WO2009040642A1

WO2009040642A1 - Exhaust device and control device for internal combustion engine

Info

Publication number: WO2009040642A1
Application number: PCT/IB2008/002496
Authority: WO
Inventors: Yasuyuki Irisawa; Shigeki Miyashita
Original assignee: Toyota Jidosha Kabushiki Kaisha
Priority date: 2007-09-27
Filing date: 2008-09-25
Publication date: 2009-04-02
Also published as: WO2009040642A9

Abstract

Each cylinder of an engine includes a first exhaust port passing through a turbocharger, and a second exhaust port passing downstream of the turbocharger. The first exhaust ports of first and third cylinders and the first exhaust ports of fifth and seventh cylinders, which are respectively adjacent to each other in a first bank, are converged initially within a cylinder head. The first exhaust ports of second and fourth cylinders and the first exhaust ports of sixth and eighth cylinders, which are respectively adjacent to each other in a second bank, are converged initially within the cylinder head.

Description

EXHAUST DEVICE AND CONTROL DEVICE FOR INTERNAL COMBUSTION

ENGINE

BACKGROUND OF THE INVENTION

1. Field of the Invention

[0001] The present invention relates to an exhaust device for an internal combustion engine that performs combustion at irregular intervals, and more particularly to an exhaust device for an internal combustion engine having a turbocharger. The present invention also relates to a control device for an internal combustion engine.

2. Description of the Related Art

[0002] In a conventional device (an independent exhaust engine), each cylinder is provided with a first exhaust valve for opening and closing a first exhaust passage that passes through a turbine and a second exhaust valve for opening and closing a second exhaust passage that does not pass through the turbine (see Japanese Patent Application Publication No. 10-89106 (JP-A-10-89106), for example).

[0003] Further, according to a conventional technique employed in an engine in which combustion is performed at irregular intervals, the surface area of an exhaust passage of a cylinder having an overlap period that does not overlap an exhaust stroke is made larger than the surface area of an exhaust passage of a cylinder having an overlap period that overlaps the exhaust stroke (see Japanese Patent Application Publication No. 2006-161581 (JP-A-2006-161581), for example).

[0004] Incidentally, turbochargers are designed to improve in efficiency when an exhaust pulse is in a certain condition. Therefore, when an exhaust pulse led to the turbocharger is reduced, supercharging responsiveness during a transition period may deteriorate. Furthermore, so-called exhaust interference, in which the exhaust action of a certain cylinder is impaired by the exhaust action of another cylinder, may occur. In this case, an amount of residual gas may increase such that a favorable combustion condition cannot be obtained, and as a result, the performance may deteriorate.

[0005] FIG. 18 is a schematic plan view showing a V-type eight cylinder engine according to the related art. As shown in the drawing, a V-type eight cylinder engine 90 includes a left-hand cylinder row (left bank) 92L and a right-hand cylinder row (right bank) 92R. The left-hand cylinder row 92L is constituted by first, third, fifth and seventh cylinders, while the right-hand cylinder row 92R is constituted by second, fourth, sixth and eighth cylinders. The cylinders of the left-hand cylinder row 92L share an exhaust manifold 94L, and the cylinders of the right-hand cylinder row 92R share an exhaust manifold 94R. [0006] Various ignition sequences may be employed in this type of V-type eight cylinder engine 90, but here, a case in which the ignition sequence is l → 8 → 7 → 3 → 6 → 5 → 4 → 2 will be described. FIG. 19 is a view showing relationships between a crank angle and the working strokes of each cylinder in the case of the above ignition sequence. As shown in the drawing, with this ignition sequence, the combustion intervals in the left-hand and right-hand cylinder rows are not regular intervals of 180°, but irregular intervals (90°, 180°, 270°).

[0007] The cylinders of the left-hand cylinder row 92L share the exhaust manifold 94L, and are therefore capable of receiving an effect from the exhaust pulse of another cylinder in the left-hand cylinder row 92L. Similarly, the cylinders of the right-hand cylinder row 92R share the exhaust manifold 94R, and are therefore capable of receiving an effect from the exhaust pulse of another cylinder in the right-hand cylinder row 92R. When the combustion intervals in the left-hand and right-hand cylinder rows are regular, the exhaust pulse effect received by each cylinder from another cylinder is uniform. In the V-type eight cylinder engine 90, however, the combustion intervals of the left-hand and right-hand cylinder rows are not regular, as described above, and therefore the exhaust pulse effect received from another cylinder varies from cylinder to cylinder.

[0008] Specifically, in the case of the left-hand cylinder row 92L, first, the first cylinder receives a blowdown effect from the seventh cylinder, making an increase in the amount of residual gas in the first cylinder more likely to occur. More specifically, as shown in FIG. 19, a valve overlap period in which an open period of an exhaust valve overlaps an open period of an intake valve occurs in the first cylinder when the exhaust valve of the seventh cylinder opens in order to discharge high-pressure exhaust gas from the cylinder into the exhaust manifold 94L, leading to an increase in the internal pressure of the exhaust manifold 94L. As a result, the exhaust gas in the exhaust manifold 94L flows back into an intake port of the first cylinder through the exhaust valve of the first cylinder, leading to an increase in the amount of residual gas in the first cylinder.

[0009] Further, in the left-hand cylinder row 92L, the third cylinder receives a blowdown effect from the fifth cylinder, making an increase in the amount of residual gas in the third cylinder more likely to occur, similarly to the case described above. On the other hand, the fifth and seventh cylinders are unlikely to receive an exhaust pulse effect from another cylinder, and hence the amount of residual gas therein decreases.

[0010] Likewise in the right-hand cylinder row 92R, the amount of residual gas in the sixth cylinder becomes likely to increase upon reception of a blowdown effect from the fourth cylinder, and the amount of residual gas in the second cylinder becomes likely to increase upon reception of a blowdown effect from the eighth cylinder, as is evident from FIG. 19. On the other hand, the fourth and eighth cylinders are unlikely to receive an exhaust pulse effect from another cylinder, and hence the amount of residual gas therein decreases. [0011] Hence, the V-type eight cylinder engine 90 possesses a characteristic whereby the amount of residual gas in a specific cylinder (the first, second, third and sixth cylinders) increases easily, regardless of the engine rotation speed and the engine load. As a result, the cylinders in which the amount of residual gas increases easily exhibit less favorable combustion and are more likely to misfire than the other cylinders. Accordingly, various adverse effects, such as deterioration of the drivability due to torque variation and the like, accelerated deterioration of an exhaust gas purification catalyst, and so on are more likely to occur.

[0012] In response to this problem, the related art proposes varying the shape of cams for driving the intake valve and the exhaust valve between the cylinders in which the amount of residual gas increases easily and the other cylinders in order to reduce the amount of residual gas in the cylinders in which the amount of residual gas increases easily. When this measure is taken, however, a valve characteristic correction value takes a fixed value, and therefore a sufficient reduction in the residual gas amount can only be achieved in a specific operating region. Hence, the problem is not solved in all operating regions.

[0013] Meanwhile, Japanese Patent Application Publication No. 2006-161619 (JP-A-2006-161619) discloses a technique in which a flow passage open/close valve is provided in the exhaust passages of blowdown gas generating cylinders (according to the above ignition sequence, the fifth, seventh, fourth and eighth cylinders) and the exhaust pressure of the blowdown gas generating cylinders is reduced by open/close controlling these flow passage open/close valves, thereby equalizing the residual gas amount in each cylinder. When this measure is taken, however, pump loss in the blowdown gas generating cylinders increases, and as a result, fuel efficiency deteriorates. Moreover, the manufacturing cost increases in proportion to the cost of the flow passage open/close valves.

[0014] Further, Japanese Patent Application Publication No. 2003-56374 (JP-A-2003-56374) discloses a technique employed in an engine in which valve characteristics of intake/exhaust valves can be controlled independently in each cylinder, wherein the valve characteristic is controlled such that either an intake valve opening timing of cylinders in which the residual gas amount increases easily is retarded in relation to the other cylinders or an exhaust valve closing timing thereof is advanced in relation to the other cylinders, thereby reducing the valve overlap period of the cylinders in which the residual gas amount increases easily in comparison with the other cylinders such that the amount of residual gas in these cylinders decreases. However, the intake valve opening timing and exhaust valve closing timing, as well as the valve overlap period, are extremely important parameters that greatly affect fuel efficiency, combustion control/emissions control corresponding to internal EGR amount control, and so on. Therefore, when the overlap period of only the cylinders in which the residual gas amount increases easily is simply shortened to avoid combustion deterioration and misfiring in these cylinders, the fuel efficiency performance and emissions performance are greatly affected, which is undesirable.

[0015] As described above, various techniques have been proposed in the related art to reduce the amount of residual gas in a specific cylinder of a V-type eight cylinder engine in which the residual gas amount increases easily, but all of these proposals have disadvantages.

[0016] Further, the following problem occurs in a V-type eight cylinder engine having a turbocharger. Typically, a turbine is provided in an exhaust passage of an engine having a turbocharger, and therefore back pressure is more likely to increase such that an increase in the residual gas amount is more likely to occur. Hence, in a specific cylinder in which the residual gas amount increases easily, the amount of residual gas increases even further, leading to further combustion deterioration and a further increase in the likelihood of a misfire.

SUMMARY OF THE INVENTION

[0017] The present invention provides an exhaust device for an internal combustion engine, which is capable of improving the supercharging responsiveness of a turbocharger by suppressing a reduction in an exhaust pulse. Further, the present invention provides an exhaust device for an internal combustion engine, which is capable of suppressing an increase in a residual gas amount in a cylinder by suppressing exhaust interference. The present invention also provides a control device for an internal combustion engine, which is capable of reliably suppressing an increase in a residual gas amount in a specific cylinder in which the residual gas amount is likely to be increased by an exhaust pulse effect from another cylinder, in an internal combustion engine that includes a turbocharger and has irregular combustion intervals between cylinders in an identical cylinder row.

[0018] A first aspect of the present invention relates to an exhaust device for an internal combustion engine, including: a turbocharger; and a plurality of cylinders, each including a first exhaust passage that passes through the turbocharger and a first exhaust valve provided in the first exhaust passage. In this exhaust device for an internal combustion engine, the first exhaust passages of a pair of cylinders, from among the plurality of cylinders, converge initially, and then converge with another first exhaust passage between a convergence point of the first exhaust passages and the turbocharger.

[0019] In the exhaust device for an internal combustion engine according to this aspect, the internal combustion engine may perform combustion at irregular intervals, each of the plurality of cylinders includes a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage, the internal combustion engine performs combustion at irregular intervals¹, and the first exhaust passages of adjacent cylinders, from among the plurality of cylinders, may converge initially, and then may converge with another first exhaust passage between a convergence point of the first exhaust passages and the turbocharger.

[0020] According to the constitution described above, the first exhaust passages of adjacent cylinders converge initially, and then converge with another first exhaust passage between the convergence point and the turbocharger. As a result, a reduction in an exhaust pulse that is led to the turbocharger can be suppressed. Accordingly, the transient responsiveness of the turbocharger can be improved.

[0021] In the exhaust device for an internal combustion engine according to this aspect, the internal combustion engine may be a V-type eight cylinder engine, and the first exhaust passage of a certain cylinder and the first exhaust passage of another cylinder having a combustion stroke that is removed from the combustion stroke of the certain cylinder by a crank angle of 180° do not have to converge initially upstream of the turbocharger. [0022] According to this constitution, the first exhaust passage of a certain cylinder of the V-type eight cylinder engine and the first exhaust passage of another cylinder that is removed from the combustion stroke of the certain cylinder by a crank angle of 180° do not converge initially upstream of the turbocharger. As a result, exhaust interference in which the exhaust action of the certain cylinder is impaired by the exhaust action of the other cylinder can be suppressed, and an increase in the amount of residual gas in the cylinder can be suppressed.

[0023] In the exhaust device for an internal combustion engine according to this aspect, the V-type eight cylinder engine may include, within a cylinder head, a first bank including a first cylinder, a third cylinder, a fifth cylinder and a seventh cylinder, and a second bank including a second cylinder, a fourth cylinder, a sixth cylinder and an eighth cylinder, and the first exhaust passages of the first cylinder and the third cylinder, the first exhaust passages of the fifth cylinder and the seventh cylinder, the first exhaust passages of the second cylinder and the fourth cylinder, and the first exhaust passages of the sixth cylinder and the eighth cylinder may respectively converge initially within the cylinder head.

[0024] According to this constitution, in the V-type eight cylinder engine, the first exhaust passages of the first cylinder and the third cylinder, the first exhaust passages of the fifth cylinder and the seventh cylinder, the first exhaust passages of the second cylinder and the fourth cylinder, and the first exhaust passages of the sixth cylinder and the eighth cylinder respectively converge initially within the cylinder head. As a result, exhaust interference in the first to eighth cylinders can be suppressed. Moreover, the layout of the exhaust passages can be made compact.

