US20050241597A1

US20050241597A1 - Air and fuel supply system for a combustion engine

Info

Publication number: US20050241597A1
Application number: US10/992,125
Authority: US
Inventors: James Weber; Scott Leman; Gerald Coleman; Kevin Duffy; Eric Fluga; Jonathan Kilkenny; Wade Robel; Sean Cornell
Original assignee: Caterpillar Inc
Current assignee: Caterpillar Inc
Priority date: 2002-05-14
Filing date: 2004-11-19
Publication date: 2005-11-03

Abstract

An internal combustion engine has an engine block defining at least one cylinder, a head, a piston, an air intake valve, and a valve actuator. The valve actuator has an actuator housing defining a tank and a bore, and a piston adapted to engages the air intake valve. The valve actuator further has a control valve disposed between the tank and the bore in the actuator housing. The control valve selectively moves between a first position where fluid is allowed to flow between the tank and the bore and a second position where the fluid is prevented from flowing to trap fluid in the bore, thereby preventing the engine valve from returning to a closed position. The internal combustion engine also has at least one turbocharger fluidly connected to the air intake port, and a fuel supply system operable to controllably inject fuel into the combustion chamber.

Description

RELATED APPLICATIONS

This application is a continuation-in-part of U.S. patent application Ser. No. 10/933,300, filed Sep. 3, 2004, which is a continuation-in-part of U.S. patent application Ser. No. 10/733,570, filed Dec. 12, 2003, which is a continuation of U.S. patent application Ser. No. 10/143,908, filed May 14, 2002, now U.S. Pat. No. 6,688,280. This application is also a continuation-in-part of U.S. patent application Ser. No. 10/733,570, filed Dec. 12, 2003, which is a continuation of U.S. patent application Ser. No. 10/143,908, filed May 14, 2002, now U.S. Pat. No. 6,688,280. This application is also a continuation-in-part of U.S. patent application Ser. No. 10/457,351, filed Jun. 10, 2003.
The entire disclosure of each of the U.S. patent applications mentioned in the preceding paragraph is incorporated herein by reference. In addition, the entire disclosure of each of U.S. Pat. No. 6,651,618 and U.S. Pat. No. 6,688,280 is incorporated herein by reference.

TECHNICAL FIELD

The present invention relates to a supply system for an internal combustion engine and, more particularly, to a fuel and air supply system for an internal combustion engine.

BACKGROUND

The operation of an internal combustion engine, such as, for example, a diesel, gasoline, or gaseous fuel driven engine such as a natural gas engine, may cause the generation of undesirable emissions. These emissions, which may include particulates and nitrous oxide (NOx), are generated when fuel is combusted in a combustion chamber of the engine. An exhaust stroke of an engine piston forces exhaust gas, which may include these emissions, from the engine. If no emission reduction measures are in place, these undesirable emissions will eventually be exhausted to the environment.
Research is currently being directed towards decreasing the amount of undesirable emissions that are exhausted to the environment during the operation of an engine and on improving engine efficiency. One such approach involves adjusting the actuation timing of the engine valves. For example, the actuation timing of the intake and exhaust valves may be modified to implement a variation on the typical diesel or Otto cycle known as the Miller cycle. In a “late intake” type Miller cycle, the intake valves of the engine are held open during a portion of the compression stroke of the piston. By holding the intake valves open during a portion of the compression stroke of the piston, the compression ratio of the engine is reduced while maintaining a high expansion ratio, which results in a temperature reduction of the fuel/air mixture within the combustion chamber. This improved thermal efficiency reduces the emission of NOx.
One system utilized to vary intake valve timing is described in U.S. Pat. No. 6,237,551 (the '551 patent) issued to Macor et al. on May 29, 2001. The '551 patent describes a hydraulic actuator that establishes a hydraulic link between a cam and an intake valve. When the link is established, the valve will be actuated according to the shape of the cam. However, when the link is broken, such as by opening a control valve, the force of a valve return spring causes the engine valve to close. Thus breaking the hydraulic link allows the engine valve to close at a different timing than would be achieved by the shape of the cam.
Although the valve actuation system of the '551 patent may provide some flexibility in the opening timing of the intake valve, it may be problematic. For example, the type of hydraulic actuator described in the '551 patent typically uses engine lubricating oil as the operating fluid. Lubricating oil may be supplied to the hydraulic actuator by a standard engine lubrication system. However, the lubricating oil may become contaminated with dirt, or debris, as the lubricating oil is circulated through the engine. Any such contamination of the lubricating oil may lead to degraded performance of the hydraulic actuator, which may translate to a reduction in the overall efficiency of the engine.
In addition, because the engine of the '551 patent can only operate by establishing the hydraulic link between the cam and the intake valve, and because the viscosity of the lubricating oil used to establish the hydraulic link may depend upon temperature, the engine of the '551 patent may not operate properly under varying environmental and operational conditions. For example, when the lubricating oil is cold, such as when the engine is starting, the hydraulic actuator may experience slow response times because of the increased viscosity of the lubricating oil. Under some environmental conditions, the engine may need to operate for a period of time to warm the lubricating oil so that the hydraulic actuator will operate as expected. The engine may experience rough running conditions or difficulty starting until the lubricating oil is warmed enough to allow the hydraulic actuator to operate properly.
Further reduction in the amount of pollutants emitted to the atmosphere and improvement in engine efficiency can be realized by combining a charged air induction system with variable valve timing. One such system is described in U.S. Pat. No. 6,273,076 (the '076 patent) issued to Beck et al. on Aug. 14, 2001. The '076 patent describes an engine having camless electro-hydraulically controlled intake valves capable of modulating a supply of air to a combustion chamber. The engine of the '076 patent also describes a combined supercharger/turbocharger assembly for directing charged air to the combustion chamber.
Although the engine of the '076 patent may reduce the amount of pollutants emitted from an engine by combining a charged induction system with variable valve timing, the engine of the '076 patent may also be problematic. In addition to the operational limitations of the engine of the '076 patent acquired because the intake valves are hydraulically controlled, the engine of the '076 patent may also be efficiency limited because the air induction system includes a supercharger. Specifically, the supercharger does not utilize the energy available in the exhaust flow from the engine of the '076 patent and requires additional driving energy from the engine to turn the compressor.
The disclosed air and fuel supply system is directed to overcoming one or more of the problems set forth above.

