JP2009114991A - Supercharging device for engine - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To provide a supercharging device for an engine capable of increasing engine torque in a low rotation region and effectively suppressing surging with simple structure. <P>SOLUTION: The supercharging device includes: an exhaust manifold having independent exhaust passages 16a, 16bc, 16d; a collecting part 31c downstream of the exhaust manifold 16 where the independent exhaust passages 16a-16d are collected; an exhaust turbocharger 50 connected downstream of the collected part 31c; a variable exhaust valve 30 capable of changing each passage cross sectional area of the independent exhaust passages 16a-16d; a variable exhaust valve control means 20; an intake relief passage 22 connecting an exhaust passage 11 to an exhaust passage 60; an intake relief valve 23 provided in the intake relief passage 22; and an intake relief control means 20. The variable exhaust valve control means 20 executes independent exhaust throttling control of the independent exhaust passages 16a-16d in a low rotation supercharging region A3, and the intake relief control means 20 controls to open the intake relief valve 23 in a surging region G1. <P>COPYRIGHT: (C)2009,JPO&INPIT

Description

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

  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.

  On the other hand, the occurrence of surging may be a problem for a turbocharger compressor. Generally, surging is likely to occur when the flow rate of air flowing through the compressor is small and the pressure ratio (ratio of pressure before and after the compressor) is high, in other words, in a low rotation / high load region. The operating region where surging occurs is called a surging region. Surging occurs due to air separation (vortex) around the impeller of the compressor. If this occurs, supercharging efficiency is reduced, and vibration may occur or the compressor material may be fatigued.

In recent years, therefore, a method for suppressing surging by returning a part of the pressurized intake air to, for example, the intake side in the surging region has been studied. Specifically, a bypass passage that connects the intake passages before and after the compressor and a bypass valve that opens and closes the bypass passage are provided. Then, the bypass valve is opened in the surging region. In this way, the amount of air flowing through the compressor can be relatively increased with respect to the substantial amount of intake air (intake amount introduced into the combustion chamber), so that the transition from the surging region to the non-surging region can be performed. it can. That is, the occurrence of surging can be suppressed.
JP-A-9-112285

  However, regarding the problem of the supercharging region, each of the conventional superchargers including the variable geometry turbo shown in Patent Document 1 has a problem that the structure is complicated or the size is increased. The solution is demanded.

  Further, with respect to surging, the above-described conventional suppression measures allow a part of the air pressurized by the compressor to escape to the upstream low-pressure part, which is wasteful and correspondingly deteriorates fuel consumption.

  The present invention has been made in view of the above circumstances, and it is possible to increase the engine torque in the low rotation range and to suppress the surging efficiently while having a simple structure. It is an object to provide a feeding device.

  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. An exhaust turbocharger including a collecting portion, a turbine connected to the downstream side of the collecting portion and driven by exhaust; a compressor for boosting intake air by the driving force of the turbine; and an upstream side of the collecting portion A variable exhaust valve capable of changing the cross-sectional area of each of the independent exhaust passages, variable exhaust valve control means for driving and controlling the variable exhaust valves, the compressor downstream side of the intake passage, and the variable exhaust of the exhaust passage. An intake relief passage connecting the valve downstream side and the turbine upstream side, an intake relief valve disposed on the intake relief passage, and the intake air Intake relief control means for controlling the opening of the leaf valve, and the variable exhaust valve control means adjusts the cross-sectional areas of the independent exhaust passages by the variable exhaust valves in a predetermined low rotation region of the supercharging region. Independent exhaust throttle control for reducing the maximum area is performed, and the intake relief control means controls opening of the intake relief valve when the compressor is in a predetermined surging region in a supercharging region. To do.

  The invention according to claim 2 is the engine supercharging device according to claim 1, wherein an EGR passage connecting the exhaust passage and the intake passage, an EGR valve disposed in the EGR passage, and a predetermined low pressure EGR control means for controlling the opening of the EGR valve in a load region, and the intake relief passage and the intake relief valve are configured to be shared by the EGR passage and the EGR valve.

  The invention according to claim 3 is the engine supercharging device according to claim 2, wherein the EGR passage is provided with a water-cooled EGR cooler capable of exchanging heat between the engine coolant and the gas in the EGR passage. It is characterized by.

  The invention according to claim 4 is the engine supercharging device according to any one of claims 1 to 3, wherein when the intake relief valve is controlled to open, the variable exhaust valve control means or the intake relief. The control means adjusts the opening of at least one of the intake relief valve and the variable exhaust valve so that the compressor is in the vicinity of the boundary between the surging region and the other non-surging region and in the non-surging region. Features.

  According to the first aspect of the present invention, as will be described below, the engine torque in the low rotation region can be increased while the structure is simple, and efficient suppression of surging can be achieved.

  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 following four points can be cited mainly as advantages of the ejector effect.

