JP4807343B2 - Engine supercharger - Google Patents

Engine supercharger Download PDF

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JP4807343B2
JP4807343B2 JP2007266533A JP2007266533A JP4807343B2 JP 4807343 B2 JP4807343 B2 JP 4807343B2 JP 2007266533 A JP2007266533 A JP 2007266533A JP 2007266533 A JP2007266533 A JP 2007266533A JP 4807343 B2 JP4807343 B2 JP 4807343B2
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exhaust
valve
independent
variable
exhaust valve
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JP2009097335A (en
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直之 山形
幹公 藤井
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マツダ株式会社
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    • YGENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y02TECHNOLOGIES OR APPLICATIONS FOR MITIGATION OR ADAPTATION AGAINST CLIMATE CHANGE
    • Y02TCLIMATE CHANGE MITIGATION TECHNOLOGIES RELATED TO TRANSPORTATION
    • Y02T10/00Road transport of goods or passengers
    • Y02T10/10Internal combustion engine [ICE] based vehicles
    • Y02T10/14Technologies for the improvement of mechanical efficiency of a conventional ICE
    • Y02T10/144Non naturally aspirated engines, e.g. turbocharging, supercharging

Description

  The present invention relates to an engine supercharging device, and more particularly to an engine that increases engine torque in a low rotation range.

  As a means for increasing the output torque of the engine, a supercharging device for increasing the intake pressure is known. A typical example is an exhaust turbocharger (hereinafter abbreviated as a turbocharger). A turbocharger is a turbine wheel (hereinafter abbreviated as a turbine) provided in an exhaust passage and a compressor wheel (hereinafter abbreviated as a compressor) provided in an intake passage by a shaft. The compressor is driven by rotating the turbine to increase the intake pressure.

  A turbocharger is characterized in that it can not increase engine torque in a wide rotation range from a low rotation region to a high rotation region, while it can efficiently obtain a high supercharging pressure. Generally, a small turbocharger increases torque in a low rotation region, and a large turbocharger increases torque in a high rotation region. Therefore, when a turbocharger is provided, it is necessary to select a turbocharger of a type suitable for the torque characteristics required for the engine.

  However, in many cases, it is desired to increase the engine torque in a wide rotation range from low rotation to high rotation. Therefore, for example, one having two turbochargers for low rotation and high rotation (so-called two-stage turbo), one having an electric supercharger for low rotation and a turbocharger for high rotation, or As shown in Patent Document 1, there has been proposed a turbine nozzle provided with a movable flap, and in a low rotation region, the flap opening degree is reduced to increase the supercharging efficiency (so-called variable geometry turbo).

In the case of a variable geometry turbo as shown in Patent Document 1, it is desirable to employ a large turbo as a base turbocharger. As a result, in the high rotation region, the flap opening is increased to increase the engine torque as an original characteristic of the large turbo, and in the low rotation region, the flap opening is decreased to increase the exhaust flow velocity. It is possible to increase the engine torque by increasing the turbine driving force. Eventually, the engine torque can be increased in a wide rotation range from low rotation to high rotation.
JP-A-9-112285

  However, each of the conventional turbochargers including the variable geometry turbo shown in Patent Document 1 has a problem in that the structure is complicated or the size is increased.

  The present invention has been made in view of the above circumstances, and an object of the present invention is to provide an engine supercharging device capable of increasing engine torque in a low rotation region while having a simple structure.

The invention according to claim 1 for solving the above-described problems is an exhaust manifold connected to an exhaust port of each cylinder and having a plurality of independent exhaust passages, and the independent exhaust passages gathered at the exhaust manifold or downstream thereof. A collecting portion, an exhaust turbocharger connected to the downstream side of the collecting portion, a variable exhaust valve capable of changing each cross-sectional area of the independent exhaust passage on the upstream side of the collecting portion, and the variable exhaust Variable exhaust valve control means for driving and controlling the valve; and exhaust valve timing changing means capable of changing the opening and closing valve timing of the exhaust valve for opening and closing the exhaust port of each cylinder , wherein the variable exhaust valve control means includes at least the above-mentioned in the low-speed region than the engine rotational speed wastegate valve of the turbocharger starts to open, the independent exhaust by the variable exhaust valve The respective cross-sectional area of the road is reduced from the maximum area, and the reduction degree of the respective cross-sectional area to perform an independent exhaust throttle control is increased the lower the engine speed, the exhaust valve timing changing means, An engine supercharging device characterized in that when the independent exhaust throttle control is executed, the opening timing of the exhaust valve is delayed in a range before exhaust bottom dead center with respect to the non-execution time .

The invention according to claim 2 is the supercharging device according to claim 1, wherein the engine, by changing at least one of the opening and closing valve timing of the intake and exhaust valves, also open both the intake valve and the said exhaust valve Valve timing changing means capable of changing the amount of overlap as a period, and the valve timing changing means expands the overlap amount when the independent exhaust throttle control is executed and when it is not executed. To do.

The invention according to claim 3, in the supercharging device according to claim 2, wherein the engine, the variable exhaust valve control means executes the independent exhaust throttle control even in the natural suction region, the valve timing changing means, Nature The overlap amount is increased when the independent exhaust throttle control is executed in the intake region, compared to when it is not executed.

  According to the first aspect of the present invention, as described below, the engine torque in the low rotation region can be increased while having a simple structure.

  First, according to the configuration of the present invention, the exhaust ejector effect can be obtained by the independent exhaust passage and the variable exhaust valve of the exhaust manifold. The ejector is a conventionally known device called a jet pump, and sucks out a fluid to be sucked out by a negative pressure by a high-speed driving fluid. For example, it is used as a vacuum pump by utilizing the sucking action. In the present specification, the exhaust suction action obtained by the same principle as the ejector is referred to as an ejector effect.

  Explaining the ejector effect according to the configuration of the present invention, exhaust, particularly blowdown gas (strong exhaust immediately after the exhaust valve is opened) flows through an independent exhaust passage in the exhaust manifold, which is a variable exhaust valve. When passing through a portion where the passage cross-sectional area is reduced, the flow velocity is increased and the pressure is reduced. This throttled exhaust corresponds to the driving fluid.

  On the other hand, in the collecting portion, an exhaust passage through which exhaust corresponding to the driving fluid flows and another exhaust passage communicate with each other. Therefore, in the gathering portion, the exhaust as the driving fluid sucks the exhaust (sucked fluid) in the other exhaust passage by the ejector effect.

  In order to further enhance the ejector effect, the exhaust of the driving fluid and the exhaust of the sucked fluid may be merged at a shallowest angle (an angle close to parallel) as much as possible.

  The advantages of the ejector effect mainly include the following three points.

  The first is an increase in the turbine flow rate of the turbocharger (the amount of exhaust gas supplied to the turbocharger). The exhaust of the driving fluid and the exhaust of the sucked fluid merge at the gathering portion and are introduced into the turbine of the turbocharger provided downstream thereof. Accordingly, the turbine flow rate is increased by the amount of exhausted air as compared with the case where the ejector effect is not provided. Thus, the turbine driving force can be increased and the supercharging pressure can be increased.

  Secondly, the exhaust gas scavenging is promoted. Since the exhaust as the sucked fluid is sucked out by the ejector effect, scavenging is promoted. Thereby, the exhaust resistance is reduced. Further, normally, when shifting from the exhaust stroke to the intake stroke, a period (overlap period) in which both the intake valve and the exhaust valve are opened is provided. Since the intake of the overlap period is also promoted by the promotion of the scavenging, the intake amount can be increased and the engine torque can be increased.

  In order to obtain this effect more, as in a normal 4-cycle 4-cylinder engine, when a certain cylinder is blown down (immediately after the exhaust valve is opened), the overlap period of other cylinders overlaps. It is good to be. However, the exhaust passages between the cylinders in that relationship need to be independent from each other upstream of the variable exhaust valve.

  Thirdly, in the case of performing dynamic pressure supercharging, it is the promotion. Here, first, dynamic pressure supercharging will be described. In the dynamic pressure supercharging, the supercharging capability of the turbocharger is increased by utilizing the pulsation of the exhaust gas. Although the detailed mechanism will be described later, dynamic pressure supercharging is promoted as the exhaust pulsation increases. In order to increase the exhaust pulsation, it is the simplest and most effective to reduce the exhaust passage volume. However, in terms of layout, there is a limit to reducing the exhaust manifold volume by reducing the entire volume of the exhaust manifold.

  In the normal case where there is no ejector effect, the exhaust gas flows around (reverses) to another exhaust passage at the collecting portion. However, if there is an ejector effect as in the present invention, the exhaust sucks the driven fluid from the other exhaust passages as the driving fluid. That is, it does not wrap around other exhaust passages. This brings about the effect of reducing the exhaust passage volume in the dynamic pressure supercharging.

