region.

  That is, in the above embodiment, the opening degree of each variable valve 26 provided in the plurality of communication passages 25A and 25B is set to 100% in a high speed range where the engine rotation speed is equal to or higher than the predetermined value Nex, and the exhaust order is not continuous. By completely communicating the independent exhaust passages (16A and 16D, 16B and 16C) of a plurality of cylinders, the flow of exhaust gas discharged from one cylinder is branched through the communication passages 25A and 25B. The exhaust gas can be passed through a plurality of independent exhaust passages. Thus, the exhaust resistance can be effectively suppressed and the pump loss can be reduced in the high speed region of the engine where the exhaust gas flow rate is particularly fast and the exhaust resistance tends to increase.

  On the other hand, in the low speed range where the engine rotation speed is lower than the predetermined value Nex, the opening degree of the variable valve 26 is set to be less than 100%, whereby the branch flow of the exhaust gas is suppressed or stopped. The Then, when the exhaust valve 6 starts to open in a certain cylinder (any one of 2A to 2D), the exhaust gas (blowdown gas) exhausted vigorously from the cylinder is connected to an independent exhaust passage ( 16A to 16D) are mainly ejected to the merging portion 18 and a large negative pressure is generated around the ejected flow. At this time, in the cylinder in which the exhaust order is one before the cylinder in which the exhaust valve 6 has started to open, there is an overlap period in which both the intake valve 5 and the exhaust valve 6 are open. The suction of the exhaust gas from the exhaust port 4 and the inflow of air from the intake port 3 are performed simultaneously, thereby promoting scavenging and increasing the intake amount.

  As described above, according to the above-described embodiment, the engine output can be effectively improved by utilizing the ejector effect in the low speed region of the engine to promote scavenging, and the exhaust gas flow rate is particularly high. In a fast engine high speed range, the pump loss can be effectively reduced by diverging the flow of the exhaust gas and suppressing the exhaust resistance.

  In particular, in the above embodiment, in the low speed range where the engine rotation speed is lower than the predetermined value Nex, the lower the rotation speed, the smaller the opening of the variable valve 26 is set to 0%. There is an advantage that it is possible to effectively prevent a sudden change in the engine output accompanying a change in the opening of the variable valve 26 while exhibiting the maximum ejector effect when the rotational speed is low.

  That is, in the above embodiment, when the engine speed is the lowest (in the range from Ne0 to Nex ′ in FIG. 7), the opening of the variable valve 26 is set to 0% and the communication passages 25A and 25B are opened. By completely blocking, scavenging can be effectively promoted by maximizing the ejector effect. On the other hand, an intermediate speed range (Nex ′ to Nex in FIG. 7) between the speed range where the opening degree of the variable valve 26 is set to 0% and the speed range where the opening degree of the variable valve 26 is set to 100%. Unlike the case where the variable valve 26 is suddenly opened and closed by linearly controlling the opening degree of the variable valve 26 according to the engine speed, the engine output accompanying the change in the opening degree of the variable valve 26 is Sudden changes can be effectively prevented.

  In the above embodiment, each variable valve 26 provided in the first and second communication passages 25A and 25B is a butterfly valve having a common rotating shaft 27. The rotating shaft 27 is rotated by the electric motor 28. As a result, each of the variable valves 26 is driven to open and close. According to such a configuration, since the plurality of variable valves 26 can be driven simultaneously by rotating only one rotating shaft 27, each variable valve 26 can be easily and reliably adjusted with the opening degrees of the variable valves 26 matched to each other. It can be opened and closed.

  In the above embodiment, the connection angles θ1 to θ4 between the independent exhaust passages 16A to 16D of the cylinders 2A to 2D and the first and second communication passages 25A and 25B connected thereto are all independent exhaust passages. Are set identically (θ1 = θ2 = θ3 = θ4). According to such a configuration, when the variable valve 26 is opened, it branches from the independent exhaust passage (any one of 16A to 16D) of a certain cylinder to another independent exhaust passage through the communication passage (25A or 25B). Since the ratio of exhaust gas to be discharged is the same regardless of the exhaust gas discharged from any cylinder, it is possible to effectively prevent the exhaust resistance reduction effect from varying from cylinder to cylinder.

  In the above embodiment, the angles θa to θd formed by the axes of the downstream ends of the independent exhaust passages 16A to 16D and the axis of the merging portion 18 are set to be the same for all the independent exhaust passages (θa = θb = θc = θd). According to such a configuration, the inflow angle of the exhaust gas flowing from each cylinder 2A to 2D through the independent exhaust passages 16A to 16D into the merging portion 18 is equal even when the exhaust gas is discharged from any cylinder. Therefore, it is possible to effectively prevent the effect of scavenging promotion due to the ejector effect from varying from cylinder to cylinder.

  In the above embodiment, the opening degree of the variable valve 26 is set to 100% (fully open) in a high speed range where the engine rotational speed is equal to or higher than the predetermined value Nex, and in the low speed range where the rotational speed is lower than the predetermined value Nex, The lower the rotational speed, the lower the opening of the variable valve 26 is gradually reduced (see FIG. 7). If the engine rotational speed is less than the predetermined value Nex, the opening of the variable valve 26 is immediately reduced to 0% ( It may be lowered to (fully closed). However, in this case, since there is a concern that the engine output suddenly changes at the boundary of the predetermined value Nex, the engine output is suddenly changed by, for example, retarding the ignition timing by the spark plug 7 in the vicinity of the predetermined value Nex. Measures such as prevention are desired.