  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.

  Fourth, promotion of intake air relief according to the present invention. Here, intake relief will be described first. In the configuration of the present invention, an intake relief passage connecting the compressor downstream side of the intake passage, the variable exhaust valve downstream side of the exhaust passage and the turbine upstream side, and an intake relief valve disposed on the intake relief passage are provided. Is provided. The intake relief control means controls the opening of the intake relief valve when the compressor is in the surging region in the supercharging region. This is the intake relief. That is, the intake air relief according to the present invention is a kind of one that suppresses surging by guiding (relieving) a part of the compressed air by the compressor to the exhaust passage side.

  According to the intake relief of the present invention, the amount of air flowing through the compressor can be relatively increased with respect to the substantial intake amount (the intake amount introduced into the combustion chamber), so that the surging region is changed to the non-surging region. Can be migrated. That is, the occurrence of surging can be suppressed.

  Moreover, the pressurized air (hereinafter referred to as relief gas) guided to the exhaust side through the intake relief passage can be effectively utilized as energy for driving the turbine of the turbocharger. Compared to the conventional type that recirculates pressurized air to the upstream side of the compressor, it is superior in that the relief gas performs effective work, and it can suppress the deterioration of fuel consumption due to relief of pressurized air. it can.

  However, it is not always possible to perform proper intake relief simply by providing an intake relief passage and opening the intake relief valve. Since the exhaust passage exhaust pressure, especially the exhaust pressure upstream of the turbine, is high, relief gas is difficult to flow from the intake side to the exhaust side even if the intake relief valve is opened, and conversely there is a concern of backflow from the exhaust side to the intake side .

  As a solution to this, it is effective to use the ejector effect by the variable exhaust valve described above (fourth advantage). In the present invention, the exhaust side of the intake relief passage is connected to the downstream side of the variable exhaust valve. Therefore, the ejector effect by the independent exhaust throttle control can be obtained at the connection portion between the intake relief passage and the exhaust passage. According to the ejector effect, the exhaust side of the intake relief passage has a low pressure, so that the relief gas can be smoothly guided from the intake passage to the exhaust passage by opening the intake relief valve. That is, intake relief can be promoted.

  Moreover, there is a large overlap between the surging area (low rotation / high load area (supercharging area)) and the independent exhaust throttle control execution area (low rotation / supercharging area) (usually the surging area is independent exhaust throttle control). Therefore, it is easy to properly achieve both independent exhaust throttle control and intake relief.

  As described above, according to the present invention, it is possible to increase the engine torque in the low rotation range while having a simple structure using one turbocharger (first to third ejector effects). ) And efficient suppression of surging can be achieved (combination of intake relief control and the fourth advantage of the ejector effect).

  According to the second aspect of the present invention, as will be described below, the intake air relief is performed without using a separate dedicated intake relief passage and intake relief valve by using the system while performing EGR (exhaust gas recirculation). It can be performed.

  Conventionally, EGR for recirculating exhaust gas to the intake side has been widely performed. When EGR is performed, the ratio of the inert gas component (recirculated exhaust gas, that is, GER gas) to the intake air increases, so that the combustion temperature can be lowered and the generation and emission of nitrogen oxides (NOx) are suppressed. be able to. Moreover, since the amount of intake gas can be increased while suppressing an increase in the amount of oxygen, the intake negative pressure can be reduced and the pumping loss can be reduced.

  According to the configuration of the present invention, an EGR passage connecting the exhaust passage and the intake passage, an EGR valve disposed in the EGR passage, and an EGR control means for controlling the opening of the EGR valve in a predetermined low load region, By providing the above, proper EGR can be performed.

  Here, the intake relief and the EGR include a passage (an intake relief passage or an EGR passage) that bypasses the engine body and connects the intake passage and the exhaust passage, and a valve that opens and closes the passage (intake relief). There are many parts that are common in terms of structure, such as a valve or an EGR valve. Therefore, in the present invention, the dedicated intake relief passage and the intake relief valve are not provided, and these are shared by the EGR passage and the EGR valve, whereby the structure can be simplified and the cost can be reduced.

  In order to properly use the EGR passage, the compressor downstream side of the intake passage and the variable exhaust valve downstream side of the exhaust passage and the upstream side of the turbine are connected so that the EGR passage also functions as an intake relief passage. Although it is necessary, there is no particular problem in the function of EGR for doing so.

  Further, the intake relief is performed in the surging region, that is, the low rotation / high load region, and the EGR is performed in the natural intake region, that is, the low load region. That is, normally, the intake air relief and the EGR are performed at the same time or there is no request for them, so that both can be performed properly without interfering with each other.

  According to the invention of claim 3, as will be described below, the water-cooled EGR cooler (hereinafter abbreviated as EGR cooler) can be used effectively not only when the EGR is performed but also when the intake relief is performed.