  As described above, if the entire exhaust passage volume (exhaust manifold volume) is the same, the configuration of the present invention having the ejector effect can further promote the dynamic pressure supercharging as compared with the case without the ejector effect.

  In the configuration of the present invention, it is desirable to employ a large turbocharger as the turbocharger. As a result, it is possible to increase the torque as an original characteristic of the large turbo in the high rotation region, and to increase the torque in the low rotation region by the ejector effect or the like. Eventually, the engine torque can be increased in a wide rotation range from the low rotation range to the high rotation range.

Further , according to the present invention , as will be described below, an effective ejector effect can be obtained in an operation region that is necessary and sufficient and has less harmful effects.

  The effect of increasing the supercharging pressure due to the ejector effect is substantially effective in a region where the rotational speed is lower than the rotational speed at which the wastegate valve of the turbocharger starts to open (so-called intercept point). In the higher rotation range, it is necessary to open the wastegate valve to prevent the boost pressure from becoming excessive, that is, to suppress the increase in boost pressure, so there is no need to increase the boost pressure due to the ejector effect. It becomes. On the contrary, if the independent exhaust throttle control is performed at a higher rotational speed than the intercept point, it may cause an increase in exhaust resistance.

  Therefore, according to the present invention, by executing the independent exhaust throttle control in the rotation region lower than the intercept point, an effective ejector effect can be obtained in the operation region that is necessary and sufficient and has less harmful effects.

  Further, even in the low rotation range where the independent exhaust throttle control is executed, the lower the rotation speed, the greater the request for increasing the supercharging pressure. On the other hand, the ejector effect increases as the exhaust passage cross-sectional area is reduced (squeezed). Therefore, the effect of increasing the boost pressure according to the required degree can be obtained by reducing the exhaust passage cross-sectional area as the engine speed increases and the lower the rotation speed.

Furthermore, according to the present invention, as described below, a more remarkable ejector effect can be obtained.

  In the normal case where the independent exhaust throttle control is not executed, the exhaust valve starts to open relatively early before the exhaust stroke bottom dead center (for example, 40 to 60 ° CA before bottom dead center; “° CA” indicates the crank angle). . By doing so, scavenging is promoted, but on the other hand, since the exhaust starts while the piston descends, the momentum of the blowdown gas is weakened. This is disadvantageous for independent exhaust throttle control that uses blowdown gas as the drive fluid for the ejector effect.

  Therefore, according to the present invention, when the independent exhaust throttle control is executed, the opening timing of the exhaust valve is delayed, so that it is possible to suppress a decrease in the momentum of the blowdown gas. In addition, after the bottom dead center, the action of pushing up the exhaust is added, so that the blowdown gas can be energized. That is, the ejector effect can be obtained more remarkably.

However, if the exhaust valve is opened after exhaust bottom dead center, the exhaust resistance increases. Therefore, even if delay the opening timing of the exhaust valve, by delaying a range of pre-exhaust bottom dead center, immediately before exhaust bottom dead center you to open the exhaust valve.

According to the invention of claim 2 , as described below, a more remarkable ejector effect can be obtained.

  As described above, according to the independent exhaust throttle control of the present invention, scavenging is promoted by the ejector effect, and intake air in the overlap period is promoted, so that the intake air amount can be increased and the engine torque can be increased. . Therefore, according to the present invention, when the independent exhaust throttle control is executed, the above effect can be obtained more significantly by increasing the overlap amount (the length of the overlap period) by the valve timing changing means.

  Normally, if the overlap amount is inadvertently increased, the exhaust gas may flow backward due to the intake negative pressure. However, in the independent exhaust throttling control of the present invention, the exhaust is sucked downstream by the ejector effect, so that such a backflow hardly occurs. That is, it is possible to extend the overlap period while suppressing the adverse effect of the backflow of exhaust.

  Further, the valve timing changing means for increasing the overlap amount may be one that advances the valve opening timing of the intake valve, delays the valve closing timing of the exhaust valve, or both. As a valve timing changing means, using a general VVT (which changes the opening / closing valve timing in a translational manner without changing the valve opening period), and at least setting the exhaust valve closing timing to be delayed, The opening timing of the exhaust valve is automatically delayed. That is, this can also be used as the exhaust valve timing changing means.

According to the third aspect of the present invention, as described below, the response of the overlap expansion can be enhanced at the time of transition from the natural intake region to the supercharging region.

  In general, those equipped with valve timing changing means are set such that the amount of overlap increases as the load increases and the rotation speed increases. Therefore, for example, even when the engine speed is the same, the overlap is relatively small in the natural intake region (low load region), and the overlap is large in the supercharging region (high load region).

  Therefore, for example, when the load increases due to an acceleration request and a rapid transition is made from the low load region to the high load region, the valve timing changing means needs to increase the overlap amount accordingly. If the increase amount of the overlap amount is large, the time required for changing the valve timing becomes long, and there is a concern that the responsiveness is lowered.

In particular, in the present invention, the independent exhaust throttle control is executed in a predetermined low rotation region of the supercharging region, and control is performed to increase the overlap amount at that time ( Claim 2 ). For this reason, the increase amount of the overlap amount tends to be larger, and there is an increasing concern about responsiveness reduction.

  Therefore, according to the present invention, the independent exhaust throttle control is executed even in the natural intake region, and the overlap amount is increased. By doing so, an increase in the overlap amount is reduced, so that the response delay can be suppressed.

  Hereinafter, preferred embodiments of the present invention will be described with reference to the accompanying drawings.

  FIG. 1 is a schematic configuration diagram of an engine supercharging device according to a first embodiment of the present invention. 2 is a partial side sectional view of FIG.

  The supercharger for an engine according to the present embodiment has a simple configuration using one turbocharger 50, but obtains high supercharging performance in a wide range from a low rotation region to a high rotation region, and is high in the entire region. The greatest feature is that torque can be generated.

As means for achieving this, the following two technical features are mainly provided.
(1) Improvement of supercharging capability by dynamic pressure supercharging (2) Independent exhaust throttle control using independent exhaust passage and variable exhaust valve 30 These will be described in detail later. First, the configuration and structure of this embodiment will be described. To do.

  The engine 1 is an in-line four-cylinder four-cycle engine. In the cylinder block 2, first to fourth cylinders 3a, 3b, 3c, 3d (referred to collectively as cylinders 3) are arranged on a straight line. The configuration of each cylinder 3 is common, and as shown in FIG. 2, an intake port 6 for sucking intake Wi and an exhaust port 8 for discharging exhaust We are provided above the combustion chamber 4. An intake valve 7 for opening and closing the intake port 6 and an exhaust valve 9 for opening and closing the exhaust port 8 are provided. Further, a spark plug 5 for generating a spark is provided at the top of the combustion chamber 4. In addition, an unillustrated fuel supply means (such as a fuel injection valve) is provided at an appropriate position.

  First to fourth exhaust passages 16a, 16b, 16c, and 16d are connected to the exhaust port 8 of each cylinder 3. As shown in FIG. 1, the second exhaust passage 16b and the third exhaust passage 16c are gathered on the downstream side to form an auxiliary collective exhaust passage 16bc. The first to fourth exhaust passages 16a to 16d and the auxiliary collective exhaust passage 16bc constitute an exhaust manifold 16 as a whole.

  That is, the exhaust manifold 16 has four independent exhaust passages (first to fourth cylinders 3a, 3b, 3c, 3d) on the upstream side, and three independent exhaust passages (first, fourth exhaust passages 16a, 16d, and The auxiliary collective exhaust passage 16bc). Unless otherwise specified below, the independent exhaust passage refers to three independent exhaust passages on the downstream side.

  A housing 31 is connected to the downstream side of the exhaust manifold 16. A variable exhaust valve 30 is provided on the inner upstream side of the housing 31 (near the connection portion with the exhaust manifold 16). Further, the housing 31 forms a collective portion 31 c where the three independent exhaust passages gather on the downstream side of the variable exhaust valve 30.

  The variable exhaust valve 30 is a valve that changes the cross-sectional areas of the three independent exhaust passages 16a, 16bc, and 16d while maintaining their independent states, and includes a flap 35, an actuator 38 that drives the flap 35, and the like. . The detailed structure will be described later with reference to FIG.

  As shown in FIG. 1, a housing 51 of a turbocharger 50 (exhaust turbine supercharger), specifically a turbine nozzle thereof, is connected to the downstream side of the housing 31 (collecting portion 31c). The exhaust turbine supercharger 50 is a well-known supercharger, in which a turbine 54 provided in the exhaust passage 60 and a compressor 52 provided in the intake passage are connected by a shaft 53. , The compressor 52 is driven to compress the intake air Wi to increase the intake pressure.

  The exhaust passage 60 is provided with a waist passage 61 that bypasses the turbine 54 and a waste gate valve 62 that opens and closes the waist passage 61.