[0025] In the exhaust device for an internal combustion engine according to this aspect, the internal combustion engine may perform combustion at irregular intervals, each of the plurality of cylinders may include a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage, the plurality of cylinders may be constituted by a first cylinder, a second cylinder, a third cylinder, a fourth cylinder, a fifth cylinder, a sixth cylinder, a seventh cylinder, and an eighth cylinder, the internal combustion engine may be a V-type eight cylinder engine including, within a cylinder head, a first bank including the first cylinder, the third cylinder, the fifth cylinder and the seventh cylinder, and a second bank including the second cylinder, the fourth cylinder, the sixth cylinder and the eighth cylinder, the first exhaust passages of the third cylinder and the fifth cylinder may be led to the turbocharger without initially converging, and the first exhaust passages of the fourth cylinder and the sixth cylinder may be led to the turbocharger without initially converging.

[0026] According to this constitution, in the V-type eight cylinder engine, the first exhaust passages of the third cylinder and the fifth cylinder are led to the turbocharger without initially converging, and the first exhaust passages of the fourth cylinder and the sixth cylinder are led to the turbocharger without initially converging. Here, the respective combustion strokes of the third cylinder and fifth cylinder, which are disposed adjacent to each other within the first bank, are removed from each other by a crank angle of 180°. Similarly, the respective combustion strokes of the fourth cylinder and sixth cylinder, which are disposed adjacent to each other within the second bank, are removed from each other by a crank angle of 180°. Hence, the first exhaust passages of cylinders that are adjacent to each other within the same bank and have combustion strokes that are removed from each other by a crank angle of 180° do not converge initially upstream of the turbocharger. As a result, exhaust interference can be suppressed, and an increase in the amount of residual gas in the cylinders can be suppressed.

[0027] In the exhaust device for an internal combustion engine according to this aspect, the internal combustion engine may perform combustion at irregular intervals, each of the plurality of cylinders may include a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage, the plurality of cylinders may be constituted by a first cylinder, a second cylinder, a third cylinder, a fourth cylinder, a fifth cylinder, a sixth cylinder, a seventh cylinder, and an eighth cylinder, the internal combustion engine may be a V-type eight cylinder engine including, within a cylinder head, a first bank including the first cylinder, the third cylinder, the fifth cylinder and the seventh cylinder, and a second bank including the second cylinder, the fourth cylinder, the sixth cylinder and the eighth cylinder, the first exhaust passages of the second cylinder and the fourth cylinder may be led to the turbocharger after initially converging, and the first exhaust passages of the third cylinder and the seventh cylinder may be led to the turbocharger after initially converging. [0028] According to this constitution, in the V-type eight cylinder engine, the first exhaust passages of the second cylinder and the fourth cylinder are led to the turbocharger after initially converging, and the first exhaust passages of the third cylinder and the seventh cylinder are led to the turbocharger after initially converging. Here, combustion occurs consecutively in the fourth cylinder and the second cylinder, and combustion occurs consecutively in the seventh cylinder and the third cylinder. Hence, the exhaust pulse of the fourth cylinder and the exhaust pulse of the second cylinder converge, thereby enlarging the amplitude of the exhaust pulse. Similarly, the exhaust pulse of the seventh cylinder and the exhaust pulse of the third cylinder converge, thereby enlarging the amplitude of the exhaust pulse. As a result, the turbocharger can be driven efficiently, enabling an improvement in supercharging responsiveness during a transition period.

[0029] In the exhaust device for an internal combustion engine according to this aspect, in a predetermined operating condition, a lift amount of the second exhaust valve or a bypass-side valve overlap period, during which an open period of the second exhaust valve and an open period of an intake valve overlap, of a certain cylinder from among the plurality of cylinders may be greater than a lift amount of the second exhaust valve or a bypass-side valve overlap period, during which an open period of the second exhaust valve and an open period of an intake valve overlap, of the other cylinder of the plurality of cylinders. Further, in the exhaust device for an internal combustion engine according to this aspect, when the first exhaust valve and the second exhaust valve are both driven, an opening timing of the second exhaust valve may be made later than an opening timing of the first exhaust valve, and a closing timing of the second exhaust valve may be made later than a closing timing of the first exhaust valve. [0030] In the exhaust device for an internal combustion engine according to this aspect, the pair of cylinders may be set on the basis of timings at which the plurality of cylinders discharge exhaust gas via the first exhaust passages.

[0031] A second aspect of the present invention relates to a control device for an internal combustion engine in which combustion intervals of respective cylinders within an identical cylinder row are irregular such that the magnitude of a residual gas effect, according to which a residual gas amount becomes more likely to increase due to an exhaust pulse effect from another cylinder, differs among the cylinders. The control device for an internal combustion engine includes: a turbocharger; a first exhaust passage that passes through a turbine inlet of the turbocharger; a first exhaust valve for opening and closing an exhaust port that communicates with the first exhaust passage; a second exhaust passage that does not pass through the turbine inlet; a second exhaust valve for opening and closing an exhaust port that communicates with the second exhaust passage; a variable valve device that is capable of varying a lift amount of the second exhaust valve or a bypass-side valve overlap period, during which an open period of the second exhaust valve and an open period of an intake valve overlap, separately in cylinders exhibiting a large residual gas effect, in which the residual gas effect is large, and other cylinders; and opening characteristic control means for controlling the variable valve device such that in a predetermined operating condition, the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting a large residual gas effect is greater than the lift amount of the second exhaust valve or the bypass-side valve overlap period of the other cylinders.

[0032] According to this constitution, the first exhaust valve that communicates with the turbine inlet of the turbocharger and the second exhaust valve that does not communicate with the turbine inlet are provided in each cylinder, and therefore, during the valve overlap period between the second exhaust valve and the intake valve (the bypass-side valve overlap period), a scavenging action in which burned gas in the cylinder is chased out by intake air that has been raised in pressure through supercharging and discharged to the low-pressure second exhaust valve is obtained. In an engine having a turbocharger, in which back pressure increases easily, the amount of residual gas typically tends to increase easily, but, the amount of residual gas can be reduced sufficiently by this scavenging action. Moreover, a lift amount of the second exhaust valve or the bypass-side valve overlap period in a specific cylinder exhibiting a large residual gas effect, in which the residual gas amount increases easily, can be made larger than the lift amount of the second exhaust valve or the bypass-side valve overlap period in the other cylinders. Hence, in the cylinders exhibiting a large residual gas effect, in which the residual gas amount increases easily, a greater scavenging action than that of the other cylinders can be exhibited. As a result, an increase in the residual gas amount of the cylinders exhibiting a large residual gas effect can be suppressed reliably, whereby combustion deterioration and misfiring in the cylinders exhibiting a large residual gas effect can be avoided reliably.

[0033] In the control device for an internal combustion engine according to this aspect, the respective cylinders of the internal combustion engine may be divided into cylinders exhibiting a large residual gas effect, cylinders exhibiting an intermediate residual gas effect, in which the residual gas effect is smaller than that of the cylinders exhibiting a large residual gas effect, and cylinders exhibiting a small residual gas effect, in which the residual gas effect is smaller than that of the cylinders exhibiting an intermediate residual gas effect, the variable valve device may be made capable of varying the lift amount of the second exhaust valve or the bypass-side valve overlap period separately in the cylinders exhibiting a large residual gas effect, the cylinders exhibiting an intermediate residual gas effect, and the cylinders exhibiting a small residual gas effect, and the opening characteristic control means may control the variable valve device such that the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting a large residual gas effect is greater than the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting an intermediate residual gas effect, and such that the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting an intermediate residual gas effect is equal to or greater than the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting a small residual gas effect.

[0034] According to this constitution, the scavenging action of the cylinders exhibiting a large residual gas effect can be made greater than the scavenging action of the cylinders exhibiting an intermediate residual gas effect, and the scavenging action of the cylinders exhibiting an intermediate residual gas effect can be made equal to or greater than the scavenging action of the cylinders exhibiting a small residual gas effect. Hence, the magnitude of the scavenging action in each cylinder can be controlled appropriately in accordance with the likelihood of residual gas remaining in each cylinder. As a result, the residual gas amount in each cylinder can be reduced and evened out at the same time, and fresh air can be prevented from blowing through the second exhaust passage.

[0035] The control device for an internal combustion engine according to this aspect may further include: second exhaust valve stopping means capable of setting the lift amount of the second exhaust valve to zero separately in the cylinders exhibiting a large residual gas effect and the other cylinders; and lift amount switching sequence control means which, when switching the lift amount of the second exhaust valve in each cylinder to zero, switches the lift amount of the second exhaust valve in the other cylinders to zero first, and switches the lift amount of the second exhaust valve in the cylinders exhibiting a large residual gas effect to zero thereafter.

[0036] According to this constitution, when the lift amount of the second exhaust valve in each cylinder is switched to zero, the lift amount of the second exhaust valve in the cylinders other than the cylinders exhibiting a large residual gas effect can be switched to zero first such that the lift amount of the second exhaust valve in the cylinders exhibiting a large residual gas effect is switched to zero thereafter. When the lift amount of the second exhaust valve of each cylinder is switched to zero, the

^■ scavenging action is no longer exhibited, and when the second exhaust valve is halted, an exhaust gas flow rate to the first exhaust valve side increases rapidly, and therefore the back pressure on the first exhaust valve side and the amplitude of the exhaust pulse become more likely to experience a large transient increase. Hence, during a transitional operation in which the lift amount of the second exhaust valve in each cylinder is switched to zero, the residual gas amount in the cylinders exhibiting a large residual gas effect is particularly likely to increase in comparison with a steady operation. On the other hand, the scavenging action can be exhibited by continuing to drive the second exhaust valves of the cylinders exhibiting a large residual gas effect until the transient large-amplitude exhaust pulse attenuates, and therefore a transient increase in the residual gas amount of the cylinders exhibiting a large residual gas effect can be suppressed reliably. [0037] Further, in the control device for an internal combustion engine according to this aspect, the lift amount switching sequence control means may switch the lift amount of the second exhaust valve in a certain cylinder, from among the other cylinders, that does not have an exhaust pulse effect on the cylinders exhibiting a large residual gas effect to zero first. [0038] According to this constitution, when the lift amounts of the second exhaust valves in the cylinders other than the cylinders exhibiting a large residual gas effect are switched to zero preferentially, the lift amount of the second exhaust valve in any cylinder other than a cylinder that has an exhaust pulse effect on a cylinder exhibiting a large residual gas effect can be switched to zero first. In so doing, an excessive increase in the exhaust pulse received by the cylinder exhibiting a large residual gas effect can be suppressed, and therefore a transient increase in the residual gas amount of the cylinder exhibiting a large residual gas effect can be suppressed reliably.

[0039] Further, in the control device for an internal combustion engine according to this aspect, in the cylinders exhibiting a large residual gas effect and cylinders that have an exhaust pulse effect on the cylinders exhibiting a large residual gas effect, the respective first exhaust valves thereof or the respective exhaust ports that pass through the second exhaust valves thereof may converge within a cylinder head.

[0040] According to this constitution, in the cylinders exhibiting a large residual gas effect and the cylinders that have an exhaust pulse effect on the cylinders exhibiting a large residual gas effect, the respective first exhaust valves thereof or the respective exhaust ports that pass through the second exhaust valves thereof converge within the cylinder head. With this constitution, the surface area of the exhaust passage can be reduced, and thermal energy dissipation from the surface of the exhaust passage can be suppressed. As a result, an improvement in turbo efficiency or early warm-up of a catalyst can be achieved. On the other hand, these cylinders are more likely to receive an exhaust pulse effect, and therefore the residual gas amount in the cylinders exhibiting a large residual gas effect is particularly likely to increase. According to the present invention, however, an increase in the residual gas amount of the cylinders exhibiting a large residual gas effect can be suppressed reliably, and therefore combustion deterioration and misfiring in the cylinders exhibiting a large residual gas effect can be avoided reliably.

[0041] Further, in the control device for an internal combustion engine according to this aspect, positions of the first exhaust valve and the second exhaust valve may be reversed in adjacent cylinders of an identical cylinder row, either the first exhaust valves or the second exhaust valves of two adjacent cylinders may be positioned adjacent to each other, exhaust ports passing respectively through the two second exhaust valves positioned adjacent to each other may converge within the cylinder head, and exhaust ports passing respectively through the two first exhaust valves positioned adjacent to each other may converge within the cylinder head.