  In general, since the EGR gas that recirculates in the EGR passage is high in temperature, it is not preferable if it is led directly to the intake passage because the intake air temperature rises. Therefore, an increase in the intake air temperature can be suppressed by providing an EGR cooler to cool the EGR gas.

  On the other hand, when the EGR passage is used as the intake relief passage during intake relief, the EGR cooler conversely has an action of heating the relief gas. The EGR cooler is a water-cooled type using engine cooling water, and the cooling water temperature is about 90 ° C. when warm. In contrast, since the temperature of the relief gas is around 50 ° C., heat is supplied from the cooling water side to the relief gas side by heat exchange in the EGR cooler. The heated relief gas increases in temperature and the internal energy increases. Therefore, the exhaust gas energy after the relief gas merges with the normal exhaust gas also increases, and the turbine drive energy of the turbocharger increases. Thus, the relief gas can be used as turbine drive energy more efficiently.

  According to the invention of claim 4, as described below, the amount of relief gas can be suppressed as much as possible while suppressing surging.

  As described above, according to the intake air relief of the present invention, it is possible to suppress surging while suppressing deterioration of fuel consumption, but it is desirable to keep the amount of relief gas to a minimum. For this purpose, as in the present invention, intake relief control may be performed so that the compressor is in the vicinity of the boundary between the surging region and the non-surging region and in the non-surging region.

  Note that the greater the opening of the intake relief valve and the greater the degree of throttle of the variable exhaust valve, the greater the amount of relief gas. Therefore, as in the present invention, the amount of relief gas can be adjusted to an optimum value by adjusting the opening of at least one of the intake relief valve and the variable exhaust valve.

  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 an 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 and that surging can be more efficiently suppressed.

As means for achieving this, it has the following three technical features.
(1) Improvement of supercharging capability by dynamic pressure supercharging (2) Independent exhaust throttle control using independent exhaust passage and variable exhaust valve 30 (3) Surge suppression by intake relief control These will be described in detail later. First, the configuration and structure of this embodiment will be described.

  The engine 1 is an in-line four-cylinder four-cycle engine provided with an exhaust turbine supercharger 50 (hereinafter abbreviated as a turbocharger). The turbocharger 50 is a well-known supercharger, in which a compressor 52 provided in the intake passage 10 and a turbine 54 provided in the exhaust passage 60 are connected by a shaft 53. In FIG. 1, the compressor 52 and the turbine 54 are shown separately for the sake of clarity, but in practice, the compressor 52 is provided on one end side of one shaft 53 and the turbine 54 is provided on the other end side. Is provided. In the vicinity of the installation position of the turbocharger 50, the intake passage 10 and the exhaust passage 60 are actually close to each other, and the turbocharger 50 is interposed therebetween. The turbocharger 50 drives the compressor 52 by rotating the turbine 54 with the exhaust gas We, and compresses the intake air Wi to increase the intake pressure.

  An air flow meter 11 is provided upstream of the compressor 52 of the turbocharger 50 to detect the intake air amount.

  On the other hand, an intercooler 13 that cools intake air that has been compressed by the compressor 52 and has risen in temperature is provided on the downstream side of the compressor 52. A throttle valve 18 that throttles intake air in accordance with the operating state is provided on the downstream side. The opening degree of the throttle valve 18 is adjusted by an actuator 19 composed of a motor or the like. Further, a surge tank 14 is provided on the downstream side. The surge tank 14 is an air reservoir that temporarily retains intake air, and suppresses intake pulsation and interference that adversely affect the accuracy of the air flow meter 11, and increases the density and increases the flow velocity by retaining air. To increase inhalation efficiency.

  A pressure sensor 25 is attached to the surge tank 14 to detect intake pressure in the surge tank 14 (hereinafter referred to as surge tank pressure). The surge tank pressure is also used in normal engine control. In this embodiment, the surge tank pressure is also used to detect surging of the compressor 52 as described later.

  An intake manifold 15 that guides intake air to each cylinder is provided on the downstream side of the surge tank 14, and its downstream end is connected to the engine main body 2 having a cylinder block and a cylinder head as main members. The engine body 2 is provided with first to fourth cylinders 3a, 3b, 3c, 3d (referred to collectively as cylinders 3) on a straight line. The configuration of each cylinder 3 is common, and as shown in FIG. 2, an intake port 6 for inhaling intake Wi connected to an intake manifold 15 at an upper portion of the combustion chamber 4 and an exhaust port for discharging exhaust We 8 are provided. 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.