  The turbocharger 50 according to the present embodiment is a large turbo with a strong torque increasing action mainly in a high rotation region. Generally, A / R (ratio between the nozzle area A of the turbine section shown in FIG. 2 and the distance R from the turbine shaft to the nozzle center section) is relatively large, and the turbine diameter D is also relatively large. It is called turbo. The turbocharger 50 of the present embodiment is set to have a relatively small A / R, although the turbine diameter D is the same as that of a general large turbocharger.

  As shown in FIG. 1, the engine 1 is provided with a variable valve timing mechanism 12 (valve timing changing means). The variable valve timing mechanism 12 according to the present embodiment is a so-called VVT (Variable Valve Timing) in which the valve timing (valve opening / closing valve timing) is moved back and forth in parallel while maintaining the valve opening periods of the intake valve 7 and the exhaust valve 9. It is. As a VVT system, the valve timing may be continuously changed or may be changed in two or more steps.

  The variable valve timing mechanism 12 of the present embodiment includes an intake side intake VVT 12i (intake valve timing changing means) and an exhaust side exhaust VVT 12e (exhaust valve timing changing means), and both the intake valve 7 and the exhaust valve 9 are provided. The valve timing is changed at.

  As shown in FIG. 1, an ECU (Engine Control Unit) 20 that electrically controls the operation of the engine 1 is provided. The ECU 20 is a control unit composed of a microprocessor having a CPU, a memory, a counter timer group, an interface, a bus connecting these units, and the like. The ECU 20 performs drive control of the variable valve timing mechanism 12 and opening / closing control of the wastegate valve 62 in addition to general combustion control such as fuel supply amount, throttle opening, or ignition timing.

  Further, the ECU 20 also functions as variable exhaust valve control means for driving and controlling the variable exhaust valve 30. Specifically, the ECU 20 controls each of the independent exhaust passages 16a, 16bc, and 16d by the variable exhaust valve 30 at least in a predetermined low rotation region of the supercharging region (a rotation region lower than the intercept point at which the wastegate valve 62 starts to open). Independent exhaust throttle control is performed to reduce the cross-sectional area more than when the maximum area (variable exhaust valve 30 is open).

  Further, the engine 1 of the present embodiment, like a general four-cylinder engine, shifts each stroke so that each cylinder 3 sequentially reaches the ignition timing at every crank angle of 90 degrees (hereinafter referred to as 90 ° CA). Driving. The ignition order is so-called # 1 → # 3 → # 4 → # 2 (#x indicates the x-th cylinder). Table 1 shows the transition of the stroke of each cylinder 3.

  In Table 1, each row represents the first cylinder 3a to the fourth cylinder 3d, and each column represents the transition of the stroke every 90 ° CA. As shown in Table 1, for example, when the first cylinder 3a is in the expansion stroke, the second cylinder 3b is in the exhaust stroke, the third cylinder 3c is in the compression stroke, and the fourth cylinder 3d is in the intake stroke.

  In the state shown in FIG. 2, the first cylinder 3a is in a transition period (near the bottom dead center) from the expansion stroke to the exhaust stroke. At this time, the exhaust valve 9 is opened and the exhaust We begins to be discharged from the combustion chamber 4 to the exhaust port 8 (blowdown).

  As shown in Table 1, the second cylinder 3b is in a transition period (near top dead center) from the exhaust stroke to the intake stroke at that time. In this transition period, as shown in the figure, a period during which both the intake valve 7 and the exhaust valve 9 are open, that is, a so-called overlap period is provided.

  FIG. 3 is an external perspective view of the exhaust manifold 16 and the housing 31. 4 is a partial perspective view of the exhaust manifold 16 on the downstream side. FIG. 5 is a perspective view of a main part of the variable exhaust valve 30. FIG. 6 is a longitudinal sectional view of the exhaust manifold 16 and the housing 31 and shows a state in which the variable exhaust valve 30 is in an open state. FIG. 7 is a cross-sectional view similar to FIG. 6 and shows a state where the variable exhaust valve 30 is in an open state. 8 is a cross-sectional view taken along line VIII-VIII in FIG. Hereinafter, the exhaust manifold 16 and the housing 31, particularly the variable exhaust valve 30 in the housing 31 will be described with reference to these drawings.

  As shown in FIG. 3, a flange 16 e is provided at the upstream end of the exhaust manifold 16 and is fixed to a cylinder head (not shown) of the engine 1. Four exhaust passages, that is, first, second, third and fourth exhaust passages 16a, 16b, 16c and 16d are connected to the flange 16e. In the flange 16e portion, each exhaust passage has a circular cross section of φ36 mm.

  The first exhaust passage 16a is connected to the exhaust port 8 of the first cylinder 3a disposed on one end side of the cylinders 3 arranged in a row. The fourth exhaust passage 16d is connected to the exhaust port 8 of the fourth cylinder 3d disposed on the other end side. The second exhaust passage 16b and the third exhaust passage 16c are connected to the exhaust ports 8 of the second cylinder 3b and the third cylinder 3c on the center side, respectively.

  As shown in FIG. 4, the first exhaust passage 16a and the fourth exhaust passage 16d maintain an independent state over their entire lengths, but the second exhaust passage 16b and the third exhaust passage 16c are located immediately before their downstream ends. So as to form an auxiliary collective exhaust passage 16bc. Accordingly, three independent exhaust passages (first exhaust passage 16a, auxiliary collective exhaust passage 16bc, and fourth exhaust passage 16d) are formed near the downstream end of the exhaust manifold 16. These are arranged in parallel at a shallow angle (preferably substantially parallel) so that the auxiliary exhaust passage 16bc is sandwiched between the first exhaust passage 16a and the fourth exhaust passage 16d.

The three independent exhaust passages are opened at a manifold outlet 17 which is a downstream end of the exhaust manifold 16. That is, the first opening 17a of the first exhaust passage 16a, the auxiliary collection opening 17bc of the auxiliary collection exhaust passage 16bc, and the fourth opening 17d of the fourth exhaust passage 16d are arranged in a straight line in this order. The opening areas of the openings 17a, 17bc, and 17d are configured to be substantially equal to each other within a range of about 380 to 616 mm 2 . This area corresponds to a circular cross-sectional area of φ22 to φ28 mm. Moreover, this is 37 to 61% in area ratio with respect to the area (equivalent to φ36 mm) of the manifold inlet portion of each exhaust passage 16a, 16b, 16c, and 16d.

  The first exhaust passage 16a and the fourth exhaust passage 16d, and the second exhaust passage 16b and the third exhaust passage 16c are symmetrical to each other. Accordingly, the first exhaust passage length La and the fourth exhaust passage length Ld are substantially equal. In the present embodiment, the first exhaust passage length La is configured to be 200 mm or less.

  Further, the first passage volume Va and the fourth passage volume Vd are substantially equal to each other, and are also substantially equal to the auxiliary collective passage volume Vbc (including a part of the second exhaust passage 16b alone and a part of the third exhaust passage 16c alone). It is configured.

  As shown in FIG. 6, a housing 31 is connected to the manifold outlet 17 of the exhaust manifold 16. The housing 31 functions as a valve housing that supports and stores the flap 35 of the variable exhaust valve 30 on the upstream side, and the exhaust We from the independent exhaust passages 16a, 16bc, and 16d merges on the downstream side. A collective portion 31c is formed.

  Here, with reference to FIG. 5, the principal part (operation part) of the variable exhaust valve 30 will be described. The variable exhaust valve 30 is provided in a direction intersecting with the flow of the exhaust We, and a flap shaft 37 supported by the housing 31, a flap 35 that can turn around the flap shaft 37, and a control signal (variable) from the ECU 20. An actuator 38 (motor or the like) that rotates the flap shaft 37 based on the opening degree command of the exhaust valve 30), and a return spring 39 that biases the flap 35 in the valve opening direction.

  The flap 35 has a fan-shaped surface 36 having a fan-shaped cross section in which the flap shaft 37 is a key of the fan as viewed from the flap shaft 37. The inside of the fan-shaped surface 36 is a hollow to reduce the weight.

  As shown in FIG. 6, the housing 31 is formed with a bulging portion 31b bulging upward, and the state in which the flap 35 is stored inside the bulging portion 31b (the state shown in FIG. 6) is variable exhaust. The valve 30 is open (fully open). When the variable exhaust valve 30 is fully open, as shown in FIG. 6, the exhaust We introduced into the housing 31 from the manifold outlet 17 is guided to the collecting portion 31c without being throttled by the flap 35 (variable exhaust valve 30).

  On the other hand, the state in which the flap 35 is rotationally driven about the flap shaft 37 and enters most inside the bulging portion 31b (the state shown in FIG. 7) is the closed (fully closed) state of the variable exhaust valve 30. The opening degree of the flap 35 is appropriately adjusted between the fully closed state and the fully open state by the actuator 38.