SUMMARY OF THE INVENTION

In one aspect, the present disclosure is directed to an internal combustion engine. The internal combustion engine includes an engine block defining at least one cylinder and a head connected with said engine block, the head including an air intake port, and an exhaust port. The internal combustion engine also includes a piston slidable in the cylinder and a combustion chamber being defined by said head, said piston, and said cylinder. The internal combustion engine further includes an air intake valve controllably movable to open and close the air intake port and a valve actuator. The valve actuator includes an actuator housing defining a tank adapted to store a supply of fluid and a bore in fluid communication with the tank. The valve actuator also includes a piston slidably disposed in the bore of the actuator housing. The piston is adapted to move between a first position and a second position where the piston engages the air intake valve. The valve actuator further includes a mechanical biasing element acting on the piston to move the piston towards the second position, and a control valve disposed between the tank and the bore in the actuator housing. The control valve selectively moves between a first position where fluid is allowed to flow between the tank and the bore, and a second position where the fluid is prevented from flowing between the bore and the tank to trap fluid in the bore. The trapped fluid prevents the piston from moving with respect to the actor housing to thereby prevent the engine valve from returning to a closed position. The internal combustion engine also includes an air supply system having at least one turbocharger fluidly connected to the air intake port, and a fuel supply system operable to controllably inject fuel into the combustion chamber.
In another aspect, the present disclosure is directed to a method of operating an internal combustion engine having at least one cylinder and a piston slidable in the cylinder. The method includes supplying pressurized air from an intake manifold to an air intake port of a combustion chamber in the cylinder. The method also includes operating a cam assembly to move an engine valve between a first position where the engine valve prevents a flow of the pressurized air between the intake manifold and the combustion chamber and a second position where the engine valve allows the flow of pressurized air during a majority portion of a compression stroke of the piston. The method further includes extending a piston from an actuator housing to engage the engine valve and directing a flow of fluid from a tank disposed within the actuator housing to a bore in the actuator housing, the bore being associated with the piston. The method additionally includes selectively preventing fluid from flowing from the bore to the tank to trap fluid in the bore and prevent the piston from moving with respect to the actuator housing, the piston engaging the engine valve to prevent the ending valve from returning to the first position.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a diagrammatic illustration of an exemplary disclosed internal combustion engine having an air supply system;
FIG. 2 is a cutaway illustration of the internal combustion engine of FIG. 1;
FIG. 3 is a cutaway view of a portion of the internal combustion engine of FIG. 1;
FIG. 4 is a cross-sectional view illustration of an exemplary disclosed variable valve closing mechanism;
FIG. 5 is a cross-sectional view illustration of another exemplary disclosed variable valve closing mechanism;
FIG. 6 is a cross-sectional view illustration of another exemplary disclosed variable valve closing mechanism;
FIG. 7 is a graph illustrating exemplary disclosed valve actuations as a function of engine crank angle;
FIG. 8 is a cross-sectional illustration of an exemplary disclosed fuel injection assembly for the internal combustion engine of FIG. 1;
FIG. 9 is a graph illustrating exemplary disclosed fuel injections as functions of engine crank angle;
FIG. 10 is a diagrammatic illustration of another exemplary disclosed internal combustion engine having an air supply system;
FIG. 11 is a diagrammatic illustration of another exemplary disclosed internal combustion engine having an air supply system; and
FIG. 12 is a diagrammatic illustration of another exemplary disclosed internal combustion engine having exhaust gas recirculation system.