  The first to fourth exhaust passages 16a, 16b, 16c, and 16d are individually connected to the exhaust ports 8 of the first to fourth cylinders 3a, 3b, 3c, and 3d. 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 described above, the turbine 54 of the turbocharger 50 is provided on the downstream side of the housing 31 (collecting portion 31c). 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. With this setting, the turbine flow rate at which the efficiency of the turbocharger 50 is maximized is shifted to a higher flow rate side than a general large turbo. Moreover, the maximum value of the efficiency also becomes high. Such a setting is suitable for the dynamic pressure supercharging (details will be described later) of the present embodiment that actively uses a region having a large turbine flow rate.

  As shown in FIG. 1, the exhaust passage 60 is provided with a waist passage 61 that bypasses the turbine 54 and a wastegate valve 62 that opens and closes the waist passage 61.

  Further, an EGR passage 22 that connects the exhaust passage 60 and the intake passage 10 is provided as a structure for executing EGR. The upstream end of the EGR passage 22 opens downstream from the variable exhaust valve 30 and the collecting portion 31 c and upstream from the turbine 54. On the other hand, the downstream end opens between the throttle valve 18 of the intake passage 10 and the surge tank 14. The EGR passage 22 has an EGR valve 23 that opens and closes it, and a water-cooled EGR cooler 24 (hereinafter abbreviated as EGR cooler) that can exchange heat between the engine coolant and the gas in the EGR passage 22 (EGR gas in the case of EGR). ) Is provided.

  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 in the intake valve 7 and the exhaust valve 9 Change the valve timing.

  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 transmits a throttle opening signal to the actuator 19 according to the operating state, and controls the opening of the throttle valve 18. Also, an appropriate fuel supply amount and ignition timing are calculated in response to an intake air amount signal from the air flow meter 11 and a surge tank pressure signal from the pressure sensor 25, and a drive signal is transmitted to a fuel injection valve and an ignition plug 5 (not shown). . The ECU 20 performs drive control of the variable valve timing mechanism 12 and opening / closing control of the waste gate valve 62.

  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 maximizes the cross-sectional areas of the independent exhaust passages 16a, 16bc, and 16d by the variable exhaust valve 30 in a predetermined low-rotation region (low-rotation supercharging region A3 shown in FIG. 9) of the supercharging region. Independent exhaust throttling control is performed to reduce the area (the variable exhaust valve 30 is open).

  Furthermore, the ECU 20 also functions as an EGR control unit that controls the opening of the EGR valve 23. Specifically, the ECU 20 opens the EGR valve 23 as appropriate in a predetermined low load region (natural intake region A1 shown in FIG. 9), and returns an appropriate amount of EGR gas Ge to the intake side.

  When the EGR is performed, the ratio of the inert gas component (EGR gas Ge) to the intake air increases, so that the combustion temperature can be lowered and the generation / exhaust of NOx can be suppressed. Moreover, since the amount of intake gas can be increased while suppressing an increase in the amount of oxygen, the intake negative pressure can be reduced and the pumping loss can be reduced.

  In general, since the EGR gas Ge is at a high temperature, if it is introduced into the intake passage 10 as it is, the intake air temperature rises, which is not preferable. The EGR cooler 24 performs heat exchange between the engine coolant and the EGR gas Ge, and cools the EGR gas Ge, thereby suppressing an increase in intake air temperature.

  The operation region where EGR has a remarkable effect is a low load region where the required intake air amount (oxygen amount) is small. In the present embodiment, the EGR execution area is the NA area A1, but this area coincides with the low load area where the EGR has a remarkable effect. On the other hand, EGR is not performed in the supercharging region A2 shown in FIG. This is because if EGR is performed in this region, the recirculated EGR gas hinders the introduction of fresh air and may cause a negative effect of reducing the supercharging effect.

  Further, the ECU 20 also functions as intake relief control means for controlling intake relief for suppressing the surging of the compressor 52. Specifically, the ECU 20 performs the surging determination of the compressor 52 based on the surge tank pressure signal from the pressure sensor 25. If it is determined that surging has occurred (is in the surging area), intake relief control is executed. That is, the EGR valve 23 is controlled to open. This intake relief control will be described in detail later.

  By the way, in the engine 1 of this embodiment, as in a general four-cylinder engine, the respective cylinders 3 are shifted from each other so that the respective ignition timings sequentially reach every 90 degrees of crank angle (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 the cylinder head of the engine body 2. 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 the area of a circle 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. Each rear edge 32a of each partition plate 32 is partially arc-shaped along the fan-shaped surface 36 of the flap 35 when the variable exhaust valve 30 is in the closed state.

  Therefore, 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, and the exhaust gas We is in an independent and parallel flow state. Squeezed while maintaining.

  Further, on the downstream side of the rear edge 32a of the partition plate 32, a collective portion 31c where independent exhausts We join each other is formed.

  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).

  The general configuration of the present embodiment has been described above. Next, the operation region of the engine 1 will be described. FIG. 9 is a diagram showing an operation region of the engine 1. 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. 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 region A3 is a low-rotation region in the supercharging region A2, and is a region that rotates at a speed lower than 2000 rpm in the present embodiment. As described above, EGR is performed in the NA area A1 in this embodiment. Independent exhaust throttling control described later is performed in the low rotation supercharging region A3.