  When the variable exhaust valve 30 is fully closed, as shown in FIG. 7, the fan-shaped surface 36 of the flap 35 blocks a part of the flow path, so that the exhaust passage sectional area is reduced. Therefore, the exhaust We introduced into the housing 31 from the manifold outlet 17 is throttled by the variable exhaust valve 30, and then guided to the collecting portion 31c.

  As shown in FIGS. 6 and 7, a partition plate 32 is provided on the upstream side of the housing 31. Two partition plates 32 are erected along (in parallel with) the flow of the exhaust gas We, and are provided in two spaced apart in the direction of the flap shaft 37. One of the two partition plates 32 is erected so as to continue from the wall surface that partitions the first opening 17a and the auxiliary assembly opening 17bc at the mating portion with the manifold outlet 17, and partitions the inside of the housing 31. Is erected so as to continue from the wall surface that partitions the auxiliary assembly opening 17bc and the fourth opening 17d, and partitions the inside of the housing 31. In other words, in the section where the exhaust gas We flows along the partition plate 32, the independent state and the parallel state of the independent exhaust passages 16a, 16bc and 16d are maintained by the two partition plates 32.

  Each rear edge 32a of each partition plate 32 is formed along the fan-shaped surface 36 of the flap 35 when the variable exhaust valve 30 is in the closed state. Therefore, when the exhaust gas We is throttled by the flap 35, the exhaust gas We is throttled while the independent state and the parallel state are maintained.

  The collecting portion 31 c is formed on the downstream side of the rear edge 32 a of the partition plate 32.

  A flange 31 a is provided on the downstream end side of the housing 31 and is joined to the housing 51 of the turbocharger 50. For convenience of the layout of the turbocharger 50, the housing 31 is bent downward halfway. Such bending is unnecessary depending on the installation position of the turbocharger 50. Different bending angles may also be used.

  As shown in FIGS. 6 to 8, a flow guide plate 33 is provided in the housing 31 along the bending direction of the housing 31. The flow guide plate 33 guides the exhaust gas We passed through the partition plate 32 so as to smoothly flow along the bend of the housing 31. In particular, as shown in FIG. 7, the flow guide plate 33 is arranged so that the housing 31 and the flow guide plate 33 surround the exhaust gas We that has passed through the partition plate 32 when the variable exhaust valve 30 is closed. .

  Further, as shown in FIGS. 6 to 8, a rectifying guide 34 is provided on the collective portion 31 c in the housing 31 so as to stand inward from the bent outer wall surface of the housing 31. The two rectifying guides 34 are provided upright (in parallel) along the flow of the exhaust gas We, and are provided in two spaced apart in the direction of the flap shaft 37. Each rectifying guide 34 is provided on substantially the same plane as each partition plate 32. A gap is provided between the rectifying guide 34 and the flow guide plate 33 on the inner side in the bending direction of the housing 31. As will be described in detail later, the rectifying guide 34 is provided to restrict the swirling flow in the direction intersecting the exhaust direction (the direction not parallel to the paper surface of FIGS. 6 and 7).

  Although the schematic configuration of the present embodiment has been described above, the advantage that the flap 35 of the variable exhaust valve 30 has a fan-shaped surface 36 (hereinafter referred to as a fan shape) will be described.

  FIG. 9 is an explanatory view of the fluid force acting on the flap of the variable exhaust valve. FIG. 9A shows a case where a fan-shaped flap 35 is used as in this embodiment, and FIG. This is shown for comparison with the present embodiment.

  First, the variable exhaust valve 130 of FIG. 9B will be described. The variable exhaust valve 130 has a plate-shaped flap plate 135, and one side thereof is supported by a flap shaft 137 so as to be swingable. This is a conventional general structure in which a fluid passage is narrowed using a flap.

  In this case, when the exhaust We hits the flap plate 135, the fluid force (exhaust pressure Pe) acts perpendicularly to the plate surface of the flap plate 135. As a result, a rotational moment Me around the flap shaft 137 is generated on the flap plate 135. If the pulsation of the exhaust We is large, the rotational moment Me varies greatly according to the pulsation. In particular, in this embodiment, since the pulsation of the exhaust gas We is positively increased in order to perform dynamic pressure supercharging described later, the fluctuation amount (pulsation amplitude) of the rotational moment Me is particularly large.

  If the fluctuation amount of the rotational moment Me is large, the flap plate 135 is likely to flutter and the operation becomes unstable, and means for preventing it is necessary. For example, it is necessary to increase the size of the actuator 38 to withstand a high load, or to increase the size of the actuator 38 in order to increase the set load of the return spring 39.

  On the other hand, the variable exhaust valve 30 of this embodiment shown in FIG. In this case as well, the exhaust pressure Pe acts on the fan-shaped surface 36 of the flap 35. However, since this exhaust pressure Pe acts perpendicularly to the fan-shaped surface 36, that is, in the radial direction of the flap shaft 37, the rotational moment Me is theoretically not generated (the radial load Fe acts on the flap shaft 37 as a resultant force). Therefore, even if the exhaust pulsation is large, the flap 35 hardly flutters, and it is not necessary to increase the size of the actuator 38 and the return spring 39 to prevent this. As a result, the variable exhaust valve 30 can be reduced in size while ensuring reliable operation.

  Next, main technical features of the present embodiment introduced at the beginning will be described.

(1) Improvement of supercharging capacity by dynamic pressure supercharging First, dynamic pressure supercharging will be described. The dynamic pressure supercharging increases the supercharging capability of the turbocharger 50 by using exhaust pulsation as described below.

  FIG. 10 is a turbine characteristic diagram of the turbocharger 50. The horizontal axis represents the turbine flow rate Qt (kg / s), and the vertical axis represents the turbine driving force Ft (kW). Normally, pulsation also occurs in the turbine flow rate Qt due to exhaust pulsation. The turbine flow rate Qt and the turbine driving force Ft shown here are an instantaneous flow rate and an instantaneous driving force that change momentarily due to the pulsation. Hereinafter, the principle of dynamic pressure supercharging will be described with reference to FIG.

  As indicated by the characteristic C11, the turbine driving force Ft increases as the turbine flow rate Qt increases. The increase rate is not constant (linear), and increases as the turbine flow rate Qt increases. As a result, the characteristic C11 becomes a downwardly curved characteristic as shown in the figure. However, FIG. 10 shows the degree of curvature exaggerated more than the actual degree for convenience of explanation.

  The characteristic C101 is a virtual characteristic shown for comparison with the characteristic C11, and is a characteristic in which the relationship between the turbine flow rate Qt and the turbine driving force Ft is in a proportional relationship (linear).

  Here, the case where the pulsation of the turbine flow rate Qt is small (peak flow rate q1) and the case where it is large (peak flow rate q2) will be considered. In both cases, the time-averaged turbine flow rate (for example, per 180 ° CA) is assumed to be the same. In that case, an effective exhaust time per one exhaust stroke (hereinafter referred to as a blow-down period) is shorter when the pulsation of the turbine flow rate Qt is larger than when the pulsation is small (see FIG. 11).

  First, the linear characteristic C101 will be described. When the pulsation of the turbine flow rate Qt is small and the peak value is the flow rate q1, the pulsation peak value of the turbine driving force Ft becomes the driving force Ft1 (point P11). On the other hand, when the pulsation of the turbine flow rate Qt is large and the peak value is the flow rate q2, the pulsation peak value of the turbine driving force Ft becomes the driving force Ft2 '(point P12'). From this, it appears that the average value of the turbine driving force Ft is increased when the pulsation of the turbine flow rate Qt is larger than when the pulsation of the turbine flow rate Qt is smaller. However, in practice, the blow-down period is shortened so as to offset it, so that the time-averaged turbine driving force is theoretically the same (there is no pulsation, and the same is true in a steady flow).

  On the other hand, the actual characteristic C11 is as follows. When the pulsation of the turbine flow rate Qt is small, it is the same as the characteristic C101 (point P11). However, when the pulsation of the turbine flow rate Qt is large and the peak value is the flow rate q2, the pulsation peak value of the turbine driving force Ft becomes the driving force Ft2 (point P12). As shown in FIG. 10, since the driving force Ft2> the driving force Ft2 ′, in this case, even when the pulsation of the turbine flow rate Qt is larger, the time average is reduced even if the decrease due to the shortening of the blowdown period is subtracted. This increases the turbine driving force.

  As described above, since the turbocharger 50 has a downwardly convex turbine driving force characteristic such as the characteristic C11, the time average value of the turbine driving force increases as the pulsation of the turbine flow rate Qt increases. The pressure can be increased. This is the principle of dynamic pressure supercharging.