[0042] According to this constitution, the exhaust ports passing through the first exhaust valves or the exhaust ports passing through the second exhaust valves of two adjacent cylinders converge within the cylinder head. With this constitution, the surface area of the exhaust passage can be reduced, and thermal energy dissipation from the surface of the exhaust passage can be suppressed. As a result, an improvement in turbo efficiency or early warm-up of a catalyst can be achieved. On the other hand, these cylinders are more likely to receive an exhaust pulse effect, and therefore the residual gas amount in the cylinder exhibiting a large residual gas effect is particularly likely to increase. According to the present invention, however, an increase in the residual gas amount of the cylinder exhibiting a large residual gas effect can be suppressed reliably, and therefore combustion deterioration and misfiring in the cylinder exhibiting a large residual gas effect can be avoided reliably.

[0043] Further, in the control device for an internal combustion engine according to this aspect, when the first exhaust valve and the second exhaust valve are both driven, an opening timing of the second exhaust valve may be made later than an opening timing of the first exhaust valve, and a closing timing of the second exhaust valve may be made later than a closing timing of the first exhaust valve.

[0044] According to this constitution, when the first exhaust valve and the second exhaust valve are both driven, the opening timing of the second exhaust valve can be made later than the opening timing of the first exhaust valve, and the closing timing of the second exhaust valve can be made later than the closing timing of the first exhaust valve. Hence, the scavenging action can be exhibited easily, and a backflow of exhaust gas from the high-back pressure first exhaust valve to the cylinder can be suppressed. As a result, the residual gas amount can be reduced even more reliably.

BRIEF DESCRIPTION OF THE DRAWINGS

[0045] The foregoing and further objects, features and advantages of the invention will become apparent from the following description of example embodiments with reference to the accompanying drawings, wherein like numerals are used to represent like elements and wherein:

FIG. 1 is a view illustrating a system configuration according to a first embodiment of the present invention;

FIG. 2 is a view showing an example of valve timings in a V-type eight cylinder engine according to the first embodiment of the present invention;

FIG. 3 is a view showing an exhaust pulse in each cylinder of the V-type eight cylinder engine according to the first embodiment of the present invention;

FIG. 4 is a view illustrating a system configuration according to a modified example of the first embodiment of the present invention; FIGS. 5A, 5B, 5 C, 5D, 5E and 5F are schematic diagrams showing the constitution of exhaust passages communicating with turbochargers 4OA, 4OB in a second embodiment of the present invention;

FIG. 6 is a schematic diagram showing the constitution of an exhaust passage communicating with a turbocharger 4OB according to a third embodiment of the present invention;

FIG. 7 is a view showing exhaust pulse convergence according to the third embodiment of the present invention;

FIG. 8 is a schematic diagram showing the constitution of an exhaust passage communicating with a turbocharger 4OB according to a modified example of the third embodiment of the present invention;

FIG. 9 is a schematic plan view illustrating an exhaust system of a V-type eight cylinder engine according to a fourth embodiment of the present invention;

FIG. 10 is a block diagram showing a system configuration according to the fourth embodiment of the present invention;

FIG. 11 is a valve lift diagram of an intake valve, a first exhaust valve EXl, and a second exhaust valve EX2 during a cold start (a catalyst warm-up region), according to the fourth embodiment of the present invention;

FIG. 12 is a valve lift diagram of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 in a high output region of a stoichiometric combustion mode, according to the fourth embodiment of the present invention;

FIG. 13 is a valve lift diagram of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 in a fuel efficiency improvement region of a lean combustion mode, according to the fourth embodiment of the present invention; FIG. 14 is a view showing an operating region of the engine according to the fourth embodiment of the present invention;

FIG. 15 is a valve lift diagram of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 in regions B and D, according to the fourth embodiment of the present invention; FIG. 16 is a flowchart showing a routine executed in the fourth embodiment of the present invention;

FIG. 17 is a flowchart showing a routine executed in a fifth embodiment of the present invention;

FIG. 18 is a schematic plan view showing a V-type eight cylinder engine according to the related art; and

FIG. 19 is a view showing relationships between a crank angle and the working strokes of each cylinder in the V-type eight cylinder engine according to the related art.

DETAILED DESCRIPTION OF EMBODIMENTS

[0046] Embodiments of the present invention will be described below with reference to the drawings. Note that elements common to the drawings have been allocated identical reference symbols, and duplicate description thereof has been omitted.

[0047] FIG. 1 is a view illustrating a system configuration according to a first embodiment of the present invention. The system shown in FIG. 1 includes an engine 1 having a plurality of cylinders 2. The engine 1 is a V-type eight cylinder engine, for example. A cylinder head 4 of the engine 1 includes a first bank 6 and a second bank 8.

[0048] The first bank 6 and second bank 8 are disposed at an incline having a predetermined angle. A first cylinder #1, a third cylinder #3, a fifth cylinder #5, and a seventh cylinder #7 are disposed in series in the first bank 6. A second cylinder #2, a fourth cylinder #4, a sixth cylinder #6, and an eighth cylinder #8 are disposed in series in the second bank 8.

[0049] A combustion sequence of the engine 1 is first cylinder #1 → eighth cylinder #8 → seventh cylinder #7 — > third cylinder #3 → sixth cylinder #6 → fifth cylinder #5 → fourth cylinder #4 → second cylinder #2. Hence, the ignition intervals in the respective banks 6, 8 are irregular.

[0050] Each cylinder 2 includes a first exhaust valve HA provided in a first exhaust port 1OA that passes through a turbine 41A, 41B of a turbocharger 4OA, 4OB, and a second exhaust valve HB provided in a second exhaust port 1OB that passes downstream of the turbine.

[0051] The first exhaust ports 1OA of adjacent cylinders 2 converge in the cylinder head 4. Note, however, that the first exhaust port 1OA of a certain cylinder and the first exhaust port 1OA of another cylinder having a combustion stroke which is removed from the combustion stroke of the certain cylinder by a crank angle of 180° do not converge in the cylinder head 4. As will be described below, the reason for this is to suppress a reduction in an exhaust pulse led to the turbocharger 4OA, 4OB.

[0052] More specifically, as shown in FIG. 1, the first exhaust ports 1OA of the adjacent first cylinder #1 and third cylinder #3 converge in the cylinder head 4. One end of an exhaust passage 21 is connected to a convergence point between the first exhaust ports 1OA. The other end of the exhaust passage 21 is led to the turbine 41A of the turbocharger 4OA. A compressor 42A connected to the turbine 41A is provided in an intake passage, not shown in the drawing.

[0053] Further, the first exhaust ports 1OA of the adjacent fifth cylinder #5 and seventh cylinder #7 converge in the cylinder head 4. One end of an exhaust passage 22 is connected to a convergence point between the first exhaust ports 1OA. The other end of the exhaust passage 22 is led to the turbine 41A. The exhaust passage 21 and the exhaust passage 22 converge immediately before the turbine 41A. A start-up catalyst 46A is provided on an exhaust passage 44A downstream of the turbine 41A. The start-up catalyst 46 A is a conventional oxidation catalyst, for example. A NOx catalyst, not shown in the drawing, is provided downstream of the start-up catalyst 46A.

[0054] Meanwhile, one end of an exhaust passage 23 is connected to the second exhaust port 1OB of the first cylinder #1. The second exhaust ports 1OB of the third cylinder #3 and fifth cylinder #5 converge in the cylinder head 4. One end of an exhaust passage 24 is connected to a convergence point between the second exhaust ports 1OB. Further, one end of an exhaust passage 25 is connected to the second exhaust port 1OB of the seventh cylinder #7. The other ends of the respective exhaust passages 23, 24, 25 converge to form an exhaust passage 26. The exhaust passage 26 is connected to the exhaust passage 44A downstream of the turbine 41A. In other words, the exhaust passage 26 is connected to the exhaust passage 44A between the turbine 41A and the start-up catalyst 46A.

[0055] Further, the first exhaust ports 1OA of the second cylinder #2 and fourth cylinder #4 converge in the cylinder head 4. One end of an exhaust passage 31 is connected to a convergence point between the first exhaust ports 1OA. The other end of the exhaust passage 31 is led to the turbine 41B of the turbocharger 4OB. Further, the first exhaust ports 1OA of the sixth cylinder #6 and eighth cylinder #8 converge in the cylinder head 4. One end of an exhaust passage 32 is connected to a convergence point between the first exhaust ports 1OA. The other end of the exhaust passage 32 is led to the turbine 41B. The exhaust passage 31 and the exhaust passage 32 converge immediately before the turbine 41B. An oxidation catalyst serving as a start-up catalyst

46B is provided on an exhaust passage 44B downstream of the turbine 41B. A NOx catalyst, not shown in the drawing, is provided downstream of the start-up catalyst 46B.

[0056] Meanwhile, one end of an exhaust passage 33 is connected to the second exhaust port 1OB of the second cylinder Wl. The second exhaust ports 1OB of the fourth cylinder #4 and sixth cylinder #6 converge in the cylinder head 4. One end of an exhaust passage 34 is connected to the convergence point between the second exhaust ports 1OB. Further, one end of an exhaust passage 35 is connected to the second exhaust port 1OB of the eighth cylinder #8. The other ends of the respective exhaust passages 33, 34, 35 converge to form an exhaust passage 36. The exhaust passage 36 is connected to the exhaust passage 44B downstream of the turbine 41B. In other words, the exhaust passage 36 is connected to the exhaust passage 44B between the turbine 41B and the start-up catalyst 46B.

[0057] In the system according to the first embodiment, the first exhaust valve HA and an intake valve (not shown) are opened and closed at timings such as those shown in FIG. 2, for example. FIG. 2 is a view showing an example of valve timings in a V-type eight cylinder engine. Here, when the first exhaust valve HA of an arbitrary cylinder is open (in particular, at a closing timing) and a pressure wave (to be referred to hereafter as an "exhaust pulse") generated by the start of exhaust from another cylinder arrives, the exhaust action of the arbitrary cylinder is impaired. As a result of this so-called exhaust interference, the exhaust action of the arbitrary cylinder can no longer be performed efficiently. Hence, increases occur in the internal pressure of the cylinder and the amount of residual gas in the cylinder. As a result, a combustion speed may decrease such that it becomes impossible to obtain a favorable combustion condition. Accordingly, a reduction in output and deterioration of the fuel efficiency may occur. When a valve overlap in which both the first exhaust valve HA and the intake valve are open exists, as shown in FIG. 2, the amount of residual gas in the cylinder increases even further, with the result that the exhaust interference problem becomes even more serious. [0058] Further, the turbocharger 4OA, 4OB is designed to increase in efficiency in accordance with the exhaust pulse. Hence, when the exhaust pulse of a certain cylinder is reduced by the exhaust pulse of another cylinder, the exhaust pulse led to the turbocharger 4OA, 4OB deteriorates, and as a result, the supercharging responsiveness of the turbocharger 4OA, 4OB during a transition period may deteriorate. [0059] FIG. 3 is a view showing the exhaust pulse of each cylinder in the V-type eight cylinder engine. As shown in FIG. 3, the combustion sequence of the V-type eight cylinder engine is first cylinder #1 → eighth cylinder #8 → seventh cylinder #7 → third cylinder #3 → sixth cylinder #6 → fifth cylinder #5 → fourth cylinder #4 → second cylinder #2, and the combustion interval is 90° CA. [0060] The arrows in FIG. 3 denote working angles of the first exhaust valves HA of the first cylinder #1, eighth cylinder #8, and seventh cylinder #7. ϊn a typical V-type eight cylinder engine, the working angle of the first exhaust valve HA is controlled between 220° and 270° CA. This working angle control can be realized by an ECU (Electronic Control Unit) and a variable valve mechanism, not shown in the drawing, which are installed in the engine 1. The variable valve mechanism is connected to the first exhaust valve HA, and may be a conventional hydraulic or mechanical variable valve mechanism or an electromagnetically driven valve mechanism.

[0061] As shown in FIG. 3, in a V-type eight cylinder engine, combustion and exhaust are performed three times each while the first exhaust valve HA of one cylinder is open. More specifically, as shown in FIG. 3, while the first exhaust valve HA of the first cylinder #1 is open, the first exhaust valves HA of the eighth cylinder #8 and seventh cylinder #7 open in succession. As a result, the exhaust action of the first cylinder #1 may be impaired by the exhaust actions of the eighth cylinder #8 and seventh cylinder #7. However, as shown in FIG. 3, when the first exhaust valve HA of the first cylinder #1 closes, the first exhaust valve HA of the seventh cylinder #7 is open. Therefore, the exhaust action of the seventh cylinder #7 has a greater effect on the exhaust action of the first cylinder #1 than the exhaust action of the eighth cylinder #8.