  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 is large. 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 than the smaller one, the time average is subtracted from the decrease due to the shortening of the blowdown period. 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 described later will also vary 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. . Further, the cylinder-to-cylinder variation also occurs in the intake amount increasing effect by the ejector effect described later.

  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. When the independent exhaust throttle control is executed, the exhaust We (blowdown 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.

  When the variable exhaust valve 30 is not provided or is fully opened and the ejector effect cannot be expected, the blowdown gas flows into the other exhaust passage (backflows) 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.

  Fourthly, the above-described promotion of intake relief will be described in detail in the section “(3) Surge suppression by intake relief control” described later.

  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 supercharging region A3 shown in FIG. The low rotation supercharging area A3 is set to a low rotation area from the intercept point at which the wastegate valve 62 starts to open, in this embodiment, 2000 pm or less. 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.

  In the present embodiment, the independent exhaust throttle control is limited to the low rotation supercharging region A3 and is not performed in the NA region A1. Even if the independent exhaust throttle control is performed in the NA region A1, there is no serious problem in itself, and there are several advantages. However, when EGR is performed in the NA region A1 as in this embodiment, it is better not to perform independent exhaust throttle control in the NA region A1, as described below.

  As shown in FIG. 1, the upstream end of the EGR passage 22 opens to the downstream side of the variable exhaust valve 30. On the other hand, the ejector effect of the independent exhaust throttle control occurs on the downstream side of the variable exhaust valve 30. Therefore, if the independent exhaust throttle control is executed during the EGR, the ejector effect by the independent exhaust throttle control may reach the upstream end of the EGR passage 22 as well. When the ejector effect reaches the upstream end of the EGR passage 22, the suction action prevents the EGR gas Ge from flowing from the exhaust side toward the intake side, or reverses the flow, so that proper EGR is not performed.

  Therefore, in the present embodiment, the independent exhaust throttle control is prohibited in the NA region A1 where EGR is performed. By doing so, the interference problem with said EGR can be solved exactly. On the other hand, most of the profits obtained by the independent exhaust throttle control are obtained in the low-rotation supercharging region A3. Therefore, even if the independent exhaust throttle control is not performed in the NA region A1, the opportunity loss is small.

  As described above, (1) dynamic pressure supercharging and (2) independent exhaust throttle control, which are the main technical features of the present embodiment, have been described, but these are closely related and cooperate to improve engine performance. Yes.

  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 opening period Pe1 and the intake valve opening period Pi1 indicated by solid lines are characteristics when the variable exhaust valve 30 is fully open (for example, the NA region A1) without performing independent exhaust throttle control. 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 (first cylinder 3a) in FIG. 15, 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. Therefore, it is desirable to keep the delay of the exhaust valve opening timing just before exhaust bottom dead center as shown in the figure.

  FIG. 16 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 (corresponding to the characteristic Tx in FIG. 9), a characteristic C124 indicated by a broken line is a characteristic when a conventional exhaust system and a general large turbocharger are employed, and A characteristic C125 is a characteristic when a conventional exhaust system and a general small turbocharger (a turbocharger having a relatively small turbine diameter D and A / R) are 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 due to the throttle control, valve timing change control, and the adoption of the small A / R turbocharger 50. 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.

  Although not directly shown in FIG. 16, appropriate EGR is performed in the NA region A1, which is a partial load region, and exhaust gas purification by NOx reduction and fuel efficiency improvement by pumping loss reduction are achieved.

  Next, (3) suppression of surging by intake relief control, which is a main technical feature of the present embodiment, will be described.

  FIG. 17 is a characteristic diagram illustrating suppression of surging by intake relief control. The horizontal axis represents the compressor flow rate Qc (the amount of air flowing through the compressor 52), and the vertical axis represents the pressure ratio (downstream pressure / upstream pressure of the compressor 52).

  The surging of the compressor 52 is a phenomenon that occurs when the pressure ratio is greater than or equal to a predetermined value at a certain compressor flow rate Qc (when in the surging region G1). It is likely to occur in a load region (for example, particularly in a low rotation / high load region in the low rotation supercharging region A3 shown in FIG. 9). Surging occurs due to separation (vortex) of air around the impeller of the compressor 52, and if this occurs, the supercharging efficiency is lowered and vibration is generated and the material of the compressor 52 is fatigued.

  Therefore, in the present embodiment, intake relief is performed in order to suppress the surging of the compressor 52. In the intake relief, a part of the intake air pressurized by the compressor 52 (hereinafter referred to as relief gas Gi) is led to the exhaust passage 60 side in the surging region G1. In the intake relief control, the ECU 20 (intake relief control means) controls the opening of the EGR valve 23 when the compressor 52 is in the surging region G1. When the EGR valve 23 is opened, the relief gas Gi is guided from the intake passage 10 to the exhaust passage 60 via the EGR passage 22 (see FIG. 1).