  FIG. 11 is an exhaust pulsation characteristic diagram (actual measurement value). The abscissa indicates the crank angle θ (deg: top dead center is 0 ° CA) of the first cylinder 3a, and the ordinate indicates the exhaust flow rate Qe (kg / s). The illustrated characteristic is a characteristic when there is no throttling effect by the variable exhaust valve 30 (when the variable exhaust valve 30 is fully open). The exhaust flow rate Qe is the sum of all the cylinders 3. Accordingly, blowdown occurs at a 180 ° CA cycle (in any cylinder 3). In the illustrated example, blowdown occurs in the first cylinder 3a between 180 ° CA and 360 ° CA. When the wastegate valve 62 is closed, the exhaust flow rate Qe is equal to the turbine flow rate Qt (FIG. 10).

  A characteristic C12 is a characteristic of the present embodiment, and has a pulsation peak value of the flow rate q2 (corresponding to the flow rate q2 of FIG. 10), that is, a characteristic of large exhaust pulsation. On the other hand, the characteristic C102 is a characteristic shown for comparison with the characteristic C12, and has a pulsation peak value of the flow rate q1 (corresponding to the flow rate q1 in FIG. 10), that is, a characteristic with small exhaust pulsation. The characteristic C12 has a larger exhaust pulsation than the characteristic C102, and the blow-down period is shortened accordingly. That is, the characteristic C12 has a higher dynamic pressure supercharging effect than the characteristic C102. Specifically, the turbine speed (measured value) of the characteristic C12 increased by 43% compared to the characteristic C102.

  Also, by performing strong dynamic pressure supercharging, the blow-down period is shortened, so the exhaust pressure after blow-down decreases, the exhaust resistance decreases, the residual gas decreases, and the intake charge amount and knock resistance are reduced. There is also an effect of improvement.

  The most effective means for obtaining a large exhaust pulsation such as the characteristic C12 is to reduce the volume of the exhaust manifold 16. For this purpose, the first passage volume Va (≈fourth passage volume Vd≈auxiliary collective passage volume Vbc) shown in FIG. 4 may be reduced. In view of the fact that reducing the cross-sectional area of the passage increases the exhaust resistance, which is undesirable, the first exhaust passage 16a can be made as short as possible to reduce the first passage volume Va. It will be. Specifically, it is desirable that the length La (shown in FIG. 4) of the first exhaust passage 16a be 6 times or less the passage diameter D1 (shown in FIG. 6) at the exhaust manifold inlet of the first exhaust passage 16a. In the present embodiment, since the diameter D1 = φ36 mm and the length La ≦ 200 mm as described above, this condition is satisfied. Therefore, effective dynamic pressure supercharging can be expected.

  Further, as described above, the passage volumes of the first passage volume Va, the fourth passage volume Vd, and the auxiliary collective passage volume Vbc of the exhaust manifold 16 are substantially equal to each other. If there is a large difference between the volumes of these independent exhaust passages, the scavenging promotion effect due to the ejector effect also varies greatly between the cylinders. As a result, there is a difference in the knocking performance depending on the scavenging performance. As a result, the setting corresponding to the cylinder 3 having the lowest knocking performance is forced, and even if the knocking performance is improved in the other cylinders 3, it is wasted. . In addition, the cylinder-to-cylinder variation also occurs in the intake amount increasing effect due to the ejector effect.

  According to the configuration of the present embodiment, the volumes of the first passage volume Va, the fourth passage volume Vd, and the auxiliary collective passage volume Vbc are substantially equal to each other. Therefore, there is no such problem, and the advantage of the ejector effect can be obtained more effectively. be able to.

  By the way, in a general in-line four-cylinder engine, if the length La of the first exhaust passage 16a and the length Ld of the fourth exhaust passage 16d are naturally laid out, the collecting portion 31c is closer to the center. The layout is almost symmetrical as in the present embodiment. Then, if the second exhaust passage 16b and the third exhaust passage 16c are independent of each other, it is natural that the length thereof becomes shorter than the length La and the length Ld. In order to forcibly align this with the length La, a layout such as unnaturally detouring is required. This is not desirable because it causes an increase in exhaust resistance or prevents the length La and the length Ld from being shortened in order to establish the layout.

  According to the present embodiment, since the second exhaust passage 16b and the third exhaust passage 16c, which tend to be small in volume, are gathered to form the auxiliary collective exhaust passage 16bc, the volume of the auxiliary collective exhaust passage 16bc can be easily increased. The first exhaust passage volume Va and the fourth exhaust passage volume Vd can be made substantially equal.

  Note that the second exhaust passage 16b and the third exhaust passage 16c are independent of each other even if they are assembled. As shown in Table 1, since the second cylinder 3b and the third cylinder 3c are not adjacent to each other, the exhaust valve 9 starts to open before the bottom dead center and closes after the top dead center. However, there is no period in which the exhaust valve 9 of the second cylinder 3b and the exhaust valve 9 of the third cylinder 3c are both open. Therefore, there is no mutual exhaust interference, and in the exhaust stroke of the second cylinder 3b, the third exhaust passage 16c can be regarded as an extension of the second exhaust passage 16b, and in the exhaust stroke of the third cylinder 3c. The second exhaust passage 16b can be regarded as an extension of the third exhaust passage 16c in a pseudo manner.

  Thus, in this embodiment, although it is a 4-cylinder engine, the mutually independent relationship is implement | achieved by the three independent exhaust passages. By doing so, the layout can be made compact, and the connection portion between the housing 31 and the turbocharger 50 can be downsized.

  By the way, if the exhaust manifold volume is reduced, the dynamic pressure supercharging effect is increased as described above, but on the other hand, the exhaust temperature tends to increase in the high rotation region. Therefore, for example, it is desirable to improve heat resistance by using cast steel having high heat resistance as the material of the exhaust manifold 16 or by cooling the exhaust manifold 16 with water.

(2) Independent exhaust throttle control using each independent exhaust passage and variable exhaust valve Next, independent exhaust throttle control using each independent exhaust passage 16a, 16bc, 16d and variable exhaust valve 30 will be described.

  This will be specifically described with reference to FIG. As described above, in the state of FIG. 2, the first cylinder 3a is in the blow-down state and the second cylinder 3b is in the overlap period. Exhaust gas We (blow-down gas) guided to the first exhaust passage 16 a is throttled by the variable exhaust valve 30. The throttled blowdown gas increases in flow rate and decreases in pressure. This throttled blowdown gas corresponds to the driving fluid that provides the ejector effect.

  On the other hand, in the collecting portion 31c, the first exhaust passage 16a through which the blowdown gas flows and the auxiliary collecting exhaust passage 16bc communicate with each other. Accordingly, the exhaust We (sucked fluid) flowing through the auxiliary collective exhaust passage 16bc (and the second exhaust passage 16b) is sucked into the blow-down gas (driving fluid) having a low pressure and introduced into the collecting portion 31c (ejector effect). ).

  Even after the exhaust valve 9 of the second cylinder 3b is closed (after the overlap period), if the ejector effect of the driving fluid continues, the second exhaust passage 16b and the auxiliary collective exhaust passage 16bc. The exhaust gas We remaining in the gas can be sucked out, and scavenging can be promoted.

  Although FIG. 2 shows a case where the first cylinder 3a is in a blow-down state, the same applies to other cases. For example, when the second cylinder 3b is in the blow-down state, as is apparent from Table 1, the fourth cylinder 3d is in the overlap state. Therefore, the exhaust gas We flowing through the second exhaust passage 16b (auxiliary collective exhaust passage 16bc) becomes the driving fluid, and the exhaust gas We flowing through the fourth exhaust passage 16d becomes the sucked fluid. For example, when the third cylinder 3c is in a blow-down state, the first cylinder 3a is in an overlap state. Therefore, the exhaust gas We flowing through the third exhaust passage 16c (auxiliary collective exhaust passage 16bc) is the driving fluid, and the exhaust gas We flowing through the first exhaust passage 16a is the sucked fluid. For example, when the fourth cylinder 3d is in the blow-down state, the third cylinder 3c is in the overlap state. Accordingly, the exhaust We flowing through the fourth exhaust passage 16d is the driving fluid, and the exhaust We flowing through the third exhaust passage 16c (auxiliary collective exhaust passage 16bc) is the sucked fluid.

  In this way, in the cylinders adjacent in the ignition order, there is a relationship that the exhaust gas We of the cylinder after the ignition order becomes the driving fluid and the exhaust We of the previous cylinder becomes the sucked fluid. On the other hand, in order to properly obtain the ejector effect, the exhaust passage for the driving fluid and the exhaust passage for the sucked fluid need to be independent upstream of the variable exhaust valve 30. In other words, the exhaust passages need to be independent from each other in the cylinders adjacent in the ignition order. In the case of the present embodiment, the first exhaust passage 16a and the fourth exhaust passage 16d are clearly independent of any other exhaust passage, so the above condition is satisfied. The second exhaust passage 16b and the third exhaust passage 16c are gathered upstream of the variable exhaust valve 30 to form an auxiliary collective exhaust passage 16bc. However, as described above, since the second cylinder 3b and the third cylinder 3c are not cylinders whose ignition order is adjacent, there is no problem in satisfying the above condition even if they are not independent. Eventually, in the present embodiment, the exhaust passages of the cylinders adjacent to each other in the ignition order are independent from each other, so that an appropriate ejector effect can be obtained.