DETAILED DESCRIPTION

FIG. 1 illustrates an exemplary power system 10 having internal combustion engine 12, an air supply system 14, and intake and exhaust manifolds 16 and 18 fluidly connecting internal combustion engine 12 and air supply system 14. Internal combustion engine 12 may include an engine block 20 defining a plurality of cylinders 22, the number of which depends upon the particular application. In one example internal combustion engine 12 may include six cylinders 22, however, it is contemplated that internal combustion engine 12 may include any number of cylinders 22 and that cylinders 22 may be disposed in an in-line configuration, a V-configuration, or any other suitable configuration. It should be appreciated that internal combustion engine 12 may be any type of internal combustion engine such as, for example, a diesel engine, a gasoline engine, a gaseous fuel driven engine such as a natural gas engine, or any other type of engine known in the art.
As illustrated in FIG. 2, internal combustion engine 12 may include numerous additional components and systems that cooperate to generate a power output. In particular, internal combustion engine 12 may include a piston 24 slidably disposed within each of the plurality of cylinders 22, a crankshaft 26, a connecting rod 28 operatively connecting each piston 24 with crankshaft 26, a cylinder head 30 associated with each of cylinders 22, a series of valve actuation assemblies 32, and a control system 34.
Crankshaft 26 may be rotatably disposed within engine block 20 and operably connected to each piston 24. In particular, one connecting rod 28 may couple each piston 24 to crankshaft 26 so that a sliding motion of pistons 24 within cylinders 22 results in a rotation of crankshaft 26. Similarly, a rotation of crankshaft 26 may result in a sliding motion of pistons 24. For example, an uppermost position of each piston 24 in cylinder 22 may correspond to a top dead center position of crankshaft 26, and a lowermost position of each piston 24 in cylinder 22 may correspond to a bottom dead center position of crankshaft 26.
Internal combustion engine 12 may be a four-stroke engine, wherein piston 24 reciprocates between the uppermost position and the lowermost position during a combustion (or expansion) stroke, an exhaust stroke, and intake stroke, and a compression stroke. As piston 24 reciprocates between the upper most and lower most positions, crankshaft 26 may rotate from the top dead center position to the bottom dead center position during the combustion stroke, from the bottom dead center to the top dead center during the exhaust stroke, from top dead center to bottom dead center during the intake stroke, and from bottom dead center to top dead center during the compression stroke. Each stroke of piston 24 correlates to about 180° of crankshaft rotation, or crank angle. Thus, the combustion stroke may begin at about 0° crank angle, the exhaust stroke at about 180°, the intake stroke at about 360°, and the compression stroke at about 540°.
Each cylinder head 30 may be connected with engine block 20 and associated with one cylinder 22 to form a combustion chamber 36. It is also contemplated that one cylinder head 30 may alternatively be associated with multiple cylinders 22 to form multiple combustion chambers 36. Each cylinder head 30 may define an exhaust port 38 associated with each cylinder 22 that leads from the respective cylinder 22 to an exhaust passageway 40, and an intake port 42 that leads from the respective cylinder 22 to an intake passageway 44. Exhaust passageway 40 may direct exhaust fluid from exhaust port 38 to exhaust manifold 18 (referring to FIG. 1), while intake passageway 44 may provide fluid, for example air or a fuel/air mixture, from intake manifold 16 (referring to FIG. 1) to intake port 42. It is contemplated that cylinder head 30 may define multiple exhaust ports 38 and/or multiple intake ports 42 for each cylinder 22. Intake manifold 16 and exhaust manifold 18 may be constructed as a single integral parts or, alternatively, may be constructed as multi-part manifolds, depending upon the particular application.
Each valve actuation assembly 32 may be configured to open and close at least one exhaust port 38 and/or at least one intake port 42. Specifically, internal combustion engine 12 may include an exhaust valve 46 disposed within each exhaust port 38. Each exhaust valve 46 may include a valve stem 48 having a valve head 50, valve head 50 being sized and arranged to selectively close exhaust port 38. Similarly, internal combustion engine 12 may include an intake valve 52 with a valve stem 54 and a head 56 at a first end of valve stem 54, head 56 being sized and arranged to selectively close intake port 42. As described in greater detail below, each exhaust valve 46 and intake valve 52 may be actuated to move or “lift” valve heads 50 and 56 to thereby open the respective exhaust and intake ports 38 and 42. In a cylinder 22 having a pair of exhaust valves 46 and a pair of intake valves 52, each of the pairs of exhaust and intake valves 46, 52 may be actuated by a single valve actuation assembly 32 or by a pair of valve actuation assemblies 32.
As illustrated in FIG. 3, each valve actuation assembly 32 may include a rocker arm 58 having a first end 60, a second end 62, and a pivot point 64. The first end 60 of rocker arm 58 may operatively engage valve head 56 of intake valve 52 through a bridge 53 operatively engaged with valve stem 54 of two intake valves 52. It is contemplated that bridge 53 may be omitted, if desired, when valve actuation assembly 32 includes only a single intake valve 52 associated with each cylinder 22. The second end 62 of rocker arm 58 may be operatively associated with a pushrod 66. Intake valve 52 may be movable between a first position permitting flow from intake passageway 44 to enter combustion chamber 36 and a second position substantially blocking flow from intake passageway 44 to combustion chamber 36. It is contemplated that pushrod 66 may be omitted, if desired.
Valve actuation assembly 32 may also include a valve spring 68 configured to bias intake valve 52. Specifically, valve spring 68 may act on valve stem 54 to move valve head 56 of intake valve 52 relative to cylinder head 30. In one embodiment, valve spring 68 may bias valve head 56 of intake valve 52 into the first position, where valve head 56 engages a valve seat 72 within cylinder head 30 to prevent a flow of fluid relative to intake port 42. Valve head 56 of intake valve 52 may be movable against the bias of valve spring 68 toward a second position where valve head 56 is away from valve seat 72 to allow a flow of fluid relative to intake port 42.
Internal combustion engine 12 may also include a camshaft 74 operatively connected to crankshaft 26 and carrying a cam 76 with one or more lobes 78. Camshaft 74 may be operatively connected with crankshaft 26 in any manner readily apparent to one skilled in the art where a rotation of crankshaft 26 may result in a corresponding rotation of the camshaft 74. For example, camshaft 74 may be connected to crankshaft 26 through a gear train (not shown) that reduces the rotational speed of camshaft 74 to approximately one half of the rotational speed of crankshaft 26.
Intake valve 52 may be arranged to operate cyclically based on the configuration of cam 76, lobe 78, and the rotation of camshaft 74 to achieve a desired predetermined intake valve timing. As will be explained in greater detail below, the shape of cam lobe 78 on cam 76 may determine, at least in part, the actuation timing of valve head 56 of intake valve 52. It is contemplated that the distance between an outer edge of lobe 78 may vary and/or that cam 76 may include a greater number of cam lobes and/or a cam lobe having a different configuration depending upon the desired intake valve actuation timing.