  In order to perform intake relief originally, an intake relief passage connecting the downstream side of the compressor 52 in the intake passage 10 to the downstream side of the variable exhaust valve 30 in the exhaust passage 60 and the upstream side of the turbine 54, and the intake air thereof An intake relief valve disposed on the relief passage is required. However, such an intake relief passage and an intake relief valve are similar in structure to the EGR passage 22 and the EGR valve 23. Therefore, in the present embodiment, the intake relief passage and the intake relief valve are shared by the EGR passage 22 and the EGR valve 23. That is, when performing intake relief, the EGR passage 22 is used as an intake relief passage, and the EGR valve 23 is used as an intake relief valve. By doing so, there is no need to provide a separate dedicated intake relief passage or intake relief valve, and the structure can be simplified and the cost can be reduced.

  In addition, intake relief is performed in the surging region G1 (low rotation / high load region), and EGR is performed in the natural intake region A1 (low load region). That is, normally, the intake air relief and the EGR are performed at the same time or there is no request for them, so that both can be performed properly without interfering with each other.

  Returning to FIG. 17, the description will be continued. A region where surging does not occur other than the surging region G1 is indicated by a non-surging region G2. Surging line C31 is a boundary line between surging region G1 and non-surging region G2.

  A characteristic C32 indicated by a broken line is an operation line indicating a transient characteristic of the compressor 52 when the engine 1 shifts from a low rotation / low load operation state (for example, an idle operation state) to a high rotation / high load operation state in a short time. The operating line (virtual operating line) is assumed when intake relief control is not executed even if surging occurs. On the other hand, a characteristic C33 indicated by a solid line is an operating line (actual operating line) when the intake relief control is executed in the surging region G1.

  If the compressor 52 is in the surging region G1 (for example, the state P5) on the virtual operation line C32, surging occurs. At that time, the ECU 20 immediately executes the intake relief control, that is, opens the EGR valve 23. Then, the relief gas Gi is guided from the intake side to the exhaust side via the EGR passage 22, and the compressor flow rate Qc increases accordingly. That is, since the state P5 changes to the state P6, the compressor 52 shifts from the surging region G1 to the non-surging region G2.

  As a result of continuously executing such intake relief control along the virtual operation line C32, an actual operation line C33 is actually obtained, and surging is appropriately suppressed.

  Next, the above effect will be described in comparison with the conventional structure. State P7 shows a conventional structure, that is, a case where a part of the pressurized air by the compressor 52 is returned to the upstream side of the compressor 52. In this case, the state changes from P5 to P7. As described above, even in the conventional structure, the compressor 52 can be shifted from the surging region G1 to the non-surging region G2 by increasing the compressor flow rate Qc. However, the pressure ratio is lowered by refluxing the pressurized air. That is, the supercharging pressure is lowered and the supercharging efficiency is lowered. On the other hand, in the case of the present embodiment, the relief gas Gi is used as turbine drive energy, so that the work amount of the compressor 52 increases and the pressure ratio can be increased compared to the state P7 (state P6). That is, a decrease in supercharging efficiency can be suppressed.

  In addition, the state P6 shown in FIG. 17 has shown the case where a pressure ratio becomes equal to the state P5. This indicates that the decrease in the pressure ratio due to relief of the pressurized air and the increase in the pressure ratio due to the increase in the work amount of the compressor 52 are offset. Actually, the pressure ratio does not always change in this way, and the pressure ratio in the state P6 can fluctuate up and down within a range higher than the pressure ratio in the state P7.

  By the way, in this intake relief control, just opening the EGR valve 23 cannot always perform proper intake relief. Since the exhaust pressure in the exhaust passage 60, particularly the exhaust pressure upstream of the turbine 54, is high, the relief gas Gi hardly flows from the intake passage 10 to the exhaust passage 60 even when the EGR valve 23 is opened. There is even a concern of backflow.

  Therefore, in the present embodiment, an appropriate intake relief is realized by the ejector effect by the variable exhaust valve 30 described above (the promotion of intake relief, which is the fourth advantage by the ejector effect). In the present embodiment, the exhaust side of the EGR passage 22 is connected to the downstream side of the variable exhaust valve 30. Therefore, the ejector effect by the independent exhaust throttle control can be obtained at the connection portion between the EGR passage 22 and the exhaust passage 60. According to the ejector effect, the exhaust side of the EGR passage 22 has a low pressure, so that the relief gas Gi can be smoothly guided from the intake passage 10 to the exhaust passage 60 by opening the EGR valve 23. That is, intake relief can be promoted.