  In order to further enhance the ejector effect, the exhaust We corresponding to the driving fluid and the exhaust We corresponding to the fluid to be sucked may be merged at a shallowest angle (an angle close to parallel) as much as possible. In the present embodiment, the three independent exhaust passages 16a, 16bc, 16d are arranged in parallel in the vicinity of the manifold outlet 17, and the parallel arrangement is maintained until reaching the collecting portion 31c after flowing into the housing 31. The condition of the merging angle is satisfied. That is, a high ejector effect can be obtained.

  The advantages of the ejector effect mainly include the following three points.

  The first is an increase in the turbine flow rate of the turbocharger 50 (the amount of exhaust gas We supplied to the turbocharger 50). The turbine flow rate at the time of blowdown is obtained by adding the amount of exhaust gas We sucked out by the ejector effect to the normal blowdown gas amount. That is, the turbine flow rate is increased accordingly. As a result, the turbine driving force increases and the supercharging pressure can be improved.

  Secondly, the scavenging of the exhaust gas We is promoted. Due to the ejector effect, the exhaust We as the fluid to be sucked out is sucked and scavenging is promoted, so that the exhaust resistance of the cylinder 3 is reduced. Further, since the intake of air in the overlap period is promoted by the promotion of scavenging, the intake air amount can be increased and the engine torque can be increased.

  Third, promotion of dynamic pressure supercharging. As described above, the effect of dynamic pressure supercharging can be obtained by reducing the volume of the exhaust manifold 16, but the effect can be further promoted as described below by the ejector effect.

  If the variable exhaust valve 30 is not present or is fully open and the ejector effect cannot be expected, the blowdown gas flows into the other exhaust passage (reverses flow) through the collecting portion 31c. This acts as if the volume of the exhaust passage is apparently increased. On the other hand, if there is an ejector effect by the variable exhaust valve 30, the blow-down gas sucks the exhaust We as the driven fluid from the other exhaust passage as the driving fluid. That is, it does not wrap around other exhaust passages. This brings about the effect of reducing the exhaust passage volume in the dynamic pressure supercharging.

  As described above, when the entire exhaust passage volume (exhaust manifold volume) is the same, the present embodiment having the ejector effect by the variable exhaust valve 30 promotes the dynamic pressure supercharging more than the one without the ejector effect. It can be done.

  As described above, the advantages of the ejector effect have been described. The ejector effect becomes more prominent as the exhaust gas We corresponding to the driving fluid is strongly reduced. The degree of throttling can be changed by adjusting the opening degree of the variable exhaust valve 30, that is, by swinging the flap 35 around the flap shaft 37 (indicated by an arrow Z1 in FIG. 2).

  FIG. 6 shows the flow of the exhaust We when the variable exhaust valve 30 is fully open in the variable exhaust valve 30 and a specific shape in the vicinity thereof. FIG. 6 shows a state in which the exhaust We from the first cylinder 3a flows into the housing 31 through the first exhaust passage 16a, but the exhaust We from the other cylinders is the same.

  When the variable exhaust valve 30 is fully open, the flap 35 is almost completely stored in the bulging portion 31 b, and a part of the flap 35 forms a wall surface of the exhaust passage that continues from the exhaust manifold 16.

  Accordingly, the exhaust We from the first exhaust passage 16 a smoothly flows into the housing 31 through the manifold outlet 17. Since the partition plate 32 is provided in the upstream part of the housing 31, the independent exhaust state is maintained until the rear edge 32a. Then, the trailing edge 32a is guided to the collecting portion 31c without hitting the fan-shaped surface 36 of the flap 35, that is, without being squeezed. The exhaust gas We is further guided from the collecting portion 31c to the housing 51 of the turbocharger 50.

  In the present embodiment, the housing 31 is bent downward for the convenience of the layout of the turbocharger 50. The exhaust gas We is smoothly guided to the turbocharger 50 by the flow guide plate 33 provided along the curved flow path of the housing 31.

  Further, the flow of the exhaust Wet becomes smoother by the two rectifying guides 34. As shown in FIG. 4, at the manifold outlet 17, a first exhaust passage 16a and a fourth exhaust passage 16d are arranged in parallel on both sides of the auxiliary collective exhaust passage 16bc. Even in the housing 31, the positional relationship is maintained until reaching the collecting portion 31 c. Therefore, the exhaust We from the first exhaust passage 16a and the fourth exhaust passage 16d flows in with an offset in plan view with respect to the axis of the collective portion 31c. Generates a flow (vortex). In addition, the turning direction is opposite between the exhaust from the first exhaust passage 16a and the exhaust from the fourth exhaust passage 16d, so that the flow of the exhaust We is greatly disturbed in the collecting portion 31c. Such turbulence of the exhaust may reduce the ejector effect.

  In order to solve this problem, the flow straightening guide 34 of the present embodiment is provided on the extended surface of the partition plate 32. Therefore, the flow from the first exhaust passage 16a toward the auxiliary collective exhaust passage 16bc and the fourth exhaust passage 16d. The flow toward the auxiliary collective exhaust passage 16bc is restricted. Accordingly, the above-mentioned turning is suppressed, and the flow of the exhaust gas We in the collecting portion 31c becomes smoother. Therefore, the ejector effect can be further enhanced.

  On the other hand, the flow of the exhaust We when the variable exhaust valve 30 is fully closed is shown in FIG. When the variable exhaust valve 30 is in the fully closed state, the flap 35 swings into the housing 31 and the fan-shaped surface 36 blocks the flow of the exhaust We. However, it is not completely shielded, and the flow path in the bent inner portion of the flow guide plate 33 is secured.

  Exhaust gas We from the first exhaust passage 16 a flows into the housing 31 through the manifold outlet 17. Since the partition plate 32 is provided in the upstream part of the housing 31, the independent exhaust state is maintained until the rear edge 32a. Then, the exhausted air We blocked by the fan-shaped surface 36 of the flap 35 at the rear edge 32a flows into the collecting portion 31c at the rear edge 32a of the partition plate 32 (the bent inner portion inlet portion of the flow guide plate 33). At that time, the exhausted We that has been throttled to high speed and low pressure acts as a driving fluid, and sucks the exhaust We from other exhaust passages (mainly the auxiliary collective exhaust passage 16bc) by the ejector effect. These exhausts We join together and are smoothly guided to the housing 51 of the turbocharger 50 by the flow guide plate 33.

  The variable exhaust valve 30 can take an intermediate opening between fully closed and fully open. In that case, the closer to full closure, the stronger the throttle action of the blow-down exhaust gas We and the higher the ejector effect.

  Next, independent exhaust throttle control will be described. As described above, the independent exhaust throttle control is a control in which the ECU 2 (variable exhaust valve control means) causes the variable exhaust valve 30 to reduce the cross-sectional areas of the independent exhaust passages 16a, 16bc, and 16d as compared with the maximum area. Specifically, the ECU 20 sends an opening signal to the actuator 38 of the variable exhaust valve 30, and the actuator 38 rotationally drives the flap shaft 37 to adjust the rotation angle of the flap 35. In the present embodiment, the independent exhaust throttle control is performed in the low rotation region A3 (see FIG. 18) of the supercharging region. Specifically, the low rotation supercharging region A3 is set to a low rotation region, for example, 2000 prm or less from the intercept point at which the wastegate valve 62 starts to open. Since the wastegate valve 62 opens in order to suppress the supercharging pressure from becoming too high in the high rotation range from the intercept point, it is not necessary to increase the supercharging pressure due to the ejector effect in the high rotation range. Therefore, the variable exhaust valve 30 is fully opened to suppress the exhaust resistance.

  In the independent exhaust throttle control, the variable exhaust valve 30 is set to a lower opening degree as the engine speed Ne is lower in the low rotation supercharging region A3.

  As mentioned above, although the dynamic pressure supercharging and independent exhaust throttle control which are the main technical features of this embodiment were demonstrated, these are closely related and cooperate and are improving supercharging performance.

  FIG. 12 is a graph showing the charging efficiency ηc in the low rotation supercharging region A3. The horizontal axis represents the engine speed Ne (rpm), and the vertical axis represents the charging efficiency ηc (%). A characteristic C13 is a characteristic of the present embodiment in which dynamic pressure supercharging and independent exhaust throttle control are used in combination. A characteristic C103 is a characteristic shown for comparison, and is a characteristic when a conventional general exhaust manifold (without the variable exhaust valve 30) is used. The filling efficiency ηc of the characteristic C13 is increased by about 20 to 30 points with respect to the characteristic C103. This is the effect of increasing the supercharging pressure by dynamic pressure supercharging and independent exhaust throttle control using the variable exhaust valve 30.