A rotation of cam 76 may cause a cam follower 88 and associated pushrod 66 to periodically reciprocate between an upper position and a lower position. The reciprocating movement of pushrod 66 may cause rocker arm 58 to pivot about pivot point 64. When pushrod 66 moves in the direction indicated by an arrow 90, rocker arm 58 may pivot and move bridge 53 in the opposite direction. The movement of bridge 53 may cause each intake valve 52 to lift from valve seat 72 and open intake port 42. As cam 76 continues to rotate, valve spring 68 may act on bridge 53 to return each intake valve 52 to the closed position. In one embodiment, lobe 78 may be configured to operate intake valve 52 in an Otto or diesel cycle, whereby intake valve 52 moves to the second or closed position from between about 10° before bottom dead center of the intake stroke and about 10° after bottom dead center of the compression stroke. In this manner, the shape and orientation of cam 76 may control the timing of the actuation of intake valves 52.
Cam 76 may be configured to coordinate the actuation of the intake valves 52 with the movement of pistons 24. For example, each intake valve 52 may be actuated to open intake port 42 when the associated piston 24 is withdrawing within cylinder 22 to allow air to flow from intake passageway 44 into the combustion chamber 36.
Exhaust valve 46 (referring to FIG. 2) may be configured in a manner similar to intake valve 52 and may be operated by a lobe (not shown) of cam 76. Alternatively, a second cam (not shown) may be connected to crankshaft 26 to control the actuation timing of the exhaust valve 46. Exhaust valve 46 may be actuated to open exhaust port 38 when piston 24 is advancing within cylinder 22 to allow exhaust to flow from cylinder 22 into exhaust passageway 40.
Valve actuation assembly 32 may also include a variable valve closing mechanism 92 that is structured and arranged to selectively interrupt cyclical movement of intake valve 52 initiated by cam 76. For example, variable valve closing mechanism 92 may be selectively operated to supply hydraulic fluid, for example, at a low pressure or a high pressure, in a manner to resist closing of intake valve 52 by the bias of valve spring 68. That is, after intake valve 52 is lifted (i.e., opened by cam 76) and when cam 76 is no longer holding intake valve 52 open, the hydraulic fluid may hold intake valve 52 open for a desired period. The desired period may change depending on the desired performance of internal combustion engine 12 to operate under a conventional Otto or diesel cycle or under a variable Miller cycle.
Variable valve closing mechanism 92 may include numerous components that cooperate to hold intake valve 52 in the open position. In particular, variable valve closing mechanism 92 may include a housing 320 that slidably receives a piston 322 having an end 324. End 324 of piston 322 may be adapted to engage end 60 of rocker arm 58. One skilled in the art will recognize that end 324 of piston 322 may engage another portion of rocker arm 58 or may be operatively engaged with intake valves 52 in another way.
FIG. 4 illustrates housing 320 of variable valve closing mechanism 92 defining a tank 326, a bore 328, and a fluid passageway 334 that connects tank 326 and bore 328. Tank 326 may be adapted to store a supply of fluid. Tank 326 may store any type of fluid such as, for example, an engine lubricating oil. Bore 328 may be adapted to slidably receive piston 322. A seal 330 may be disposed between piston 322 and bore 328 and may be any type of sealing element such as, for example, an o-ring, that is adapted to minimize fluid leakage from bore 328 past piston 322. Fluid may flow within fluid passageway 334 from tank 326 to bore 328 as spring 332 biases piston 322 away from housing 320 in the direction of arrow 335.
Spring 332 may be any mechanical biasing element adapted to bias piston 322 away from housing 320. The force exerted by spring 332 may be less than the force exerted by springs 68 on bridge 53 (referring to FIG. 3).
Variable valve closing mechanism 92 may also include a control valve 336 disposed within fluid passageway 334. Control valve 336 may be movable between a first position where fluid is allowed to flow through fluid passageway 334 and a second position where fluid is prevented from flowing through fluid passageway 334. Thus, by controlling the position of control valve 336, the rate of fluid flow between tank 326 and bore 328 may be controlled.
A snubbing valve 338 may be disposed between bore 328 and control valve 336. Snubbing valve 338 may be configured to decrease the rate at which fluid exits bore 328 to thereby slow the rate at which piston 322 retracts within bore 328. Snubbing valve 338 may include one or more passageway 340 having openings that connect bore 328 with the fluid line leading to control valve 336.
A bleed valve 342 may be disposed in housing 320. Bleed valve 342 may be adapted to allow air, or any other gas, that finds its way into tank 326 to be released from tank 326. This will prevent air from being passed from tank 326 to bore 328. Bleed valve 342 may also purge air from anywhere within variable valve closing mechanism 92.
Variable valve closing mechanism 92 may also include an accumulator 344 having a piston 346 disposed within a chamber 348, and spring 350 configured to act on piston 346. Fluid entering chamber 348 may act to move piston 346 and compress spring 350 if the force exerted by the fluid on piston 346 is great enough to overcome the force of spring 350. Spring 350 may act to move piston 346 and force fluid out of chamber 348 when the force of spring 350 is greater than the force exerted by the pressurized fluid on piston 346.
A fluid passageway 352 may connect accumulator 344 to fluid passageway 334 between tank 326 and bore 328. A restrictive orifice 354 may be disposed in an inlet to accumulator 344. As described in greater detail below, accumulator 344 may act to dampen oscillations in bore 328 and fluid passageway 334, which could otherwise cause piston 322 to oscillate relative to housing 320.
Housing 320 of variable valve closing mechanism 92 may also include one or more leak passageways 356 and 358. Leak passageways 356 and 358 may be adapted to allow fluid that leaks from either control valve 336 or accumulator 344 to return to tank 326. As described in greater detail below, both control valve 336 and accumulator 344 may be exposed to fluid having a substantial pressure. Leak passageways 356 and 358 may help prevent any fluid that leaks through control valve 336 or accumulator 344 from leaking from variable valve closing mechanism 92.
Housing 320 of variable valve closing mechanism 92 may be connected to cylinder head 30. For example, a pair of supports 357 may extend from housing 320 to cylinder head 30. Supports 357 may be attached to cylinder head 30 by any connecting member readily apparent to one skilled in the art. For example, bolts 359 may connected supports 357 to cylinder head 30.
FIG. 5 illustrates an alternate embodiment of variable valve closing mechanism 92, wherein a spring loaded piston 360 is disposed within tank 326. Spring-loaded piston 360 act to exert a force on fluid contained within tank 326. The force of spring-loaded piston 360 may increase the pressure of the fluid within tank 326 to thereby move fluid through fluid passageway 334 to bore 328.