  In addition, the surging region G1 is a particularly low rotation / high load region in the low rotation supercharging region A3 in which the independent exhaust throttle control is normally performed. That is, when the intake relief control is executed, the independent exhaust throttle control is also executed at the same time or before. Therefore, the ECU 20 can properly balance the intake relief control and the independent exhaust throttle control by controlling the opening degree of the variable exhaust valve 30.

  Further, the relief gas Gi guided to the exhaust side through the EGR passage 22 can be effectively used as energy for driving the turbine 54. Compared with the conventional type that recirculates pressurized air to the upstream side of the compressor 52, the relief gas Gi is superior in that it performs an effective work, and suppresses deterioration of fuel consumption due to relief of the pressurized air. be able to.

  However, it is desirable to keep the amount of relief gas to a minimum. For this purpose, the actual operating line C33 may be made as close as possible to the surging line C31 in the non-surging region G2. The amount of relief gas increases as the opening degree of the EGR valve 23 increases and as the throttle degree of the variable exhaust valve 30 increases. Therefore, the relief gas amount can be adjusted to the optimum value by adjusting the opening of at least one of the EGR valve 23 and the variable exhaust valve 30. In the case of this embodiment, the throttle degree of the variable exhaust valve 30 is determined from the requirements of the independent exhaust throttle control, and the opening degree of the EGR valve 23 is set so that the relief gas amount becomes the optimum value (minimum) accordingly. (Both may be adjusted separately while balancing both).

  In the present embodiment, the opening degree of the EGR valve 23 (or the throttle degree of the variable exhaust valve 30) is mapped and stored in advance from the characteristic diagram of FIG. And when setting the opening degree of the actual EGR valve 23, etc., it correct | amends according to the generation | occurrence | production state of surging. By doing so, the actual operating line C33 can be brought closer to the surging line C31 by correction while the rough value such as the opening degree of the EGR valve 23 is quickly set to improve control responsiveness.

  Furthermore, the EGR cooler 24 cools the EGR gas Ge that recirculates in the EGR passage 22 when GER is executed, but conversely heats the relief gas Gi flowing in the EGR passage 22 during intake relief. The EGR cooler 24 is a water-cooled type using engine cooling water, and the cooling water temperature is about 90 ° C. when warm. In contrast, since the temperature of the relief gas Gi is around 50 ° C., heat is supplied from the cooling water side to the relief gas Gi side by heat exchange in the EGR cooler 24. The heated relief gas Gi rises in temperature and the internal energy increases. Therefore, the exhaust gas energy after the relief gas Gi merges with the normal exhaust gas We also increases, and the turbine drive energy of the turbocharger 50 increases. Thus, the relief gas Gi can be used more efficiently as turbine drive energy.

  In the present embodiment, the ECU 20 detects the occurrence of surging as follows. Surging is a self-excited vibration of the air around the impeller blades of the compressor 52, and is always generated at a substantially constant frequency independent of the rotation speed of the compressor 52 and the engine rotation speed. Therefore, the surge tank pressure pulsation when surging occurs is characteristic, and can be clearly distinguished from, for example, a surge tank pressure change or pulsation that accompanies a normal accelerator opening / closing operation by the driver. Therefore, the ECU 20 receives the surge tank pressure signal detected by the pressure sensor 25 and detects the pulsation frequency using a band pass filter or the like as appropriate. When the pulsation frequency is a frequency that is peculiar to the occurrence of surging (preliminarily examined by experiment or the like and stored in the ECU 20), it is determined that surging has occurred (is in the surging area).

  FIG. 18 is a flowchart of control selection according to the operation region by the ECU 20. In this control, first, signals from various sensors such as the air flow meter 11, the pressure sensor 25, and a rotational speed sensor (not shown) are read by the ECU 20 (step S1). Next, a determination is made as to which of the operation regions shown in FIG. 9 is the current operation state (steps S2 and S3). If it is determined in step S2 that the area is the NA area A1, that is, the EGR area (YES in step S2), the ECU 20 executes EGR. That is, the EGR valve 23 is opened, and the opening degree control according to the operating state is performed (step S15). In this case, since the independent exhaust throttle control is not performed, the ECU 20 fully opens the variable exhaust valve 30 (step S16).

  On the other hand, if NO in step S2, that is, if it is determined that the region is the supercharging region A2 and not the EGR region, it is further determined whether or not it is the low rotation supercharging region A3 (step S3). If NO in step S3, that is, if it is determined that the engine is in the supercharging region A2 but not the low rotation supercharging region A3, the ECU 20 does not perform EGR. That is, the EGR valve 23 is fully closed (step S11). Since the independent exhaust throttle control is not performed, the process proceeds to step S16, and the variable exhaust valve 30 is fully opened.