[0062] Hence, to suppress exhaust interference in the first cylinder #1, the exhaust pulse action of the seventh cylinder #7 is preferably suppressed. More specifically, the exhaust pulse action of the seventh cylinder #7 having a combustion stroke that is later than the combustion stroke of the first cylinder #1 by 180° CA is preferably suppressed.

This applies similarly to other cylinder groups having combustion strokes that are removed from each other by 180° CA, namely the eighth cylinder #8 and third cylinder #3, the seventh cylinder #7 and sixth cylinder #6, the third cylinder #3 and fifth cylinder

#5, the sixth cylinder #6 and fourth cylinder #4, the fifth cylinder #5 and second cylinder

#2, the fourth cylinder #4 and first cylinder #1, and the second cylinder #2 and eighth cylinder #8.

[0063] In the case of cylinders disposed in different banks, such as the eighth cylinder #8 and the third cylinder #3, for example, the exhaust pulse attenuates, and therefore exhaust interference does not pose a problem. Likewise, exhaust interference does not pose a problem in the seventh cylinder #7 and sixth cylinder #6, the fifth cylinder #5 and second cylinder #2, and the fourth cylinder #4 and first cylinder #1.

[0064] Hence, the cylinder groups in which exhaust interference poses an actual problem are the first cylinder #1 and seventh cylinder #7, the third cylinder #3 and fifth cylinder #5, the sixth cylinder #6 and fourth cylinder #4, and the second cylinder #2 and eighth cylinder #8.

[0065] In the system shown in FIG. 1, the first exhaust port 1OA of the first cylinder #1 converges with the first exhaust port 1OA of the third cylinder #3 within the cylinder head 4. Accordingly, the first exhaust port 1OA of the first cylinder #1 and the first exhaust port 1OA of the seventh cylinder #7 do not converge within the cylinder head 4. In other words, the first exhaust port 1OA of the first cylinder #1 is led to the turbine 41A without initially converging with the first exhaust port 1OA of the seventh cylinder #7 having a combustion stroke that is removed therefrom by 180° CA. [0066] Further, the first exhaust port 1OA of the third cylinder #3 and the first exhaust port 1OA of the fifth cylinder #5 do not converge within the cylinder head 4. In other words, the first exhaust port 1OA of the third cylinder #3 is led to the turbine 41A without initially converging with the first exhaust port 1OA of the fifth cylinder #5. [0067] This applies similarly to the second bank 8. As shown in FIG. 1, the first exhaust port 1OA of the sixth cylinder #6 and the first exhaust port 1OA of the fourth cylinder #4 do not converge within the cylinder head 4. In other words, the first exhaust port 1OA of the sixth cylinder #6 is led to the turbine 41B without initially converging with the first exhaust port 1OA of the fourth cylinder #4. [0068] Further, the first exhaust port 1OA of the second cylinder #2 and the first exhaust port 1OA of the eighth cylinder #8 do not converge within the cylinder head 4.

In other words, the first exhaust port 1OA of the second cylinder #2 is led to the turbine

41B without initially converging with the first exhaust port 1OA of the eighth cylinder #8.

[0069] As described above, in the first embodiment, the first exhaust ports 1OA of the first cylinder #1 and third cylinder #3 and the first exhaust ports 1OA of the fifth cylinder #5 and seventh cylinder #7, which are respectively adjacent to each other within the first bank 6, converge initially. Further, the first exhaust ports 1OA of the second cylinder #2 and fourth cylinder #4 and the first exhaust ports 1OA of the sixth cylinder #6 and eighth cylinder #8, which are respectively adjacent to each other within the second bank 8, converge initially. In other words, the first exhaust port 1OA of a certain cylinder does not converge initially with the first exhaust port 1OA of another cylinder having a combustion stroke that is removed therefrom by 180° CA, and therefore exhaust interference in the certain cylinder can be suppressed. Hence, an increase in the amount of residual gas within the cylinder can be suppressed, and a favorable combustion condition can be obtained. As a result, output and fuel efficiency can be improved. Moreover, a reduction in the exhaust pulse of the turbochargers 4OA, 4OB can be suppressed, and therefore an improvement in the supercharging responsiveness during a transition period can be achieved. Further, by causing the first exhaust ports 1OA to converge within the cylinder head 4 as shown in FIG. 1, maximum compactness can be achieved in the layout of the exhaust passages.

[0070] Incidentally, in the system shown in FIG. 1, the exhaust passages 21, 22 converge before being led to the turbine 41A, and the exhaust passages 31, 32 converge before being led to the turbine 41B. In other words, an inlet part of the turbocharger 4OA serves as a convergence portion of the exhaust passages 21, 22, while an inlet part of the turbocharger 4OB serves as a convergence portion of the exhaust passages 31, 32. Furthermore, in the system shown in FIG. 1, the exhaust ports 1OA, 1OB are provided independently in each cylinder. Hence, the exhaust passage volume from the combustion chamber to the turbines 41A, 41B is small, and therefore, even though the exhaust pulse increases, it is still likely to be affected by the exhaust pulse of another cylinder (in other words, exhaust interference is likely to occur). To avoid this, the exhaust passage 21 and the exhaust passage 22 may be led to the turbine 41 A independently and the exhaust passage 31 and the exhaust passage 32 may be led to the turbine 41B independently, as shown in FIG. 4. FIG. 4 is a view illustrating the constitution of a system according to a modified example of the first embodiment. As shown by the broken line circles in FIG. 4, the turbochargers 4OA, 4OB in this modified example have a twin-entry constitution. As a result, exhaust interference can be reduced even further than in the first embodiment while obtaining a similar exhaust pulse increase effect to that of the first embodiment. [0071] Further, in the first embodiment, a case was described in which the combustion sequence of the engine 1 is first cylinder #1 → eighth cylinder #8 → seventh cylinder #7 → third cylinder #3 → sixth cylinder #6 → fifth cylinder #5 → fourth cylinder #4 → second cylinder #2. However, the present invention is not limited to this sequence, and may be applied to a case in which the combustion sequence is first cylinder #1 → eighth cylinder #8 → fourth cylinder #4 → third cylinder #3 → sixth cylinder #6 → fifth cylinder #5 → seventh cylinder #7 → second cylinder #2 (likewise in second and third embodiments to be described below). Similar effects to those of the first embodiment can also be obtained in this case.

[0072] Note that in the first embodiment, the first exhaust port 1OA is an example of a "first exhaust passage", the second exhaust port 1OB is an example of a "second exhaust passage", and the engine 1 is an example of an "internal combustion engine".

[0073] Next, referring to FIGS. 5A to 5F, a second embodiment of the present invention will be described. In the first embodiment, the first exhaust ports 1OA of adjacent cylinders converge initially within the cylinder head 4. More specifically, the respective first exhaust ports 1OA of the first cylinder #1 and third cylinder #3, the second cylinder #2 and fourth cylinder #4, the fifth cylinder #5 and seventh cylinder #7, and the sixth cylinder #6 and eighth cylinder #8 converge initially. Thus, the first exhaust ports 1OA of each group of cylinders having combustion strokes that are removed from each other by 180° CA within the same bank 6, 8, such as the first cylinder #1 and seventh cylinder #7, for example, are prevented from converging initially within the cylinder head 4.

[0074] Incidentally, although the respective combustion strokes of the first cylinder #1 and seventh cylinder #7 are removed from each other by 180° CA, these two cylinders #1, #7 are disposed at either end of the first bank 6, and therefore the degree of exhaust interference therebetween can be kept comparatively small even when the respective first exhaust ports 1OA thereof are caused to converge at a minimum length. The reason for this is that at the closing timing of the first exhaust valve HA of the first cylinder #1, a comparatively small exhaust pressure wave arrives from the seventh cylinder #7. This applies similarly to the second cylinder #2 and eighth cylinder #8 disposed at either end of the second bank 8.

[0075] In contrast, not only are the respective combustion strokes of the third cylinder #3 and fifth cylinder #5 removed from each other by 180° CA, but also, these two cylinders #3, #5 are disposed adjacent to each other within the first bank 6. When the first exhaust ports 1OA of the third cylinder #3 and fifth cylinder #5 are caused to converge at a minimum length, the degree of exhaust interference is far greater than when the first exhaust ports 1OA of the first cylinder #1 and seventh cylinder #7 are caused to converge at a minimum length. The reason for this is that at the closing timing of the first exhaust valve HA of the third cylinder #3, a large exhaust pressure wave arrives from the fifth cylinder #5.

[0076] Likewise, not only are the respective combustion strokes of the fourth cylinder #4 and sixth cylinder #6 removed from each other by 180° CA, but also, these two cylinders #4, #6 are disposed adjacent to each other within the second bank 8. When the first exhaust ports 1OA of the fourth cylinder #4 and sixth cylinder #6 are caused to converge at a minimum length, a large exhaust pressure wave arrives from the fourth cylinder #4 at the closing timing of the first exhaust valve of the sixth cylinder #6, and therefore the degree of exhaust interference increases greatly.

[0077] Hence, in the second embodiment, as shown in FIGS. 5 A to 5F, exhaust passages 52, 53 of the third cylinder #3 and fifth cylinder #5 in the V-type eight cylinder engine are led to the turbocharger 4OA without initially converging. Further, exhaust passages 56, 51 of the fourth cylinder #4 and sixth cylinder #6 are led to the turbocharger

4OB without initially converging.

[0078] FIGS. 5 A to 5F are schematic diagrams showing the constitution of exhaust passages communicating with the turbochargers 4OA, 4OB in the second embodiment.

More specifically, FIGS. 5A and 5B, FIGS. 5C and 5D, and FIGS. 5E and 5F show the constitutions of exhaust passages according to a first modified example, a second modified example, and a third modified example of the second embodiment, respectively.

Note that in FIGS. 5A to 5F, the second exhaust port provided with the second exhaust valve HB, the exhaust passage connected to the second exhaust port, and so on have been omitted. The omitted parts are similar to their counterparts in FIG. 1, and are not therefore described here.

[0079] In the first modified example of the second embodiment, as shown in FIG. 5 A, exhaust passages 51, 52 of the first cylinder #1 and third cylinder #3 in the first bank 6 converge initially, whereupon a converged exhaust passage 61 is led to the turbocharger 4OA. Further, exhaust passages 53, 54 of the fifth cylinder #5 and seventh cylinder #7 converge initially, whereupon a converged exhaust passage 62 is led to the turbocharger 4OA.

[0080] This applies likewise to the second bank 8. More specifically, as shown in FIG. 5B, exhaust passages 55, 56 of the second cylinder #2 and fourth cylinder #4 converge initially, whereupon a converged exhaust passage 63 is led to the turbocharger 4OB. Further, exhaust passages 57, 58 of the sixth cylinder #6 and eighth cylinder #8 converge initially, whereupon a converged exhaust passage 64 is led to the turbocharger 4OB.

[0081] Note that in the first modified example, when the exhaust passages 51, 52, the exhaust passages 53, 54, the exhaust passages 55, 56 and the exhaust passages 57, 58 are respectively caused to converge at a minimum length (i.e. within the cylinder head 4), an identical constitution to that of the first embodiment is obtained. In this case, the exhaust passages 51 to 58 correspond to the first exhaust ports 1OA. Further, the exhaust passages 61, 62 may be converged and then led to the turbocharger 4OA as a converged exhaust passage, rather than being led to the turbocharger 4OA independently (i.e. rather than in a twin-entry constitution). Similarly, the exhaust passages 63, 64 may be converged and then led to the turbocharger 4OB as a converged exhaust passage. [0082] In the second modified example of the second embodiment, as shown in FIG.

5C, the exhaust passages 51, 53, 54 of the first cylinder #1, fifth cylinder #5, and seventh cylinder #7 in the first bank 6 converge initially, whereupon a converged exhaust passage 65 is led to the turbocharger 4OA. Meanwhile, the exhaust passage 52 of the third cylinder #3 is led to the turbocharger 4OA without converging with another exhaust passage.

[0083] This applies likewise to the second bank 8. More specifically, as shown in FIG. 5D, the exhaust passages 55, 57, 58 of the second cylinder #2, sixth cylinder #6, and eighth cylinder #8 converge initially, whereupon a converged exhaust passage 66 is led to the turbocharger 4OB. Meanwhile, the exhaust passage 56 of the fourth cylinder #4 is led to the turbocharger 4OB without converging with another exhaust passage.

[0084] Note that the exhaust passages 52, 65 may be converged and then led to the turbocharger 4OA as a converged exhaust passage, rather than being led to the turbocharger 4OA independently (i.e. rather than in a twin-entry constitution). Similarly, the exhaust passages 56, 66 may be converged and then led to the turbocharger 4OB as a converged exhaust passage.