FIG. 6 illustrates an additional embodiment of variable valve closing mechanism 92. Variable valve closing mechanism 92 of FIG. 6 includes a third chamber 362 disposed between tank 326 and bore 328. Third chamber 362 may include a spring-loaded piston 364 and a check valve 366 disposed within spring-loaded piston 364. Check valve 366 may be configured to allow fluid to flow from tank 326 towards bore 328. In this manner, check valve 366 may be adapted to allow for the replacement of fluid that may leak from variable valve closing mechanism 92.
Control system 34 may include a controller 104 electrically connected to variable valve closing mechanism 92 of FIGS. 4-6 via a communication line 368. Controller 104 may be configured to control operation of variable valve closing mechanism 92 based on one or more current engine operating conditions. In particular, controller 104 may be programmed to receive information from one or more sensors (not shown) operatively connected with internal combustion engine 12. Each of the sensors may be configured to sense one or more operational parameters of internal combustion engine 12. For example, internal combustion engine 12 may be equipped with sensors configured to sense one or more of the following: a temperature of an engine coolant, a temperature of internal combustion engine 12, an ambient air temperature, an engine speed, a load on internal combustion engine 12, and an intake air pressure. It is contemplated that controller 104 may also be configured to control operation of variable valve closing mechanism 92 based on instructions received from an operator. It should be appreciated that the functions of controller 104 may be performed by a single controller or by a plurality of controllers. Controller 104 may include an electronic control module (not shown) that has a microprocessor and a memory. As is known to those skilled in the art, the memory may be connected to the microprocessor and configured to store an instruction set and variables. Associated with the microprocessor and part of the electronic control module may be various other known circuits such as, for example, power supply circuitry, signal conditioning circuitry, and solenoid driver circuitry, among others.
Control system 34 may be further equipped with a sensor configured to monitor the crank angle of crankshaft 26 to thereby determine the position of pistons 24 within their respective cylinders 22. The crank angle of crankshaft 26 may also be related to actuation timing of intake valves 52 and exhaust valves 46. For example, exhaust valve actuation may be timed to substantially coincide with the exhaust stroke of piston 24 and, as indicated in an exemplary graph 106 of FIG. 7, intake valve actuation 110 may be timed to substantially coincide with the intake stroke of the piston 24. FIG. 7 also illustrates valve lift for an exemplary conventional closing 314 and an exemplary late closing 316 of intake valve 52.
Intake valve 52 may begin to open at about 360° crank angle, that is, when the crankshaft 26 is at or near a top dead center position of the intake stroke. The closing of intake valve 52 may be selectively extended by the engagement of end 324 of piston 322 with first end 60 of rocker arm 58 from the cyclical closing initiated by the rotation of cam 76. That is, after intake valve 52 is lifted (i.e., opened by cam 76), and when cam 76 is no longer holding intake valve 52 open, variable valve closing mechanism 92 may hold intake valve 52 open for a desired period. The desired period may change depending on the desired performance of internal combustion engine 12. In one example, the closing of intake valve 52 may be extended from about 540° crank angle, that is, when the crank shaft is at or near a bottom dead center position of the compression stroke, to about 650° crank angle, that is, about 70° before top center of the combustion stroke. Thus, intake valve 52 may be held open for a majority portion of the compression stroke, that is, for the first half of the compression stroke and a portion of the second half of the compression stroke. It is contemplated that variable valve closing mechanism 92 may keep intake valve 52 open at an intermediate position between fully open and fully closed for a period of time or may allow continuous movement of intake valve 52 with a delayed and/or slowed closing.
Although some examples described herein involve late intake valve closure, it should be understood that certain examples in accordance with the present invention might involve engine operation where both late and early intake valve closure is selectively provided, or engine operation where only early intake valve closure is selectively provided. For example, in some exemplary engines cam 76 could have an alternative profile providing cyclical early intake valve closure and the variable valve closing mechanism 92 may be controlled to selectively delay the intake valve closing so that the delayed intake valve closing occurs before, at, and/or after bottom dead center of the intake stroke.
As illustrated in FIG. 8, internal combustion engine 12 may also include a fuel injector assembly 116 configured to inject fuel or otherwise spray fuel, for example, diesel fuel, directly into each combustion chamber 36 via a fuel port 118 within cylinder head 30 in accordance with a desired timing. Fuel injector assembly 116 may embody a mechanically-actuated, electronically-controlled unit injector, in fluid communication with a common fuel rail (not shown). Alternatively, fuel injector assembly 116 may be any common rail type injector and may be actuated and/or operated hydraulically, mechanically, electrically, piezo-electrically, or any combination thereof. The common fuel rail may provide fuel to the fuel injector assembly 116 associated with each combustion chamber 36. Fuel injector assembly 116 may be associated with an injector rocker arm 120 and pivotally coupled to a rocker shaft 122. Each fuel injector assembly 116 may include an injector body 124, a solenoid 126, a plunger assembly 128, and an injector tip assembly 130. A first end 132 of injector rocker arm 120 may be operatively coupled to plunger assembly 128. Plunger assembly 128 may be biased by a spring 134 toward the first end 132 of injector rocker arm 120 in the general direction of an arrow 138.
A second end 140 of injector rocker arm 120 may be operatively coupled to a camshaft 142. More specifically, camshaft 142 may include a cam lobe 144 having a first bump 146 and a second bump 148. Camshafts 74, 142 and their respective lobes 78, 144 may be combined into a single camshaft (not shown) if desired. First and second bumps 146, 148 may be moved into and out of contact with the second end 140 of injector rocker arm 120 during rotation of the camshaft 142. First and second bumps 146, 148 may be structured and arranged such that second bump 148 may provide a pilot injection of fuel at a predetermined crank angle before first bump 146 provides a main injection of fuel. It should be appreciated that cam lobe 144 may have only a first bump 146 that injects all of the fuel per cycle.
When one of first and second bumps 146, 148 is rotated into contact with injector rocker arm 120, the second end 140 of injector rocker arm 120 may be urged in the general direction of arrow 138. As the second end 140 is urged in the general direction of arrow 138, injector rocker arm 120 may pivot about rocker shaft 122 thereby causing the first end 132 to be urged in the general direction of an arrow 150. The force exerted on the second end 140 by first and second bumps 146, 148 may be greater in magnitude than the bias generated by spring 134, thereby causing plunger assembly 128 to be likewise urged in the general direction of arrow 150. When camshaft 142 is rotated beyond the maximum height of first and second bumps 146, 148, the bias of spring 134 may urge plunger assembly 128 in the general direction of arrow 138. As plunger assembly 128 is urged in the general direction of arrow 138, the first end 132 of injector rocker arm 120 is likewise urged in the general direction of arrow 138, which causes injector rocker arm 120 to pivot about rocker shaft 122, thereby causing the second end 140 to be urged in the general direction of arrow 150.