  Further, the opening degree of the variable valve 26 in the high speed range above the predetermined value Nex is not necessarily set to 100% (fully open). In other words, the opening degree of the variable valve 26 in the high speed range may be less than 100% as long as the opening degree allows the exhaust gas to sufficiently branch through the communication passages 25A and 25B.

  In the above embodiment, the control as shown in FIG. 7 for increasing / decreasing the opening degree of the variable valve 26 according to the rotational speed is executed regardless of the engine load. 7 is a high engine speed and high load region where the intake air amount is large and the exhaust gas flow rate is high. Therefore, it is not always necessary to execute the control shown in FIG. 7 in the low load region of the engine. . For this reason, for example, in the low load region of the engine, the variable valve 26 may be set to be fully closed (or a predetermined low opening) regardless of the rotational speed. Alternatively, the ejector effect is useful particularly in the low rotation and high load range, and since a sufficient ejector effect cannot be obtained in the low load range, contrary to the above, in the low load range of the engine, regardless of the rotation speed, The variable valve 26 may be set to fully open (or a predetermined high opening). In any case, the control for increasing or decreasing the opening degree of the variable valve 26 according to the rotational speed is executed only in the high load region of the engine.

  Further, in the above embodiment, regardless of the engine operating region, as shown in FIG. 6, the overlap period of one cylinder between the cylinders in which the exhaust order is continuous (the intake valve 5 and the exhaust valve 6 are both opened). The operation timing (valve timing) of the intake / exhaust valves 5 and 6 of each cylinder 2A to 2D is set so that the period O / L) overlaps with the opening period of the exhaust valve 6 of the other cylinder. In the case of an engine having a variable mechanism that can change the valve timing, the valve timing is set as shown in FIG. 6 only in the high load region of the engine where the intake air amount increases, and in the low load region of the engine, The valve timing may be controlled so that there is no overlap period.

  Moreover, in the said embodiment, although the example which applied the exhaust apparatus of this invention to the inline 4-cylinder engine was demonstrated, if the engine which can apply the exhaust apparatus of this invention can obtain the same effect, It is not limited to an in-line four-cylinder engine, and may be a multi-cylinder engine having four or more cylinders.

2A to 2D Cylinder 3 Intake port 4 Exhaust port 5 Intake valve 6 Exhaust valve 15 Exhaust device 16A to 16D Independent exhaust passage 17 Concentration portion 18 Merge portion 25A First communication passage (communication passage)
25B Second communication path (communication path)
26 Variable valve 27 Rotating shaft 30 ECU (control means)

Claims (5)

An exhaust system provided in a multi-cylinder engine having four or more cylinders and having an intake port, an exhaust port, an intake valve, and an exhaust valve provided in each cylinder,
Four or more independent exhaust passages individually connected to the exhaust ports of the cylinders;
An aggregating portion aggregated so that the downstream ends of the independent exhaust passages are close to each other;
A merging portion that is connected to the downstream side of the aggregating portion and has a common space communicating with all of the independent exhaust passages;
Two or more communication passages communicating with independent exhaust passages of a plurality of cylinders whose exhaust order is not continuous;
A variable valve that can be opened and closed provided in each of the communication paths;
Control means for controlling each of the variable valves,
The downstream end portion of each independent exhaust passage aggregated in the aggregation portion is formed so that the cross-sectional area of the passage becomes smaller toward the downstream side,
At least in the high load region of the engine, the opening period of the intake valve and the exhaust valve of each cylinder overlaps with a predetermined overlap period, and the overlap period of one cylinder between the cylinders in which the exhaust sequence is continuous is the other cylinder The operation timing of the intake valve and the exhaust valve of each cylinder is set so as to overlap with the opening period of the exhaust valve of
The control means maintains the opening of each variable valve at a predetermined high opening when the engine rotation speed is equal to or higher than a predetermined value at least in a high engine load range, and the engine rotation speed is lower than the predetermined value. In some cases, the opening of each variable valve is set to a value smaller than the predetermined high opening. The exhaust system for a multi-cylinder engine according to claim 1,
When the engine rotational speed is lower than the predetermined value, at least in the high load region of the engine, the control means reduces the opening of each variable valve as the rotational speed is low, and sets the minimum to 0%. An exhaust system for a multi-cylinder engine. The exhaust system for a multi-cylinder engine according to claim 1 or 2,
Each of the variable valves comprises a butterfly valve having a common rotating shaft,
An exhaust system for a multi-cylinder engine, wherein the control means drives the variable valves to open and close by rotating the common rotating shaft. The exhaust device for a multi-cylinder engine according to any one of claims 1 to 3,
An exhaust system for a multi-cylinder engine, wherein the connection angle between each independent exhaust passage and the communication passage connected thereto is set to be the same for all the independent exhaust passages. The exhaust device for a multi-cylinder engine according to any one of claims 1 to 4,
An exhaust system for a multi-cylinder engine, characterized in that the angle formed by the axis of the downstream end of each independent exhaust passage and the axis of the merging portion is set to be the same for all independent exhaust passages.

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