[0085] In the third modified example of the second embodiment, as shown in FIG. 5E, the exhaust passages 51, 53 of the first cylinder #1 and fifth cylinder #5 in the first bank 6 converge initially, whereupon a converged exhaust passage 67 is led to the turbocharger 4OA. Further, the exhaust passages 52, 54 of the third cylinder #3 and seventh cylinder #7 converge initially, whereupon a converged exhaust passage 68 is led to the turbocharger 4OA.

[0086] This applies likewise to the second bank 8. More specifically, as shown in FIG. 5F, the exhaust passages 55, 57 of the second cylinder #2 and sixth cylinder #6 converge initially, whereupon a converged exhaust passage 69 is led to the turbocharger 4OB. Further, the exhaust passages 56, 58 of the fourth cylinder #4 and eighth cylinder #8 converge initially, whereupon a converged exhaust passage 70 is led to the turbocharger 4OB.

[0087] Note that the exhaust passages 67, 68 may be converged and then led to the turbocharger 40A as a converged exhaust passage, rather than being led to the turbocharger 4OA independently (i.e. rather than in a twin-entry constitution). Similarly, the exhaust passages 69, 70 may be converged and then led to the turbocharger 4OB as a converged exhaust passage.

[0088] As described above, in the second embodiment, the exhaust passages 52, 53 of the third cylinder #3 and fifth cylinder #5 in the first bank 6 are prevented from converging initially, and the exhaust passages 56, 57 of the fourth cylinder #4 and sixth cylinder #6 in the second bank 8 are prevented from converging initially. Hence, exhaust interference between cylinders disposed adjacent to each other in the same bank and having combustion strokes that are removed from each other by 180° CA can be suppressed. Therefore, similarly to the first embodiment, an increase in the amount of residual gas within the cylinder can be suppressed, and a favorable combustion condition can be obtained. As a result, output and fuel efficiency can be improved.

[0089] Note that in the second embodiment, the exhaust passages 52, 53 serve as an example of "first exhaust passages of the third cylinder and fifth cylinder", the exhaust passages 56, 57 serve as an example of "first exhaust passages of the fourth cylinder and sixth cylinder", and the exhaust passages 52, 54 serve as an example of "first exhaust passages of the third cylinder and seventh cylinder".

[0090] Next, referring to FIGS. 6 and 7, a third embodiment of the present invention will be described. FIG. 6 is a schematic diagram showing the constitution of exhaust passages communicating with the turbocharger 4OB in the third embodiment. The constitution shown in FIG. 6 is identical to the exhaust constitution of the second bank 8 shown in FIG. 5B.

[0091] As described above, the performance of the turbochargers 4OA, 4OB is determined by the peak pressure of the exhaust pulse, or in other words the amplitude of the exhaust pulse.

[0092] In the third embodiment, an aspect in which the amplitude of the exhaust pulse is increased by raising the peak pressure of the exhaust pulse will be described. In the third embodiment, a V-type eight cylinder engine is employed as the internal combustion engine. As described above, the combustion sequence of a V-type eight cylinder engine is first cylinder #1 → eighth cylinder #8 → seventh cylinder #7 → third cylinder #3 → sixth cylinder #6 → fifth cylinder #5 → fourth cylinder #4 → second cylinder #2 (see FIG. 3). Hence, in the second bank 8, the fourth cylinder #4 and the second cylinder #2 combust consecutively. [0093] Therefore, in the third embodiment, the exhaust passages 55, 56 of the consecutively combusting second cylinder #2 and fourth cylinder #4 of the second bank 8 converge initially, as shown in FIG. 6. The converged exhaust passage 63 formed as a result of convergence between the exhaust passages 55, 56 is then led to the turbocharger 4OB. [0094] Thus, as shown by a dot-dash line in FIG. 7, the exhaust pulses of the consecutive fourth cylinder #4 and second cylinder #2 converge, leading to an increase in the peak pressure of the exhaust pulse. FIG. 7 is a view showing exhaust pulse convergence in the third embodiment. As shown in FIG. 7, the amplitude of the converged exhaust pulse is greater than the amplitude of the non-converged exhaust pulses of the eighth cylinder #8 and the sixth cylinder #6. As a result, the turbocharger 4OB can be driven efficiently, and therefore the performance of the turbocharger 4OB, such as the supercharging responsiveness thereof during a transition period, can be improved. [0095] In the third embodiment, an aspect in which the exhaust pulses of the fourth cylinder #4 and second cylinder #2 of the second bank 8 are converged was described, but the exhaust pulses of the consecutively combusting seventh cylinder #7 and third cylinder #3 of the first bank 6 may also be converged. More specifically, as shown in FIG. 5E, the exhaust passages 52, 54 of the third cylinder #3 and seventh cylinder #7 are converged initially. Thus, the exhaust pulses of the consecutive seventh cylinder #7 and third cylinder #3 are converged, although this is not shown in the drawings. As a result, the amplitude of the converged exhaust pulses becomes greater than that of the non-converged exhaust pulses of the first cylinder #1 and fifth cylinder #5. Accordingly, the turbocharger 4OA can be driven efficiently, and therefore the performance of the turbocharger 40A, such as the transient responsiveness thereof, can be improved.

[0096] Further, in the third embodiment, the exhaust passage 63 and the exhaust passage 64 are led to the turbocharger 4OB without being converged, but these exhaust passages 63, 64 may be converged immediately before the turbocharger 4OB. Further, as shown in FIG. 8, the exhaust passages 55 to 58 may be converged, whereupon a converged exhaust passage 71 is led to the turbocharger 4OB. FIG. 8 is a schematic diagram showing the constitution of exhaust passages communicating with the turbocharger 4OB in a modified example of the third embodiment. In these cases, the exhaust passage volume increases, and therefore the degree of amplitude enlargement due to exhaust pulse convergence may decrease. [0097] Note that in the third embodiment, the exhaust passages 55, 56 serve as an example of the "first exhaust passages of the second cylinder and fourth cylinder".

[0098] FIG. 9 is a schematic plan view illustrating the exhaust system of a V-type eight cylinder engine according to a fourth embodiment of the present invention. Note that FIG. 9 omits the intake system. Further, numerals preceded by # in the drawing indicate cylinder numbers.

[0099] A V-type eight cylinder engine (to be referred to simply as "engine" hereafter)

10 shown in FIG. 9 includes a left-hand cylinder row (left bank) 12L and a right-hand cylinder row (right bank) 12R. The left-hand cylinder row 12L is constituted by first, third, fifth and seventh cylinders, while the right-hand cylinder row 12R is constituted by second, fourth, sixth and eighth cylinders.

[0100] The engine 10 is provided with turbochargers 14L, 14R. In this embodiment, the turbocharger 14L relating to the left-hand cylinder row 12L and the turbocharger 14R relating to the right-hand cylinder row 12R are provided separately. The turbochargers 14L, 14R each include a turbine 14a and a compressor 14b. The turbine 14a of the turbocharger 14L, 14R is activated by exhaust gas. The compressor 14b is driven by the turbine 14a to be capable of compressing intake air.

[0101] The engine 10 is also provided with first exhaust passages 16L, 16R that pass through an inlet of the turbine 14a in the turbochargers 14L, 14R, and second exhaust passages 18L, 18R that do not pass through the inlet of the turbine 14a. Each cylinder of the engine 10 is provided with two exhaust valves, namely a first exhaust valve EXl and a second exhaust valve EX2. The first exhaust valve EXl opens and closes an exhaust port communicating with the first exhaust passage 16L, 16R, while the second exhaust valve EX2 opens and closes an exhaust port communicating with the second exhaust passage 18L, 18R.

[0102] Catalysts 2OL, 2OR are provided in the exhaust system to purify harmful components contained in the exhaust gas. More specifically, on the left-hand cylinder row 12L side, the exhaust passage on the downstream side of the turbine 14a of the turbocharger 14L and the second exhaust passage 18L are connected to the catalyst 2OL. On the right-hand cylinder row 12R side, the exhaust passage on the downstream side of the turbine 14a of the turbocharger 14R and the second exhaust passage 18R are connected to the catalyst 2OR.

[0103] In the engine 10 of this embodiment, the positions of the first exhaust valve EXl and the second exhaust valve EX2 are reversed in adjacent cylinders of the same cylinder row. Hence, in adjacent cylinders, either first exhaust valves EXl or second exhaust valves EX2 are positioned adjacent to each other. More specifically, in the left-hand cylinder row 12L, the respective second exhaust valves EX2 of the first cylinder and third cylinder are adjacent, the respective first exhaust valves EXl of the third cylinder and fifth cylinder are adjacent, and the respective second exhaust valves EX2 of the fifth cylinder and seventh cylinder are adjacent. In the right-hand cylinder row 12R, the respective second exhaust valves EX2 of the second cylinder and fourth cylinder are adjacent, the respective first exhaust valves EXl of the fourth cylinder and sixth cylinder are adjacent, and the respective second exhaust valves EX2 of the sixth cylinder and eighth cylinder are adjacent.

[0104] Exhaust ports of the two adjacent exhaust valves of the same type converge within respective cylinder heads 22L, 22R and open onto a side face of the cylinder head 22L, 22R as a single exhaust outlet. For example, an exhaust port 224 communicating with the first exhaust valve EXl of the third cylinder and an exhaust port 226 communicating with the first exhaust valve EXl of the fifth cylinder converge within the cylinder head 22L, while an exhaust port 228 communicating with the first exhaust valve EXl of the fourth cylinder and an exhaust port 230 communicating with the first exhaust valve EXl of the sixth cylinder converge within the cylinder head 22R.

[0105] The ignition sequence of the engine 10 according to this embodiment is set at l → 8 → 7 -> 3 → 6 → 5 → 4 → 2. As noted above, in a V-type eight cylinder engine having this ignition sequence, the amount of residual gas in the first cylinder is likely to increase due to a blowdown effect from the seventh cylinder, the amount of residual gas in the third cylinder is likely to increase due to a blowdown effect from the fifth cylinder, the amount of residual gas in the sixth cylinder is likely to increase due to a blowdown effect from the fourth cylinder, and the amount of residual gas in the second cylinder is likely to increase due to a blowdown effect from the eighth cylinder.

[0106] Moreover, in the engine 10 according to this embodiment, the amount of residual gas in the third cylinder and sixth cylinder is particularly likely to increase for the following reason. [0107] As described above, in the engine 10 according to this embodiment, the exhaust port 224 communicating with the first exhaust valve EXl of the third cylinder and the exhaust port 226 communicating with the first exhaust valve EXl of the fifth cylinder converge within the cylinder head 22L. Hence, blowdown gas discharged from the first exhaust valve EXl of the fifth cylinder is more likely to circulate to the third cylinder, and therefore the amount of residual gas in the third cylinder is particularly likely to increase.

[0108] Similarly in the right-hand cylinder row 12R, the exhaust port 228 communicating with the first exhaust valve EXl of the fourth cylinder and the exhaust port 230 communicating with the first exhaust valve EXl of the sixth cylinder converge within the cylinder head 22R. Hence, blowdown gas discharged from the first exhaust valve EXl of the fourth cylinder is more likely to circulate to the sixth cylinder, and therefore the amount of residual gas in the sixth cylinder is particularly likely to increase.

[0109] Hence, in the engine 10 according to this embodiment, the residual gas amount of the first, third, second and sixth cylinders is more likely to increase than that of the other cylinders, and the residual gas amount of the third and sixth cylinders is particularly likely to increase. The third and sixth cylinders serve as examples of

"cylinders exhibiting a large residual gas effect", the first and second cylinders serve as examples of "cylinders exhibiting an intermediate residual gas effect", and the fourth, fifth, seventh and eighth cylinders serve as examples of "cylinders exhibiting a small residual gas effect".

[0110] The engine 10 according to this embodiment is a lean burn engine, and is constituted to be capable of switching between a stoichiometric combustion mode, in which fuel is burned in the vicinity of the stoichiometric air-fuel ratio, and a lean combustion mode, in which the fuel is burned at a leaner air-fuel ratio than the stoichiometric air-fuel ratio.

[0111] FIG. 10 is a block diagram showing the system constitution of the fourth embodiment of the present invention. As shown in the drawing, the system of this embodiment includes a crank angle sensor 232 for detecting a rotation angle of a crankshaft (output shaft) of the engine 10, an accelerator position sensor 234 for detecting an accelerator pedal position (accelerator opening) of a vehicle installed with the engine 10, an air flow meter 236 for detecting an intake air amount of the engine 10, a supercharging pressure sensor 238 for detecting a supercharging pressure (an intake pipe pressure on the downstream side of the compressor 14b of the turbocharger 14L, 14R), a second exhaust valve lift varying mechanism 240, an exhaust valve phase varying mechanism (exhaust VVT mechanism) 242, a first exhaust valve stopping mechanism 244, and an intake valve phase varying mechanism (intake WT mechanism) 246. These sensors and actuators are electrically connected to an ECU (Electronic Control Unit) 250.