Injector body 124 may define a fuel port 152. Fuel, such as diesel fuel, may be drawn or otherwise aspirated into fuel port 152 from the fuel rail when plunger assembly 128 is moved in the general direction of arrow 138. Fuel port 152 may be in fluid communication with a fuel valve 154 via a first fuel channel 156. Fuel valve 154 may be, in turn, in fluid communication with a plunger chamber 158 via a second fuel channel 160.
Controller 104 may be configured to affect operation of fuel injector assembly 116. Specifically, solenoid 126 may be electrically coupled to controller 104 and mechanically coupled to fuel valve 154. Actuation of solenoid 126 by a signal from controller 104 may cause fuel valve 154 to be switched from an open position to a closed position. When fuel valve 154 is in its open position, fuel may advance from fuel port 152 to plunger chamber 158, and vice versa. However, when fuel valve 154 is in its closed positioned, fuel port 152 may be isolated from plunger chamber 158.
Injector tip assembly 130 may include a check valve assembly 162. Fuel may be advanced from plunger chamber 158, through an inlet orifice 164, a third fuel channel 166, an outlet orifice 168, and into cylinder 22 of internal combustion engine 12.
Thus, it should be appreciated that when one of first and second bumps 146, 148 is not in contact with injector rocker arm 120, plunger assembly 128 may be urged in the general direction of arrow 138 by spring 134, thereby causing fuel to be drawn into fuel port 152, which in turn fills plunger chamber 158 with fuel. As camshaft 142 is further rotated, one of first and second bumps 146, 148 may be moved into contact with injector rocker arm 120, thereby causing plunger assembly 128 to be urged in the general direction of arrow 150. If controller 104 is not generating an injection signal, fuel valve 154 may remain in its open position, thereby causing the fuel which is in plunger chamber 158 to be displaced by plunger assembly 128 through fuel port 152. However, if controller 104 is generating an injection signal, fuel valve 154 may be positioned in its closed position thereby isolating plunger chamber 158 from fuel port 152. As plunger assembly 128 continues to be urged in the general direction of arrow 150 by camshaft 142, fluid pressure within fuel injector assembly 116 may increase. At a predetermined pressure magnitude, for example, at about 5500 psi (38 MPa), fuel may be injected into combustion chamber 36. Fuel may continue to be injected into combustion chamber 36 until controller 104 signals solenoid 126 to return fuel valve 154 to its open position.
As shown in the exemplary graph of FIG. 9, the pilot injection of fuel may commence when crankshaft 26 is at about 675° crank angle, that is, about 45° before top dead center of the compression stroke. The main injection of fuel may occur when crankshaft 26 is at about 710° crank angle, that is, about 10° before top dead center of the compression stroke and about 45° after commencement of the pilot injection. Generally, the pilot injection may commence when crankshaft 26 is about 40-50° before top dead center of the compression stroke and may last for about 10-15° of crankshaft rotation. The main injection may commence when the crankshaft 26 is between about 10° before top dead center of the compression stroke and about 12° after top dead center of the combustion stroke. The main injection may last for about 20-45° of crankshaft rotation. The pilot injection may use a desired portion of the total fuel used, for example about 10%.
As illustrated in FIG. 1, air supply system 14 may include components that fluidly communicate with intake manifold 16 and the exhaust manifold 18. In particular, air supply system 14 may include a first turbocharger 170 and a second turbocharger 172. First and second turbochargers 170, 172 may be arranged in series with one another such that second turbocharger 172 provides a first stage of pressurization and first turbocharger 170 provides a second stage of pressurization. For example, second turbocharger 172 may be a low pressure turbocharger and first turbocharger 170 may be a high pressure turbocharger. First turbocharger 170 may include a turbine 174 and a compressor 176. Turbine 174 may be fluidly connected to exhaust manifold 18 via an exhaust duct 178 and may include a turbine wheel 180 carried by a shaft 182. Shaft 182 may be rotatably carried by a housing 184, for example, a single-part or multi-part housing. The fluid flow path from exhaust manifold 18 to turbine 174 may include a variable nozzle (not shown) or other variable geometry arrangement adapted to control the velocity of exhaust fluid impinging on turbine wheel 180. Compressor 176 may be fluidly connected to intake manifold 16 and may include a compressor wheel 186 carried by shaft 182. Thus, rotation of shaft 182 by turbine wheel 180 may cause rotation of compressor wheel 186.
First turbocharger 170 may include a compressed air duct 188 for receiving compressed air from the second turbocharger 172 and an air outlet line 190 for receiving compressed air from compressor 176 and supplying the compressed air to intake manifold 16 of the internal combustion engine 12. First turbocharger 170 may also include an exhaust duct 192 for receiving exhaust fluid from turbine 174 and supplying the exhaust fluid to second turbocharger 172.
Second turbocharger 172 may include a turbine 194 and a compressor 196. Turbine 194 may be fluidly connected to exhaust duct 192 and may include a turbine wheel 198 carried by a shaft 200, which in turn may be rotatably carried by housing 184. Compressor 196 may include a compressor wheel 202 also carried by shaft 200. Thus, rotation of shaft 200 by the turbine wheel 198 may in turn cause rotation of compressor wheel 202.
Second turbocharger 172 may include an air intake line 204 providing fluid communication between the atmosphere and compressor 196. Second turbocharger 172 may also supply compressed air to first turbocharger 170 via compressed air duct 188. Second turbocharger 172 may include an exhaust outlet 206 for receiving exhaust fluid from turbine 194 and providing fluid communication with the atmosphere. In one embodiment, first turbocharger 170 and second turbocharger 172 may be sized to provide substantially similar compression ratios. For example, first turbocharger 170 and second turbocharger 172 may both provide compression ratios of between 2:1 and 3:1, resulting in a system compression ratio of at least 4:1 with respect to atmospheric pressure. Alternatively, second turbocharger 172 may provide a compression ratio of 3:1 and first turbocharger 170 may provide a compression ratio of 1.5:1, resulting in a system compression ratio of 4.5:1 with respect to atmospheric pressure.
Air supply system 14 may include an air cooler 208, for example, an aftercooler, between compressor 176 and the intake manifold 16. Air cooler 208 may extract heat from the air to lower the intake manifold temperature and increase the air density. Optionally, air supply system 14 may include an additional air cooler 210, for example, an intercooler, between compressor 196 of second turbocharger 172 and compressor 176 of first turbocharger 170. Intercooling may use techniques such as jacket water, air to air, and the like. Alternatively, air supply system 14 may optionally include an additional air cooler (not shown) between air cooler 208 and intake manifold 16. The optional additional air cooler may further reduce the intake manifold temperature. A jacket water pre-cooler (not shown) may be used to protect air cooler 208.