  If YES in step S3, that is, if it is determined that the engine is in the low rotation supercharging region A3, the ECU 20 executes independent exhaust throttle control. That is, the variable exhaust valve 30 is throttled according to the operating state (step S4). EGR is not performed.

  Further, when occurrence of surging is recognized (YES in step S5), the ECU 20 executes intake relief control. That is, the EGR valve 23 is controlled to open (step S6), and surging is suppressed. If surging does not occur (NO in step S5), the EGR non-execution state is continued and the process returns.

  In FIG. 18, for the sake of convenience, it is assumed that surging has occurred in step S5, but actually, the ECU 20 always determines whether surging has occurred. However, since surging usually occurs in the low rotation supercharging region A3 particularly in the case of low rotation and high load, the result is almost the same as the case where the surging determination is performed at the timing of step S5. Become.

  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, in the present embodiment, a dedicated intake relief passage or intake relief valve for performing intake relief is not provided, but the EGR passage 22 and the EGR valve 23 are also used. However, it is not always necessary to do so, and a dedicated intake relief passage or intake relief valve may be provided separately. In addition, the engine is not necessarily limited to an engine that performs EGR, and an intake relief passage or an intake relief valve may be provided for an engine that does not perform EGR to perform intake relief control.

  In the above embodiment, as a means for detecting surging, the pressure sensor 25 attached to the surge tank 14 is used and the pulsation of the surge tank pressure is used. In this case, the pressure sensor 25 may be provided at a separate position (for example, between the throttle valve 18 and the surge tank 14), and the surging determination may be performed based on the pulsation of the intake negative pressure.

  Further, it is not always necessary to detect surging by pressure pulsation. For example, surging may be detected based on the pulsation of the intake flow rate detected by the air flow meter 11.

  Further, it does not necessarily have to be determined that the surging region is present when surging is actually detected. For example, an operation region where surging occurs or is likely to occur is stored in advance as a surging region, and the ECU 20 causes the intake relief to enter the surging region (regardless of actual occurrence). You may make it perform control. Or you may use together such a method and the method of performing intake-relief control by actually detecting surging like this embodiment.

1 is a schematic configuration diagram of an engine supercharging device according to an 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 a figure which shows the driving | operation area | region of an engine. 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 an engine torque characteristic. It is a characteristic view explaining suppression of surging by intake relief control. 4 is a flowchart of control selection according to an operation region by an ECU (variable exhaust valve control means / intake relief control means / EGR control means).

Explanation of symbols

1 Engine 3 Cylinder 8 Exhaust Port 16 Exhaust Manifold 16a First and Fourth Exhaust Paths (Independent Exhaust Path)
16bc Auxiliary exhaust passage (independent exhaust passage)
20 ECU (variable exhaust valve control means, intake relief control means, EGR control means)
22 EGR passage (also used as intake relief passage)
23 EGR valve (also used as intake relief valve)
24 Water-cooled EGR cooler 30 Variable exhaust valve 31c Collecting part 50 Exhaust turbocharger 52 Compressor 54 Turbine A1 Natural intake area (EGR area)
A2 Supercharging area A3 Low rotation supercharging area (predetermined low rotation area)
C31 Surging line (between surging area and non-surging area)
G1 Surging area G2 Non-surging area Ge EGR gas (gas in the EGR passage)
Gi relief gas (gas in the EGR passage)

Claims (4)

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 comprising a turbine connected to the downstream side of the collecting portion and driven by exhaust, and a compressor for boosting the intake air by the driving force of the turbine;
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 intake relief passage connecting the compressor downstream side of the intake passage and the variable exhaust valve downstream side of the exhaust passage and the turbine upstream side;
An intake relief valve disposed on the intake relief passage;
Intake relief control means for controlling the opening of the intake relief valve,
The variable exhaust valve control means performs independent exhaust throttle control for reducing the cross-sectional area of the independent exhaust passage by the variable exhaust valve from a maximum area in a predetermined low rotation region of the supercharging region,
The supercharging device for an engine, wherein the intake relief control means controls the opening of the intake relief valve when the compressor is in a predetermined surging region in the supercharging region. An EGR passage connecting the exhaust passage and the intake passage;
An EGR valve disposed in the EGR passage;
EGR control means for controlling the opening of the EGR valve in a predetermined low load region,
2. The supercharging device for an engine according to claim 1, wherein the intake relief passage and the intake relief valve are configured to be shared by the EGR passage and the EGR valve.   3. The engine supercharging device according to claim 2, wherein a water-cooled EGR cooler capable of exchanging heat between engine cooling water and gas in the EGR passage is disposed in the EGR passage.   When the intake relief valve is controlled to open, the variable exhaust valve control means or the intake relief control means is such that the compressor is near the boundary between the surging area and the other non-surging area and in the non-surging area. 4. The engine supercharging device according to claim 1, wherein an opening degree of at least one of the intake relief valve and the variable exhaust valve is adjusted.

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