  FIG. 13 is a graph showing the average effective pressure BMEP of the engine in the low rotation supercharging region A3. The horizontal axis represents the engine speed Ne (rpm), and the vertical axis represents the average effective pressure BMEP (kPa). A characteristic C14 is a characteristic of the present embodiment in which dynamic pressure supercharging and independent exhaust throttle control are used together (a characteristic corresponding to the characteristic C13 in FIG. 12). A characteristic C104 is a characteristic shown for comparison, and corresponds to the characteristic C103 in FIG. The average effective pressure BMEP of the characteristic C14 is increased by about 200 to 400 kPa with respect to the characteristic C104. This is an effect that the charging efficiency is increased (FIG. 12) by the dynamic pressure supercharging and the independent exhaust throttle control using the variable exhaust valve 30, that is, the engine torque is increased.

  Next, a description will be given of a further technique employed in the present embodiment in order to achieve the ejector effect more remarkably.

FIG. 14 is a graph showing the relationship between the exhaust passage restriction degree and the volumetric efficiency ηv in the present embodiment. The upper part of the horizontal axis indicates the aperture diameter D2 (mm). The throttle diameter D2 is a diameter of a circle corresponding to the flow path cross-sectional area S2 at the rear edge 32a of the partition plate 32 when the variable exhaust valve 30 shown in FIG. The inlet portion of the exhaust manifold 16 of the first exhaust passage 16a has a circular shape of φD1 (: original diameter = 36 mm), and the passage cross-sectional area S1 of the portion is an area corresponding to φD1 (about 1018 mm 2 ).

The lower part of the horizontal axis indicates the sectional area drawing ratio Rd (%). This cross-sectional area drawing ratio Rd is an area ratio of the drawing diameter D2 to the original diameter D1. That is, Rd = (D2 / D1) 2 × 100 (%), or Rd = (S2 / S1) × 100 (%).

  A characteristic C15 shown in FIG. 14 is a characteristic at an engine speed Ne = 1500 rpm, and C16 is a characteristic at 2000 rpm. As is clear from these characteristics, there is a particularly preferable range in which the volumetric efficiency ηv is particularly high in the range of the aperture diameter D2 = 22 to 28 mm (the cross-sectional area aperture ratio Rd = 37 to 61%). This indicates that a particularly remarkable ejector effect can be obtained in this preferred range. Therefore, by setting the aperture diameter D2 within this preferable range, a higher supercharging effect can be obtained and the engine torque can be further increased.

  Next, valve timing change control by the variable valve timing mechanism 12 will be described.

  FIG. 15 is an explanatory diagram of the valve timing change control. The abscissa represents the crank angle θ (deg: ° CA), and the top dead center TDC of the first cylinder 3a is set to 0 ° CA. The vertical axis shows a schematic valve opening amount of the intake and exhaust valves 7 and 9. The upper stage shows the cylinder that ignites later among the cylinders that are adjacent in the ignition order, and the lower stage shows the cylinder that ignites first. As an example, the first cylinder 3a is shown in the upper stage, and the second cylinder 3b is shown in the lower stage. The first cylinder 3a is in a transition period from the expansion stroke to the exhaust stroke (near the bottom dead center), and the second cylinder 3b is in a transition period from the exhaust stroke to the intake stroke (near the top dead center). ing. This corresponds to the state shown in FIG.

  The exhaust valve open period Pe1 and the intake valve open period Pi1 indicated by the solid lines are characteristics when the independent exhaust throttle control is not performed and the variable exhaust valve 30 is fully open (in the present embodiment, for example, a natural intake region). Here, an overlap L2 in which the exhaust valve opening period Pe1 and the intake valve opening period Pi1 overlap is set near the top dead center of the second cylinder 3b.

  Generally, the overlap is provided in order to sufficiently scavenge the exhaust gas We and to suck more intake air Wi. There is also an aim of pushing out the exhaust We with the intake air Wi. Similar to general variable valve timing control, the overlap L2 is changed so as to increase as the engine speed Ne increases. Specifically, the overlap L2 is expanded by delaying the closing timing of the exhaust valve 9 by the exhaust VVT 12e and advancing the opening timing of the intake valve 7 by the intake VVT 12i (performed by either the exhaust VVT 12e or the intake VVT 12i). May be).

  On the other hand, the exhaust valve opening period Pe2 and the intake valve opening period Pi2 indicated by broken lines are characteristics when the independent exhaust throttle control is being executed, that is, when the variable exhaust valve 30 throttles the exhaust We in the low-rotation supercharging region A3. The overlap L3 in this case is enlarged as shown in the figure than the overlap L2 when the independent exhaust throttle control is not performed even with the same load and the same engine speed Ne. Specifically, the closing timing of the exhaust valve 9 is delayed, and the opening timing of the intake valve 7 is advanced.

  Originally, according to the independent exhaust throttling control, scavenging is promoted by the ejector effect as described above, and intake during the overlap period is promoted, so that the intake amount is increased and the engine torque is increased. Therefore, as shown in FIG. 15, the above effect can be obtained more significantly by expanding the overlap L <b> 2 to the overlap L <b> 3 by the variable valve timing mechanism 12.

  Normally, if the overlap L2 is expanded carelessly, there is a possibility that the exhaust gas We will flow backward due to the intake negative pressure. However, in the independent exhaust throttle control, the exhaust We is sucked downstream by the ejector effect, and thus such a backflow hardly occurs. In other words, the overlap amount can be increased while suppressing the adverse effect of the backflow of the exhaust We.

  By the way, as shown in the exhaust valve opening period Pe1 in the upper stage of FIG. 15 (first cylinder 3a), the exhaust valve 9 when the independent exhaust throttle control is not executed is relatively early, for example, bottom dead, before the bottom dead center of the exhaust stroke. Start opening at 40-60 ° CA before the spot. By doing so, scavenging is promoted, but on the other hand, since the exhaust gas We starts while the piston descends, the momentum of the blow-down gas is reduced accordingly. This is disadvantageous for the independent exhaust throttle control of the present embodiment that uses blowdown gas as the drive fluid for the ejector effect.

  However, in this embodiment, when the independent exhaust throttle control is executed, the overlap amount is increased by delaying the exhaust valve closing timing. This also means that the exhaust valve opening timing is delayed at the same time (because the valve opening period itself is changed in translation and is not changed). That is, as shown in the upper part of FIG. 15, the exhaust valve opening timing is delayed by the period L1. Thereby, the fall of the momentum of the said blowdown gas is suppressed. After the bottom dead center, the action of pushing up the exhaust gas Wep is added, so that the blowdown gas can be energized. Thus, the ejector effect can be obtained more remarkably.

However, if the exhaust valve 9 is opened after the bottom dead center of the exhaust, there is an adverse effect that the exhaust resistance increases. Thus the delay of the exhaust valve opening timing is such that keep kept to just before the exhaust bottom dead center as shown.

  Next, the effect of the turbocharger 50 of this embodiment is demonstrated. FIG. 16 is a graph showing the efficiency of the turbocharger 50. The horizontal axis represents the exhaust flow rate Qe (kg / s), and the vertical axis represents the efficiency ηt (%). A characteristic C21 indicated by a solid line is a characteristic of the turbocharger 50 of the present embodiment, and a characteristic C121 indicated by a broken line is a characteristic of a general large turbocharger. In the general characteristic C121, the efficiency ηt is set to have a peak near the center of the application range of the exhaust flow rate Qe.

  On the other hand, the turbocharger 50 of the present embodiment is a large turbocharger as described above, but has a relatively small A / R as compared to a general large turbocharger. In addition, the peak of the efficiency ηt of the characteristic C21 is shifted to the high flow rate side as compared with the characteristic C121. Moreover, the peak value also becomes high. Since the turbine flow rate Qt increases as the exhaust flow rate Qe increases, such a setting is suitable for the dynamic pressure supercharging of the present embodiment that actively uses the region where the turbine flow rate Qt is large.

  FIG. 17 is a graph showing engine torque characteristics. The horizontal axis indicates the engine speed Ne (rpm), and the vertical axis indicates the engine torque Te (N · m). A characteristic C24 indicated by a solid line is a characteristic of the present embodiment, a characteristic C124 indicated by a broken line is a characteristic when a conventional exhaust system and a general large turbocharger are used, and a characteristic C125 is a characteristic of a conventional exhaust system. This is a characteristic when a typical small turbocharger (a turbocharger having a relatively small turbine diameter D and A / R) is employed.