FIG. 10 illustrates an alternate air supply system 212 for internal combustion engine 12. Air supply system 212 may include a turbocharger 214, for example, a high-efficiency turbocharger capable of producing at least about a 4:1 compression ratio with respect to atmospheric pressure. Turbocharger 214 may include a turbine 216 and a compressor 218. Turbine 216 may be fluidly connected to exhaust manifold 18 via an exhaust duct 220. Turbine 216 may include a turbine wheel 222 carried by a shaft 224, which in turn may be rotatably carried by a housing 226, for example, a single-part or multi-part housing. The fluid flow path from exhaust manifold 18 to the turbine 216 may include a variable nozzle (not shown), which may control the velocity of exhaust fluid impinging on the turbine wheel 222.
Compressor 218 may include a compressor wheel 228 carried by shaft 224. Thus, rotation of shaft 224 by the turbine wheel 222 in turn may cause rotation of compressor wheel 228. Turbocharger 214 may include an air inlet 230 providing fluid communication between the atmosphere and compressor 218 and an air outlet 232 for supplying compressed air to intake manifold 16 of internal combustion engine 12. Turbocharger 214 may also include an exhaust outlet 234 for receiving exhaust fluid from turbine 216 and providing fluid communication with the atmosphere.
Air supply system 212 may include an air cooler 236 between compressor 218 and the intake manifold 16. Optionally, air supply system 212 may include an additional air cooler (not shown) between air cooler 236 and intake manifold 16.
FIG. 11 illustrates another alternate air supply system 238 for internal combustion engine 12. Air supply system 238 may include a turbocharger 240 having a turbine 242 and first and second compressors 244, 246. Turbine 242 may be fluidly connected to exhaust manifold 18 via an inlet duct 248. Turbine 242 may include a turbine wheel 250 carried by a shaft 252, which in turn may be rotatably carried by a housing 254, for example, a single-part or multi-part housing. The fluid flow path from exhaust manifold 18 to turbine 242 may include a variable nozzle (not shown), which may control the velocity of exhaust fluid impinging on the turbine wheel 250.
First compressor 244 may include a compressor wheel 260 carried by shaft 252, and second compressor 246 may include a compressor wheel 262 carried by shaft 252. Thus, rotation of shaft 252 by turbine wheel 250 in turn may cause rotation of first and second compressor wheels 260, 262. First and second compressors 244, 246 may provide first and second stages of pressurization, respectively.
Turbocharger 240 may include an air intake line 264 providing fluid communication between the atmosphere and first compressor 244 and a compressed air duct 266 for receiving compressed air from first compressor 244 and supplying the compressed air to second compressor 246. Turbocharger 240 may also include an air outlet line 268 for supplying compressed air from second compressor 246 to intake manifold 16 of internal combustion engine 12. Turbocharger 240 may further include an exhaust outlet 270 for receiving exhaust fluid from turbine 242 and providing fluid communication with the atmosphere.
For example, first compressor 244 and second compressor 246 may both provide compression ratios of between 2:1 and 3:1, resulting in a system compression ratio of at least 4:1 with respect to atmospheric pressure. Alternatively, second compressor 246 may provide a compression ratio of 3:1 and first compressor 244 may provide a compression ratio of 1.5:1, resulting in a system compression ratio of 4.5:1 with respect to atmospheric pressure.
Air supply system 238 may include an air cooler 272 between second compressor 246 and intake manifold 16. Optionally, air supply system 238 may include an additional air cooler 274 between first compressor 244 and second compressor 246 of turbocharger 240. Alternatively, air supply system 238 may optionally include an additional air cooler (not shown) between air cooler 272 and intake manifold 16.
FIG. 12 illustrates an exemplary exhaust gas recirculation (EGR) system 276 associated with an exhaust system 278 of internal combustion engine 12. In this embodiment, internal combustion engine 12 includes an air supply system 280 having two-stage turbocharging similar to air supply system 14 of FIG. 1. Air supply system 280 may include a first turbocharger 282 having a turbine 284 and a compressor 286. Air supply system 280 may also include second turbocharger 288 having turbine 290 and compressor 292. The two-stage turbocharger system operates to increase the pressure of the air and exhaust gases being delivered to cylinders 22 via intake manifold 16, and to maintain a desired air to fuel ratio during extended open durations of intake valves 52 of cylinders 22 (referring to FIG. 2). It is noted that a two-stage turbocharger system is not required for operation of the present invention. Other types of turbocharger systems, such as the high pressure ratio single-stage turbocharger system of FIG. 10, a variable geometry turbocharger system, and the like, may be used instead.
A throttle valve 294, may be located between compressor 286 and intake manifold 16 and used to control the amount of air and recirculated exhaust gases being delivered to cylinders 22. Although throttle valve 294 is shown between compressor 286 and an aftercooler 296, throttle valve 294 may be alternatively positioned at other locations such as, for example, after aftercooler 296.
Although EGR system 276 is a low pressure EGR system, variations of EGR system 276 may be equally used with the present invention, including both low pressure loop and high pressure loop EGR systems. Other types of EGR systems, such as for example by-pass, venturi, piston-pumped, peak clipping, and back pressure, could be used.
An oxidation catalyst 298 may receive exhaust gases from turbine 290 and serve to reduce HC emissions. Oxidation catalyst 298 may also be coupled with a de-NOx, catalyst to further reduce oxides of nitrogen (NOx) emissions. A particulate matter (PM) filter 300 may receive exhaust gases from oxidation catalyst 298. Although oxidation catalyst 298 and PM filter 300 are shown as separate items, they may alternatively be combined into one package.
Some of the exhaust gases may be delivered to the atmosphere from PM filter 300, while a portion of the exhaust gases may be routed to intake manifold 16 through an EGR cooler 302, through an EGR valve 304, and through first and second turbochargers 282, 288. EGR cooler 302 may be of a type well known in the art, for example a jacket water or an air to gas heat exchanger type.
A means 306 for determining pressure within PM filter 300 is shown. In the preferred embodiment, the means 306 for determining pressure may include a pressure sensor 308. However, other alternate means for determining pressure may be employed. For example, the pressure of the exhaust gases in PM filter 300 may be estimated from a model based on one or more parameters associated with internal combustion engine 12. Parameters may include, but are not limited to, engine load, engine speed, temperature, fuel usage, and the like.
A means 310 for determining flow of exhaust gases through PM filter 300 may be used. Preferably, means 310 for determining flow of exhaust gases may include a flow sensor 312. Flow sensor 312 may be used alone to determine pressure in PM filter 300 based on changes in flow of exhaust gases, or may be used in conjunction with pressure sensor 308 to provide more accurate pressure change determinations.