[0112] The second exhaust valve lift varying mechanism 240 functions to vary the lift amount of the second exhaust valve EX2 continuously in each of the cylinders exhibiting a large residual gas effect, the cylinders exhibiting an intermediate residual gas effect, and the cylinders exhibiting a small residual gas effect. The second exhaust valve lift varying mechanism 240 according to this embodiment varies the lift amount and opening timing (operation angle) of the second exhaust valve EX2 while maintaining the closing timing of the second exhaust valve EX2. Further, the second exhaust valve lift varying mechanism 240 is capable of reducing the lift amount of the second exhaust valve EX2 of each cylinder to zero. In other words, by having the second exhaust valve lift varying mechanism 240 reduce the lift amount of the second exhaust valve EX2 to zero, the second exhaust valve EX2 can be stopped in a closed state.

[0113] The exhaust valve phase varying mechanism 242 is capable of retarding and advancing the valve timing of the first exhaust valve EXl and the second exhaust valve EX2 continuously by continuously varying the phase of a camshaft for driving the first exhaust valve EXl and second exhaust valve EX2. In the exhaust valve phase varying mechanism 242, a most advanced state serves as an initial state, and a retardation amount from the most advanced state serves as a control parameter.

[0114] The first exhaust valve stopping mechanism 244 has a function for stopping an operation of the first exhaust valve EXl in a closed state. The intake valve phase varying mechanism 246 is a substantially identical mechanism to the exhaust valve phase varying mechanism 242, which is capable of retarding and advancing the valve timing of an intake valve (not shown) of the engine 10 continuously by continuously varying the phase of a camshaft for driving the intake valve. In the intake valve phase varying mechanism 246, a most retarded state serves as an initial state, and an advancement amount from the most retarded state serves as a control parameter.

[0115] Note that the second exhaust valve lift varying mechanism 240, the exhaust valve phase varying mechanism 242, the first exhaust valve stopping mechanism 244, and the intake valve phase varying mechanism 246 all have conventional structures, and therefore description thereof has been omitted from this specification. The ECU 250 controls the opening characteristics of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 in the following manner by controlling the condition of the second exhaust valve lift varying mechanism 240, the exhaust valve phase varying mechanism 242, the first exhaust valve stopping mechanism 244, and the intake valve phase varying mechanism 246 in accordance with the operating condition of the engine 10.

[0116] FIG. 11 is a valve lift diagram of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 during a cold start (a catalyst warm-up region). As shown in FIG. 11, during a cold start, the lift of the second exhaust valve EX2 is set to be large, and the first exhaust valve EXl is stopped in a closed state. Further, by retarding the valve timing of the intake valve, a valve overlap period in which an open period of the second exhaust valve EX2 and an open period of the intake valve overlap is substantially eliminated.

[0117] According to the opening characteristic shown in FIG. 11, burned gas in the respective cylinders can be discharged in its entirety to the second exhaust passages 18L, 18R through the second exhaust valves EX2. In other words, all of the exhaust gas can be caused to flow into the catalysts 2OL, 2OR without passing through the turbines 14a of the turbochargers 14L, 14R. Hence, during a cold start in which the temperature of the catalysts 2OL, 2OR is to be raised as quickly as possible, an exhaust gas temperature reduction in the turbine 14a can be avoided, and therefore high-temperature exhaust gas can be caused to flow into the catalysts 2OL, 2OR. As a result, the catalysts 2OL, 2OR can be warmed up to an active temperature at an early stage following start-up, whereby emissions generated during start-up can be reduced sufficiently. [0118] In this embodiment, as shown in FIG. 9, exhaust ports communicating with the second exhaust valves EX2 of two cylinders converge within the cylinder heads 22L, 22R in each of the . following cylinder groups: the first cylinder and third cylinder; the fifth cylinder and seventh cylinder; the second cylinder and fourth cylinder; and the sixth cylinder and eighth cylinder. Hence, the length and surface area of a flow passage for the exhaust gas that is discharged from the second exhaust valves EX2 can be reduced, and therefore the amount of thermal energy lost from the exhaust gas flowing to the catalysts 2OL, 2OR from the second exhaust valves EX2 can be reduced to a minimum, making it possible to maintain the temperature of the exhaust gas flowing into the catalysts 2OL, 2OR at a maximum. As a result, the catalysts 2OL, 2OR can be warmed up even earlier and the temperature thereof can be maintained, enabling a further improvement in the emissions performance.

[0119] FIG. 12 is a valve lift diagram of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 in a high output region of the stoichiometric combustion mode. In this region, both the first exhaust valve EXl and the second exhaust valve EX2 open. In this case, the second exhaust valve EX2 opens in the latter half of an exhaust stroke at a smaller lift amount and for a shorter open period than the first exhaust valve EXl. In other words, the second exhaust valve EX2 opens after the first exhaust valve EXl and closes after the first exhaust valve EXl has closed (after top dead center). The valve timing of the intake valve is advanced such that the intake valve opens before top dead center. The second exhaust valve EX2 and the intake valve have a sufficient valve overlap period during which the open periods thereof overlap. The first exhaust valve EXl and the intake valve, on the other hand, have substantially no valve overlap period during which the open periods thereof overlap.

[0120] According to the opening characteristic shown in FIG. 12, only the first exhaust valve EXl is open during the first half of the exhaust stroke, and therefore high-energy exhaust gas can be supplied sufficiently to the turbines 14a of the turbochargers 14L, 14R. As a result, the turbochargers 14L, 14R can be activated efficiently such that a high supercharging pressure is obtained. [0121] In this embodiment, as shown in FIG. 9, exhaust ports communicating with the first exhaust valves EXl of two cylinders converge within the cylinder heads 22L, 22R in each of the following cylinder groups: the third cylinder and fifth cylinder; and the fourth cylinder and sixth cylinder. Hence, the length and surface area of a flow passage for the exhaust gas that is discharged from the first exhaust valves EXl can be reduced, and therefore the amount of thermal energy lost from the exhaust gas flowing to the turbines 14a of the turbochargers 14L, 14R from the first exhaust valves EXl can be reduced to a minimum, making it possible to maintain the energy of the exhaust gas flowing into the turbines 14a at a maximum. As a result, more energy can be recovered by the turbines 14a such that the turbochargers 14L, 14R can be activated with even greater efficiency.

[0122] Meanwhile, in the latter half of the exhaust stroke, the second exhaust valve EX2 opens, and therefore the amount of residual gas can be reduced greatly. More specifically, the second exhaust valve EX2 side does not pass through the turbine 14a, and therefore the back pressure is low and burned gas is discharged from the cylinder easily. Further, due to supercharging, the intake pipe pressure rises above the back pressure on the second exhaust valve EX2 side. Therefore, during the valve overlap state in which the second exhaust valve EX2 and the intake valve are both open, the burned gas in the cylinder is chased out (scavenged) by inflowing high-pressure fresh air from the intake valve, and thus the burned gas can be discharged to the second exhaust passage 18L, 18R efficiently through the second exhaust valve EX2. This action will be referred to hereafter as a "scavenging action". Further, substantially no valve overlap exists between the first exhaust valve EXl and the intake valve, and therefore backflow of the exhaust gas into the cylinder or the intake port from the high-back pressure first exhaust valve EXl can be prevented reliably. Hence, according to the opening characteristic shown in FIG. 12, the residual gas amount in the cylinder can be reduced greatly, and the air amount can be increased correspondingly, leading to an increase in output.

[0123] Also according to the opening characteristic shown in FIG. 12, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect (the first and second cylinders) is made greater than that of the cylinders exhibiting a small residual gas effect (the fifth, seventh, fourth and eighth cylinders), and the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) is made greater than that of the cylinders exhibiting an intermediate residual gas effect. Hence, the cylinders exhibiting an intermediate residual gas effect exhibit a greater scavenging action than the cylinders exhibiting a small residual gas effect, and the cylinders exhibiting a large residual gas effect exhibit a greater scavenging action than the cylinders exhibiting an intermediate residual gas effect. In other words, a steadily greater scavenging action is obtained as the likelihood of an increase in the amount of residual gas in the cylinder rises, and therefore the residual gas amount can be reduced sufficiently. Hence, according to the opening characteristic shown in FIG. 12, the residual gas amount can also be reduced sufficiently in the cylinders exhibiting a large residual gas effect and the cylinders exhibiting an intermediate residual gas effect, and therefore problems such as combustion deterioration and misfiring can be prevented reliably.

[0124] Also according to the opening characteristic shown in FIG. 12, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a small residual gas effect and the cylinders exhibiting an intermediate residual gas effect can be reliably prevented from increasing more than necessary. As a result, a situation in which the scavenging action is exhibited excessively such that fresh air blows through the second exhaust passages 18L, 18R can be avoided reliably.

[0125] FIG. 15 is a valve lift diagram of the intake valve, the first exhaust valve EXl, and the second exhaust valve EX2 in a fuel efficiency improvement region of the lean combustion mode. In this region, only the second exhaust valves EX2 of the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) are open, in contrast to the opening characteristic shown in FIG. 12, and in the cylinders exhibiting an intermediate residual gas effect and the cylinders exhibiting a small residual gas effect, the lift amount of the second exhaust valve EX2 is set at zero. In the lean combustion mode, the exhaust gas temperature typically tends to be low. In other words, exhaust energy decreases easily. Therefore, to activate the turbochargers 14L, 14R sufficiently, the amount of exhaust gas flowing to the second exhaust passages 18L, 18R must be reduced to a minimum, and the amount of exhaust gas flowing to the first exhaust passages 16L, 16R must be increased to a maximum. According to the opening characteristic shown in FIG. 13, the lift amount of the second exhaust valves EX2 in the cylinders other than the cylinders exhibiting a large residual gas effect is set at zero, and therefore the amount of exhaust gas flowing to the first exhaust passages 16L, 16R can be increased to a maximum. As a result, the turbochargers 14L, 14R can be activated sufficiently, and a sufficient supercharging pressure can be obtained. In the cylinders exhibiting a large residual gas effect, meanwhile, the second exhaust valve EX2 is open, and therefore the scavenging action can be exhibited, enabling a reduction in the amount of residual gas. Hence, in the cylinders exhibiting a large residual gas effect, where the residual gas amount is particularly likely to increase, the residual gas amount can be reduced sufficiently, and as a result, problems such as combustion deterioration and misfiring can be prevented reliably.

[0126] FIG. 14 is a view showing an operating region of the engine 10. As shown in the drawing, the operating region of the engine 10 in this embodiment is divided into four regions, namely A, B, C and D. The operating regions of A and B constitute a stoichiometric combustion region in which the engine 10 is operated in the stoichiometric combustion mode. The operating regions of C and D constitute a lean combustion region in which the engine 10 is operated in the lean combustion mode.

[0127] In the operating region A of the stoichiometric combustion region, the opening characteristic of the intake and exhaust valves is controlled to correspond to the opening characteristic shown in FIG. 12. In the operating region B of the stoichiometric combustion region, on the other hand, the opening characteristic of the intake and exhaust valves is controlled to correspond to the opening characteristic shown in FIG. 15. According to the opening characteristic shown in FIG. 15, the lift amount of the second exhaust valves EX2 in all of the cylinders is set at zero. Further, the valve timing of the exhaust valve is retarded by the exhaust valve phase varying mechanism 242 such that a valve overlap period is provided between the open period of the first exhaust valve EXl and the open period of the intake valve.

[0128] As shown in FIG. 14, the operating region B is a region including a low-rotation high-load region. The low-rotation high-load region is a region in which a so-called turbo lag (a response delay in the supercharging pressure) is most likely to occur. Hence, in this embodiment, the opening characteristic shown in FIG. 15 is set in the operating region B such that the lift amount of the second exhaust valves EX2 in all of the cylinders is set at zero. In so doing, all of the exhaust gas is caused to flow into the turbines 14a of the turbochargers 14L, 14R. As a result, the turbo lag in the low-rotation high-load region can be reduced to a minimum.

[0129] In the operating region C of the lean combustion region, the opening characteristic of the intake and exhaust valves is controlled to correspond to the opening characteristic shown in FIG. 13. Meanwhile, the operating region D of the lean combustion region is a low-load region in which the amount of exhaust gas is small and blowdown is weak, and therefore even in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) the residual gas increasing effect of blowdown from other cylinders (the fifth and fourth cylinders) is small. Hence, even in the cylinders exhibiting a large residual gas effect, there is little need to open the second exhaust valves EX2 to generate the scavenging effect. Furthermore, in the low-load operating region D, the supercharging pressure, or in other words the intake pipe pressure, is also unlikely to increase, and therefore even when the second exhaust valve EX2 is open, the obtained scavenging action is small. Hence, in this embodiment, the opening characteristic of the intake and exhaust valves is set to correspond to the opening characteristic shown in FIG. 15 in the operating region D. In other words, the lift amount of the second exhaust valves EX2 in all of the cylinders is set at zero.