  As shown in the figure, in the characteristic C124, the torque increase action in the high rotation range by the large turbocharger is strong, and in the characteristic C125, the torque increase action in the low rotation range by the small turbocharger is strong.

  On the other hand, the characteristic C24 of the present embodiment has a strong torque increasing effect by adopting the large turbocharger 50 in the high rotation region, and the independent exhaust using the dynamic pressure supercharging and the variable exhaust valve 30 in the low rotation region. The torque increasing action is strong by the throttle control, the valve timing change control, the adoption of the small A / R turbocharger 50, and the like. As a result, although it is a simple configuration using one turbocharger 50, a large supercharging effect is obtained in a wide range from the low rotation region to the high rotation region, and an increase in engine torque is achieved.

  Next, a second embodiment of the present invention will be described. 2nd Embodiment is the same as that of 1st Embodiment about the point which performs the apparatus structure shown in FIGS. 1-9, dynamic pressure supercharging and independent exhaust throttle control, valve timing change control, etc. . The difference from the first embodiment is that the first embodiment performs the independent exhaust control only in the low-rotation supercharging region A3, but in the present embodiment, it is also executed in the natural intake region, and the overlap amount is also achieved. It is the point which performs control which increases. This utilizes the effect of independent exhaust throttling control in which the back flow of exhaust is suppressed by the ejector effect even when the overlap amount is increased. As a result, it is possible to improve the overlap expansion response when the acceleration is requested.

  FIG. 18 is an engine torque characteristic diagram for explaining the characteristics of the second embodiment. The horizontal axis indicates the engine speed Ne (rpm), and the vertical axis indicates the engine torque Te (N · m). The characteristic Tx indicates the maximum load torque, and the characteristic RL indicates the torque corresponding to the running resistance (so-called road load). Symbol A1 indicates a natural intake region (hereinafter also referred to as NA region), symbol A2 indicates a supercharging region, and characteristic B indicates a boundary between the NA region A1 and the supercharging region A2. The low rotation supercharging area A3 is included in the supercharging area A2.

  In the present embodiment, independent exhaust throttle control is executed also in the NA region A1, and the exhaust We is throttled by the variable exhaust valve 30. At the same time, the variable valve timing mechanism 12 enlarges the overlap L2 (FIG. 15) more than usual (when the independent exhaust throttle control is not executed).

  As described above, if the overlap L2 is inadvertently enlarged, the exhaust gas may flow backward, but in this embodiment, the exhaust gas backflow is suppressed by the ejector effect by the independent exhaust throttle control. The overlap L2 can be enlarged without any problem.

  For example, in FIG. 18, a case is assumed in which a steady operation state P21 corresponding to load / load in the NA region A1 rapidly shifts to an operation state P22 in the low rotation supercharging region A3 due to an acceleration request. In normal cases, in the operating state P21, the overlap L2 is set to 0 to 15 ° CA, for example. On the other hand, in the operating state P22, the overlap L3 is enlarged as shown in FIG. 15 as in the first embodiment, and is set to, for example, around 65 ° CA. Therefore, at the time of transition from the operating state P21 to P22, the variable valve timing mechanism 12 needs to change the overlap amount greatly, and response delay may be a problem.

  On the other hand, in the present embodiment, the overlap L2 in the operating state P21 is enlarged more than usual, and is set to 10 to 40 ° CA, for example. As a result, the overlap expansion allowance at the time of transition to the operating state P22 is reduced, so that the time required for the change can be shortened. That is, responsiveness can be improved.

  Note that, in the high rotation region where the engine speed Ne> 2000 rpm, the independent exhaust throttle control is not performed in the supercharging region A2, and the variable exhaust valve 30 is fully opened, as in the first embodiment. On the other hand, the valve timing change control is executed, and the amount of overlap increases as the rotation speed increases and increases. Therefore, in the present embodiment, the overlap L2 is expanded as compared with the normal case while performing the independent exhaust throttle control also in the NA region A1 of such a high rotation region. Thereby, the responsiveness of the overlap change in the high rotation region is improved. It is appropriate to increase the overlap amount within a range not exceeding the overlap amount in the high load region after the transition.

  As mentioned above, although embodiment of this invention was described, each said embodiment can be suitably changed in the range which does not deviate from the summary of this invention. For example, the engine 1 does not necessarily have to be an in-line 4-cylinder engine, and may be an in-line 6-cylinder engine or a V-type 6-cylinder engine. And the shape of the exhaust manifold according to it should just be taken.

  Further, the exhaust manifold does not necessarily have an auxiliary collective exhaust passage, and may be completely independent until each exhaust passage reaches the variable exhaust valve.

  The variable exhaust valve may be any valve that can change the cross-sectional area of each independent exhaust passage, and is not necessarily limited to the structure of the variable exhaust valve 30. For example, a valve that can independently change the cross-sectional areas of the independent exhaust passages may be used. The flap 35 does not necessarily have a fan shape, and may be a plate-shaped flap plate, for example.

1 is a schematic configuration diagram of an engine supercharging device according to a first embodiment of the present invention. It is a partial sectional side view of FIG. It is an external appearance perspective view of the housing of an exhaust manifold and a variable exhaust valve. It is a downstream partial perspective view of an exhaust manifold. It is a principal part perspective view of a variable exhaust valve. It is a longitudinal cross-sectional view of the housing of an exhaust manifold and a variable exhaust valve, and is a view showing a state where the variable exhaust valve is in an open state. FIG. 7 is a cross-sectional view similar to FIG. 6, showing a state where the variable exhaust valve is in an open state. It is the VIII-VIII sectional view taken on the line of FIG. It is explanatory drawing of the fluid force which acts on the flap of a variable exhaust valve, (a) shows the case where a fan-shaped flap is used, (b) shows the case where a general flap board is used. It is a turbine characteristic view of a turbocharger. It is an exhaust pulsation characteristic figure. It is a graph which shows the filling efficiency in a low rotation supercharging area | region. It is a graph which shows the average effective pressure of the engine in a low rotation supercharging area | region. It is a graph which shows the relationship between the throttle degree of an exhaust passage, and volumetric efficiency. It is explanatory drawing of valve timing change control. It is a graph which shows the efficiency of a turbocharger. It is a graph which shows an engine torque characteristic. It is an engine torque characteristic figure, Comprising: It is explanatory drawing of 2nd Embodiment.

1 Engine 3 Cylinder 7 Intake Valve 8 Exhaust Port 9 Exhaust Valve 12 Variable Valve Timing Mechanism (Valve Timing Changing Means)
12e Exhaust VVT (Exhaust valve timing changing means)
16 Exhaust manifolds 16a, 16d First and fourth exhaust passages (independent exhaust passages)
16bc Auxiliary exhaust passage (independent exhaust passage)
20 ECU (variable exhaust valve control means)
30 Variable exhaust valve 31c Collecting part 50 Exhaust turbocharger 62 Wastegate valve A1 Natural intake area A2 Supercharge area A3 Low rotation supercharge area (predetermined low rotation area)
L2, L3 Overlap amount

Claims (3)

  1. An exhaust manifold connected to the exhaust port of each cylinder and having a plurality of independent exhaust passages;
    A collecting portion where the independent exhaust passages are gathered on the exhaust manifold or downstream thereof;
    An exhaust turbocharger connected to the downstream side of the collecting section;
    A variable exhaust valve capable of changing a cross-sectional area of each of the independent exhaust passages upstream of the collecting portion;
    Variable exhaust valve control means for driving and controlling the variable exhaust valve ;
    An exhaust valve timing changing means capable of changing an opening / closing valve timing of an exhaust valve for opening / closing the exhaust port of each cylinder ,
    The variable exhaust valve control means is configured to reduce the passage cross-sectional area of the independent exhaust passage by the variable exhaust valve at a maximum area at least in an engine speed lower than an engine speed at which the wastegate valve of the turbocharger starts to open. And the independent exhaust throttle control is executed to increase the degree of reduction of each passage cross-sectional area as the engine speed is lower .
    The exhaust valve timing changing means delays the opening timing of the exhaust valve within a range before exhaust bottom dead center when the independent exhaust throttle control is executed, when the independent exhaust throttle control is not executed. apparatus.
  2. A valve timing changing means capable of changing the amount of overlap in which both the intake valve and the exhaust valve are open by changing the opening / closing valve timing of at least one of the intake valve and the exhaust valve;
    2. The supercharging device for an engine according to claim 1, wherein the valve timing changing means increases the overlap amount when the independent exhaust throttle control is executed and when the independent exhaust throttle control is not executed .
  3. The variable exhaust valve control means performs the independent exhaust throttle control also in the natural intake region,
    3. The supercharging device for an engine according to claim 2, wherein the valve timing changing means increases the overlap amount when the independent exhaust throttle control is executed in the natural intake region when the independent exhaust throttle control is executed .
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