[0130] FIG. 16 is a flowchart showing a routine that is executed in this embodiment by the ECU 250 to realize the functions described above. This routine is executed repeatedly at either predetermined time intervals or predetermined crank angle intervals. According to the routine shown in FIG. 16, first, an accelerator opening and an engine rotation speed are calculated on the basis of detection signals from the accelerator position sensor 234 and the crank angle sensor 232 (step 100). Next, an output required of the engine 10 is determined on the basis of the accelerator opening and engine rotation speed (step 102). Next, a determination is made as to whether or not an operating point of the engine 10 is in the region A of FIG. 14 on the basis of the required output, accelerator opening, and engine rotation speed (step 104).

[0131] When the operating point of the engine 10 is determined to be in the region A in the step 104, a basic lift amount of the second exhaust valve EX2, a retardation amount of the exhaust valve phase varying mechanism 242, and an advancement amount of the intake valve phase varying mechanism 246 are calculated respectively on the basis of a stoichiometric combustion mode map, which is stored in the ECU 250 in advance (step 106). Next, a lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect (the first and second cylinders) and a lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) are calculated respectively on the basis of a predetermined map or an equation (step 108). In the step 108, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect is calculated to a larger value than the basic lift amount of the step 106, and the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect is calculated to a larger value than the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect.

[0132] Following the processing of the step 108, drive control of the second exhaust valve lift varying mechanism 240 is executed. More specifically, the lift amount of the second exhaust valves EX2 of the cylinders exhibiting a small residual gas effect (the fifth, seventh, fourth and eighth cylinders) is controlled to the basic lift amount calculated in the step 106, and the respective lift amounts of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect and cylinders exhibiting a large residual gas effect are controlled to the predetermined lift amounts calculated in the step 108 (step 110). Thus, in the operating region A, an opening characteristic such as that shown in FIG. 12^U is realized.

[0133] When it is determined in the step 104 that the operating point of the engine 10 is not in the region A, on the other hand, a determination is made as to whether or not the operating point of the engine 10 is in the region C of FIG. 14 (step 112). [0134] When the operating point of the engine 10 is determined to be in the region C in the step 112, the basic lift amount of the second exhaust valve EX2, the retardation amount of the exhaust valve phase varying mechanism 242, and the advancement amount of the intake valve phase varying mechanism 246 are calculated respectively on the basis of a lean combustion mode map, which is stored in the ECU 250 in advance (step 114). In this embodiment, the basic lift amount of the second exhaust valve EX2 calculated in the step 114 is set at zero.

[0135] Following the processing of the step 114, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect (the first and second cylinders) and the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) are calculated respectively on the basis of a predetermined map or an equation (step 116). In the step 116 of this embodiment, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect is calculated as zero, and the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect is calculated to a larger value than zero.

[0136] Following the processing of the step 116, drive control of the second exhaust valve lift varying mechanism 240 is executed. More specifically, the lift amount of the second exhaust valves EX2 of the cylinders exhibiting a small residual gas effect (the fifth, seventh, fourth and eighth cylinders) is controlled to the basic lift amount calculated in the step 114, i.e. zero, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect is controlled to the lift amount calculated in the step 116, i.e. zero, and the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect is controlled to the predetermined lift amount calculated in the step 116 (step 110). Thus, in the operating region C, an opening characteristic such as that shown in FIG. 13 is realized.

[0137] When it is determined in the step 112 that the operating point of the engine 10 is not in the region C, on the other hand, this indicates that the operating point of the engine 10 is in either the region B or the region D of FIG. 14. In this case, an opening characteristic such as that shown in FIG. 15 is realized by executing drive control of the second exhaust valve lift varying mechanism 240 such that the lift amount of the second exhaust valves EX2 in all of the cylinders reaches zero (step 118).

[0138] In the fourth embodiment described above, a variable valve device of the intake and exhaust valves is constituted by the second exhaust valve lift varying mechanism 240, the exhaust valve phase varying mechanism 242, the first exhaust valve stopping mechanism 244, and the intake valve phase varying mechanism 246. However, the variable valve device of the present invention is not limited to this constitution, and may be replaced by another device having an arbitrary constitution, which is capable of performing similar functions. For example, a variable valve device that can control the opening/closing timing of a valve arbitrarily by driving a camshaft to rotate using an electric servo motor, an electromagnetically driven or hydraulically driven variable valve device, or similar may be used.

[0139] Further, in this embodiment, the magnitude of the scavenging action is controlled by varying the lift amount of the second exhaust valve EX2, but the magnitude of the scavenging action may be controlled by varying the valve overlap period in which the open period of the second exhaust valve EX2 and the open period of the intake valve overlap (to be referred to hereafter as "bypass-side valve overlap period"). Hence, in the present invention, the bypass-side valve overlap period of the cylinders exhibiting a large residual gas effect and the cylinders exhibiting an intermediate residual gas effect may be controlled to be longer than the bypass-side valve overlap period of the cylinders exhibiting a small residual gas effect instead of varying the lift amount of the second exhaust valve EX2. In this case, either the closed period of the second exhaust valve EX2 may be controlled separately in the cylinders exhibiting a large residual gas effect, the cylinders exhibiting an intermediate residual gas effect, and the cylinders exhibiting a small residual gas effect, or the open period of the intake valve may be controlled separately in the cylinders exhibiting a large residual gas effect, the cylinders exhibiting an intermediate residual gas effect, and the cylinders exhibiting a small residual gas effect. [0140] Further, in this embodiment, a V-type eight cylinder engine having an ignition sequence of l → 8 → 7 → 3 → 6 → 5 → 4 → 2 was described, but a V-type eight cylinder engine has various ignition sequences, and is not limited to this sequence. Even when the ignition sequence is different, it is possible to identify the cylinders exhibiting a large residual gas effect, the cylinders exhibiting an intermediate residual gas effect, and the cylinders exhibiting a small residual gas effect using similar principles to those described with reference to FIGS. 19 and 9.

[0141] Further, the number of cylinders and the cylinder arrangement of the internal combustion engine according to the present invention are not limited to those of a V-type eight cylinder engine, and as long as the combustion intervals between the respective cylinders in the same cylinder row are irregular, the present invention may be applied to an engine having a different number of cylinders and a different cylinder arrangement.

[0142] Further, in the fourth embodiment, the "opening characteristic control means" of the first invention described above are realized by having the ECU 250 execute the processing of the routine shown in FIG. 16. [0143] Next, referring to FIG. 17, a fifth embodiment of the present invention will be described. The following description centers on differences with the fourth embodiment, and description of identical items has been simplified or omitted. This embodiment can be realized by having the ECU 250 execute a routine shown in FIG. 17, to be described below, using a similar hardware constitution to that of the fourth embodiment. [0144] As a feature of the fifth embodiment, in the region A of FIG. 14, the scavenging action is exhibited by opening the second exhaust valve EX2 of each cylinder, as described above, whereas in the regions B and D, the lift amount of the second exhaust valve EX2 of each cylinder is set at zero such that the scavenging action is not exhibited. Hence, when the operating point of the engine 10 shifts from the region A to the region B or the region D, a stop command relating to the second exhaust valves EX2 is issued, thereby halting the scavenging action, and as a result, the residual gas amount in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) is particularly likely to increase. Moreover, by stopping the second exhaust valves EX2, an exhaust gas flow rate to the first exhaust valve EXl side increases rapidly such that the back pressure on the first exhaust valve EXl side and the likelihood of a large increase in the amplitude of the exhaust pulse increase transiently. Therefore, during a transitional operation in which the operating point of the engine 10 shifts from the region A to the region B or the region D, the residual gas amount in the cylinders exhibiting a large residual gas effect is particularly likely to increase in comparison with a steady operation, and as a result, drivability becomes more likely to be adversely affected by combustion deterioration and misfiring.

[0145] In this embodiment, to improve these points, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect is not set at zero during a shift from the region A to the region B or D until the transient exhaust pulse having a large amplitude attenuates and stabilizes, thereby maintaining the scavenging action. In so doing, a transient increase in the residual gas amount of the cylinders exhibiting a large residual gas effect can be prevented.

[0146] Further, while switching the respective lift amounts of the second exhaust valves EX2 in the cylinders other than the cylinders exhibiting a large residual gas effect to zero, the lift amounts are typically switched successively in accordance with the ignition sequence. In this embodiment, however, control is performed to avoid switching the cylinders (the fifth and fourth cylinders) that affect the exhaust pulse of the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) first. When the lift amount of the second exhaust valve EX2 in the fifth cylinder or the fourth cylinder is switched to zero first, the exhaust pulse of the exhaust port 226 or 228 communicating with the first exhaust valve EXl of the fifth cylinder or fourth cylinder becomes particularly large, and this large exhaust pulse is transmitted to the third cylinder or the sixth cylinder via the exhaust port 224 or 230. As a result, the residual gas amount in the third cylinder or sixth cylinder, i.e. the cylinders exhibiting a large residual gas effect, is likely to increase > excessively. In this embodiment, on the other hand, switching the lift amounts of the second exhaust valves EX2 in the fifth and fourth cylinders to zero first is avoided, and therefore this situation can be avoided. Thus, the residual gas amount of the third cylinder or the sixth cylinder, i.e. the cylinders exhibiting a large residual gas effect, can be reduced reliably.

[0147] FIG. 17 is a flowchart of a routine executed by the ECU 250 in this embodiment to realize the functions described above. In FIG. 17, identical numerals have been allocated to identical steps to the steps shown in FIG. 16, and description thereof has been omitted or simplified. According to the routine shown in FIG. 17, first, the accelerator opening and the engine rotation speed are calculated (step 100), whereupon the required output of the engine 10 is determined on the basis of the accelerator opening and the engine rotation speed (step 102).

[0148] Next, on the basis of the required output, the accelerator opening, and the engine rotation speed, the basic lift amount of the second exhaust valve EX2, the retardation amount of the exhaust valve phase varying mechanism 242, and the advancement amount of the intake valve phase varying mechanism 246 are respectively calculated by referring to a predetermined map stored in the ECU 250 in advance (step 120). Next, on the basis of the calculation results obtained in the step 120, a determination is made as to whether or not a request to shift to the region B, in which the basic lift amount of the second exhaust valve EX2 is zero, has been issued (step 122).

[0149] When it is determined in the step 122 that a request to shift to the region B has been issued, a determination is made as to whether or not an operation cycle of the engine 10 has advanced by a predetermined cycle number a following issuance of the shift request (step 124). The predetermined cycle number α is a preset value serving as a cycle number required for a large-amplitude transient exhaust pulse to attenuate during a shift from the region A to the region B. When it is determined in the step 124 that the predetermined cycle number α has not elapsed, the respective lift amounts of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) are calculated on the basis of a predetermined map (step 126). The lift amounts of the second exhaust valves EX2 calculated in the step 126 are set such that a sufficient scavenging action is obtained to avoid a transient residual gas increase in the cylinders exhibiting a large residual gas effect. [0150] Following the processing of the step 126, a determination is made on the basis of a signal from the crank angle sensor 232 or the like as to whether or not a current timing corresponds to a timing for switching the lift amount of the second exhaust valve EX2 in the fifth cylinder or fourth cylinder (step 128). When it is determined in the step 128 that the switching timing of the fifth cylinder or fourth cylinder has arrived, the routine enters standby to avoid switching the lift amount of the second exhaust valve EX2 in the fifth cylinder or fourth cylinder to zero first. When it is determined in the step 128 that the switching timings of the fifth cylinder and fourth cylinder have not arrived, on the other hand, drive control of the second exhaust valve lift varying mechanism 240 is executed to switch the lift amounts of the second exhaust valves EX2 in the respective cylinders in succession (step 130). In other words, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect is controlled to the predetermined lift amount calculated in the step 126, while the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect and cylinders exhibiting a small residual gas effect is controlled to zero. [0151] When the predetermined cycle number α has elapsed following issuance of the request to shift from the region A to the region B, or in other words when it can be determined that the large-amplitude transient exhaust pulse has attenuated, the determination of the step 124 becomes affirmative, and therefore the lift amounts of the second exhaust valves EX2 in all of the cylinders are controlled to zero (step 132). In other words, the lift amounts of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect are also switched to zero.

[0152] When it is determined in the step 122 that the request to shift to the region B has not been issued, on the other hand, a determination is made as to whether or not a request to shift to the region D, in which the basic lift amount of the second exhaust valve EX2 is zero, has been issued (step 134). When it is determined in the step 134 that a request to shift to the region D has been issued, a determination is made as to whether or not the operation cycle of the engine 10 has advanced by a predetermined cycle number β following issuance of the shift request (step 136). The predetermined cycle number β is a preset value serving as a cycle number required for a large-amplitude transient exhaust pulse to attenuate during a shift from the region A to the region D. When it is determined in the step 136 that the predetermined cycle number β has not elapsed, a determination is made as to whether or not the supercharging pressure (intake pipe pressure) detected by the supercharging pressure sensor 238 is equal to or greater than a predetermined pressure (step 138). The predetermined pressure takes a preset value serving as a lower limit supercharging pressure at which the scavenging action is exhibited effectively. The region D is a low-load region in which the turbochargers 14L, 14R are not activated easily, and therefore the supercharging pressure occasionally falls to or below the predetermined pressure. In this case, the scavenging action is not exhibited effectively even when the second exhaust valves EX2 of the cylinders exhibiting a large residual gas effect are open. Hence, when it cannot be determined in the step 138 that the supercharging pressure is equal to or greater than the predetermined pressure, the lift amount of the second exhaust valve EX2 in all of the cylinders is immediately controlled to zero (step 132). [0153] On the other hand, when it is determined in the step 138 that the supercharging pressure is equal to or greater than the predetermined pressure, the lift amounts of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect (the third and sixth cylinders) are calculated on the basis of a predetermined map (step 140). The lift amounts of the second exhaust valves EX2 calculated in the step 140 are set such that the scavenging action required to avoid a transient residual gas increase in the cylinders exhibiting a large residual gas effect is obtained.

[0154] Following the processing of the step 140, a determination is made on the basis of a signal from the crank angle sensor 232 or the like as to whether or not the current timing corresponds to the timing for switching the lift amount of the second exhaust valve EX2 in the fifth cylinder or fourth cylinder (step 128). When it is determined in the step 128 that the switching timing of the fifth cylinder or fourth cylinder has arrived, the routine enters standby to avoid switching the lift amount of the second exhaust valve EX2 in the fifth cylinder or fourth cylinder to zero first. When it is determined in the step 128 that the switching timings of the fifth cylinder and fourth cylinder have not arrived, on the other hand, drive control of the second exhaust valve lift varying mechanism 240 is executed to switch the lift amounts of the second exhaust valves EX2 in the respective cylinders in succession (step 130). In other words, the lift amount of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect is controlled to the predetermined lift amount calculated in the step 140, while the lift amount of the second exhaust valves EX2 in the cylinders exhibiting an intermediate residual gas effect and cylinders exhibiting a small residual gas effect is controlled to zero.

[0155] When the predetermined cycle number β has elapsed following issuance of the request to shift from the region A to the region D, or in other words when it can be determined that the large-amplitude transient exhaust pulse has attenuated, the determination of the step 136 becomes affirmative, and therefore the lift amounts of the second exhaust valves EX2 in all of the cylinders are controlled to zero (step 132). In other words, the lift amounts of the second exhaust valves EX2 in the cylinders exhibiting a large residual gas effect are also switched to zero.

[0156] According to the processing of the routine shown in FIG. 17 and described above, when the operating point of the engine 10 shifts from the region A to the region B or the region D, a transient increase in the residual gas amount of the cylinders exhibiting a large residual gas effect can be prevented reliably. As a result, adverse effects such as combustion deterioration and misfiring in the cylinders exhibiting a large residual gas effect can be avoided such that favorable drivability is obtained.

[0157] Note that in the fifth embodiment described above, the ECU 250 serves as an example of "second exhaust valve stopping means" when using the second exhaust valve lift varying mechanism 240 to set the lift amount of the second exhaust valve EX2 to zero, and serves as an example of "lift amount switching sequence control means" when executing the processing of the routine shown in FIG. 17.

[0158] While example embodiments of the invention have been described above, it is to be understood that the invention is not limited to the particulars of the described embodiments, and may be embodied with various changes, modifications or improvements, which may occur to those skilled in the art, without departing from the scope of the invention.

Claims

1. An exhaust device for an internal combustion engine, characterized by comprising: a turbocharger; and a plurality of cylinders, each including a first exhaust passage that passes through the turbocharger and a first exhaust valve provided in the first exhaust passage, wherein the first exhaust passages of a pair of cylinders, from among the plurality of cylinders, converge initially, and then converge with another first exhaust passage between a convergence point of the first exhaust passages and the turbocharger.

2. The exhaust device for an internal combustion engine according to claim 1, wherein: the internal combustion engine performs combustion at irregular intervals; each of the plurality of cylinders comprises a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage; and the first exhaust passages of adjacent cylinders, from among the plurality of cylinders, converge initially, and then converge with another first exhaust passage between a convergence point of the first exhaust passages and the turbocharger.

3. The exhaust device for an internal combustion engine according to claim 2, wherein: the internal combustion engine is a V-type eight cylinder engine; and the first exhaust passage of a certain cylinder and the first exhaust passage of another cylinder having a combustion stroke that is removed from the combustion stroke of the certain cylinder by a crank angle of 180° do not converge initially upstream of the turbocharger.

4. The exhaust device for an internal combustion engine according to claim 3, wherein: the internal combustion engine performs combustion at irregular intervals; each of the plurality of cylinders comprises a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage; the plurality of cylinders is constituted by a first cylinder, a second cylinder, a third cylinder, a fourth cylinder, a fifth cylinder, a sixth cylinder, a seventh cylinder, and an eighth cylinder; the V-type eight cylinder engine comprises, within a cylinder head, a first bank including the first cylinder, the third cylinder, the fifth cylinder and the seventh cylinder, and a second bank including the second cylinder, the fourth cylinder, the sixth cylinder and the eighth cylinder; and the first exhaust passages of the first cylinder and the third cylinder, the first exhaust passages of the fifth cylinder and the seventh cylinder, the first exhaust passages of the second cylinder and the fourth cylinder, and the first exhaust passages of the sixth cylinder and the eighth cylinder respectively converge initially within the cylinder head.

5. The exhaust device for an internal combustion engine according to claim 1, wherein: the internal combustion engine performs combustion at irregular intervals; each of the plurality of cylinders comprises a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage; the plurality of cylinders is constituted by a first cylinder, a second cylinder, a third cylinder, a fourth cylinder, a fifth cylinder, a sixth cylinder, a seventh cylinder, and an eighth cylinder; the internal combustion engine is a V-type eight cylinder engine comprising, within a cylinder head, a first bank including the first cylinder, the third cylinder, the fifth cylinder and the seventh cylinder, and a second bank including the second cylinder, the fourth cylinder, the sixth cylinder and the eighth cylinder; and the first exhaust passages of the third cylinder and the fifth cylinder are led to the turbocharger without initially converging, and the first exhaust passages of the fourth cylinder and the sixth cylinder are led to the turbocharger without initially converging.

6. The exhaust device for an internal combustion engine according to claim 1, wherein: the internal combustion engine performs combustion at irregular intervals; each of the plurality of cylinders comprises a second exhaust passage passing downstream of the turbocharger and a second exhaust valve provided in the second exhaust passage; the plurality of cylinders is constituted by a first cylinder, a second cylinder, a third cylinder, a fourth cylinder, a fifth cylinder, a sixth cylinder, a seventh cylinder, and an eighth cylinder; the internal combustion engine is a V-type eight cylinder engine comprising, within a cylinder head, a first bank including the first cylinder, the third cylinder, the fifth cylinder and the seventh cylinder, and a second bank including the second cylinder, the fourth cylinder, the sixth cylinder and the eighth cylinder; and the first exhaust passages of the second cylinder and the fourth cylinder are led to the turbocharger after initially converging, and the first exhaust passages of the third cylinder and the seventh cylinder are led to the turbocharger after initially converging.

7. The exhaust device for an internal combustion engine according to any one of claims 2 to 6, in a predetermined operating condition, a lift amount of the second exhaust valve or a bypass-side valve overlap period, during which an open period of the second exhaust valve and an open period of an intake valve overlap, of a certain cylinder from among the plurality of cylinders is greater than a lift amount of the second exhaust valve or a bypass-side valve overlap period, during which an open period of the second exhaust valve and an open period of an intake valve overlap, of the other cylinder of the plurality of cylinders.

8. The exhaust device for an internal combustion engine according to any one of claims 2 to 7, wherein when the first exhaust valve and the second exhaust valve are both driven, an opening timing of the second exhaust valve is later than an opening timing of the first exhaust valve, and a closing timing of the second exhaust valve is later than a closing timing of the first exhaust valve.

9. The exhaust device for an internal combustion engine according to claim 1, wherein the pair of cylinders is set on the basis of timings at which the plurality of cylinders discharge exhaust gas via the first exhausts passages.

10. A control device for an internal combustion engine in which combustion intervals of respective cylinders within an identical cylinder row are irregular such that the magnitude of a residual gas effect, according to which a residual gas amount becomes more likely to increase due to an exhaust pulse effect from another cylinder, differs among the cylinders, characterized by comprising: a turbocharger; a first exhaust passage that passes through a turbine inlet of the turbocharger; a first exhaust valve for opening and closing an exhaust port that communicates with the first exhaust passage; a second exhaust passage that does not pass through the turbine inlet; a second exhaust valve for opening and closing an exhaust port that communicates with the second exhaust passage; a variable valve device that is capable of varying a lift amount of the second exhaust valve or a bypass-side valve overlap period, during which an open period of the second exhaust valve and an open period of an intake valve overlap, separately in cylinders exhibiting a large residual gas effect, in which the residual gas effect is large, and other cylinders; and opening characteristic control means for controlling the variable valve device such that in a predetermined operating condition, the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting a large residual gas effect is greater than the lift amount of the second exhaust valve or the bypass-side valve overlap period of the other cylinders.

11. The control device for an internal combustion engine according to claim 10, wherein: the respective cylinders of the internal combustion engine are divided into the cylinders exhibiting a large residual gas effect, cylinders exhibiting an intermediate residual gas effect, in which the residual gas effect is smaller than that of the cylinders exhibiting a large residual gas effect, and cylinders exhibiting a small residual gas effect,

, in which the residual gas effect is smaller than that of the cylinders exhibiting an intermediate residual gas effect; the variable valve device is capable of varying the lift amount of the second exhaust valve or the bypass-side valve overlap period separately in the cylinders exhibiting a large residual gas effect, the cylinders exhibiting an intermediate residual gas effect, and the cylinders exhibiting a small residual gas effect; and the opening characteristic control means controls the variable valve device such that the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting a large residual gas effect is greater than the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting an intermediate residual gas effect, and such that the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting an intermediate residual gas effect is equal to or greater than the lift amount of the second exhaust valve or the bypass-side valve overlap period of the cylinders exhibiting a small residual gas effect.

12. The control device for an internal combustion engine according to claim 10 or 11, further comprising: second exhaust valve stopping means capable of setting the lift amount of the second exhaust valve to zero separately in the cylinders exhibiting a large residual gas effect and the other cylinders; and lift amount switching sequence control means which, when switching the lift amount of the second exhaust valve in each cylinder to zero, switches the life amount of the second exhaust valve in the other cylinders to zero first, and switches the lift amount of the second exhaust valve in the cylinders exhibiting a large residual gas effect to zero thereafter.

13. The control device for an internal combustion engine according to claim 12, wherein the lift amount switching sequence control means switches the lift amount of the second exhaust valve in a certain cylinder, from among the other cylinders, that does not have an exhaust pulse effect On the cylinders exhibiting a large residual gas effect to zero first.

14. The control device for an internal combustion engine according to any one of claims 10 to 13, wherein in the cylinders exhibiting a large residual gas effect and cylinders that have an exhaust pulse effect on the cylinders exhibiting a large residual gas effect, the respective first exhaust valves thereof or the respective exhaust ports that pass through the second exhaust valves thereof converge within a cylinder head.

15. The control device for an internal combustion engine according to any one of claims 10 to 14, wherein: positions of the first exhaust valve and the second exhaust valve are reversed in adjacent cylinders of an identical cylinder row; either the first exhaust valves or the second exhaust valves of two adjacent cylinders are positioned adjacent to each other; and exhaust ports passing respectively through the two second exhaust valves positioned adjacent to each other converge within the cylinder head, and exhaust ports passing respectively through the two first exhaust valves positioned adjacent to each other converge within the cylinder head.

16. The control device for an internal combustion engine according to any one of claims 10 to 15, wherein, when the first exhaust valve and the second exhaust valve are both driven, an opening timing of the second exhaust valve is later than an opening timing of the first exhaust valve, and a closing timing of the second exhaust valve is later than a closing timing of the first exhaust valve.