INDUSTRIAL APPLICABILITY

An air and fuel supply system for an internal combustion engine in accordance with the exemplary embodiments of the invention may extract additional work from the engine's exhaust. The system may achieve fuel efficiency and reduce NOx emissions, while maintaining work potential and ensuring that the system reliability meets with operator expectations. The operation of internal combustion engine 12 will now be explained.
Internal combustion engine 12 may function in a known manner using, for example, the diesel principle of operation. Internal combustion engine 12 may be used with each of the exemplary air supply systems 14, 212, 238, and 280 of FIGS. 1 and 7-9, respectively. Compressed air may be supplied from air supply systems 14, 212, 238, and 280 to combustion chambers 36 via intake port 42, and exhaust air may exit combustion chambers 36 to air supply systems 14, 212, 238, and 280 via exhaust port 38. Intake valve 52 and exhaust valve 46 may be controllably moved to direct airflow into and exhaust out of combustion chambers 36.
In a conventional Otto or diesel cycle mode, intake valve 52 may move from the second or closed position to the first or open position in a cyclical fashion to allow compressed air to enter combustion chamber 36 of cylinder 22 at or near top dead center of the intake stroke (about 360° crank angle), as shown in FIG. 7. At or near bottom dead center of the compression stroke (about 540° crank angle), intake valve 52 may move from the first position to the second position to block additional air from entering combustion chamber 36. Fuel may then be injected from fuel injector assembly 116 (referring to FIG. 8) at or near top dead center of the compression stroke (about 720° crank angle).
In accordance with the Miller cycle, the Otto or diesel cycle may be modified by moving intake valve 52 from the first position to the second position at either some predetermined time before bottom dead center of the intake stroke (i.e., before 540° crank angle) or some predetermined time after bottom dead center of the compression stroke (i.e., after 540° crank angle). In a late-closing Miller cycle, intake valve 52 may be moved from the first position to the second position during a first portion of the first half of the compression stroke.
Variable valve closing mechanism 92 may enable internal combustion engine 12 to be operated in an early-closing Miller cycle, a late-closing Miller cycle, and a conventional Otto or diesel cycle. Further, injecting a substantial portion of fuel after top dead center of the combustion stroke, as illustrated in FIG. 9, may reduce NOx emissions and increase the amount of energy rejected to exhaust manifold 18 in the form of exhaust fluid. Use of a high- efficiency turbocharger 214, 240 or series turbochargers 170-172, 282-288 may enable recapture of at least a portion of the rejected energy from the exhaust. The rejected energy may be converted into increased air pressures delivered to the intake manifold 16, which may increase the energy pushing piston 24 against crankshaft 26 to produce useable work. In addition, delaying movement of intake valve 52 from the open position to the closed position may reduce the compression temperature in combustion chamber 36. The reduced compression temperature may further reduce NOx emissions.
Controller 104 may operate variable valve closing mechanism 92 to vary the timing of intake valve 52 to achieve desired engine performance based on one or more engine conditions, for example, engine speed, engine load, engine temperature, boost, and/or manifold intake temperature. Variable valve closing mechanism 92 may also allow more precise control of the air/fuel ratio. By delaying closing of intake valve 52, controller 104 may control the cylinder pressure during the compression stroke of piston 24. For example, late closing of intake valve 52 may reduce the compression work that piston 24 must perform without compromising cylinder pressure and while maintaining a standard expansion ratio and a suitable air/fuel ratio.
The following discussion describes the implementation of a late intake Miller cycle in a single cylinder 22 of internal combustion engine 12. One skilled in the art will recognize that the system of the present invention may be used to selectively implement a late intake Miller cycle in all cylinders 22 of internal combustion engine 12 in the same or a similar manner. In addition, the disclosed system may be used to implement other valve actuation variations on the conventional diesel cycle, such as, for example, an exhaust Miller cycle, an early closing Miller cycle, and other variations known in the art.
When internal combustion engine 12 is operating under normal operating conditions, controller 104 may implement a late intake Miller cycle by controlling the position of control valve 336 in variable valve closing mechanism 92. The rotation of cam 76 may cause rocker arm 58 to pivot to thereby actuate intake valves 52 (referring to FIG. 3). The force of spring 332 may cause piston 322 to extend in the direction of arrow 335 (referring to FIG. 4), to thereby follow the motion of end 60 of rocker arm 58.
The movement of piston 322 in bore 328 draws fluid into bore 328 from fluid passageway 334 and tank 326. The flow of fluid into bore 328 may be aided by spring-loaded piston 360, which may be disposed within tank 326 (referring to FIG. 5) or by spring-loaded piston 364 disposed within third chamber 360 (referring to FIG. 6). Spring-loaded pistons 360 and 364 may force fluid through fluid passageway 334 into bore 328 to ensure that bore 328 is filled with fluid.
Controller 104 may send a signal to adjust the position of control valve 336 to close passageway 334 and thereby trap fluid in bore 328 when piston 322 is partially or fully extended from housing 320. For example, controller 104 may close control valve 336 when intake valves 52 are at or near a maximum lift position such as, for example, at a peak 218 (referring to FIG. 7). Also, controller 104 may time the closing of control valve 336 to ensure that bore 328 is filled with fluid before control valve 336 is moved to the closed position.
As cam 76 continues to rotate, springs 68 may urge intake valves 52 towards their closed positions until end 60 of rocker arm 58 engages end 324 of piston 322. The fluid trapped in bore 328 may prevent piston 322 from moving with respect to housing 320 and will, therefore, prevent intake valves 52 from closing. As long as control valve 336 remains in the closed position, the trapped fluid within bore 328 will prevent springs 68 from returning intake valves 52 to the closed position. Thus, variable valve closing mechanism 92 will hold intake valves 52 in position between the open and closed positions, independently of the actuation of cam 76.
When rocker arm 58 engages piston 322, the force of springs 68 acting through rocker arm 58 may cause an increase in the pressure of the fluid within variable valve closing mechanism 92. In response to the increased pressure, fluid may flow through restrictive orifice 354 in fluid passageway 352 and into accumulator 344, which may absorb the pressure spike. In this manner, accumulator 344 may act to dampen any oscillations that may result from the engagement of rocker arm 58 and piston 322.
Controller 104 may close intake valves 52 by sending a signal to adjust the position of control valve 336 to open passageway 334. This allows the trapped fluid to flow out of bore 328. The force of springs 68 may overcome the force of spring 332 and force the fluid from bore 328 towards tank 326. The release of the trapped fluid may allow piston 322 to move within housing 320. The allows rocker arm 58 to pivot so that intake valves 52 may move to the closed position.
Snubbing valve 338 may reduce the rate at which intake valve 52 moves to the closed position. As piston 322 moves within bore 328, fluid may flow through passageways 340. The body of piston 322 will eventually block the openings to passageways 340, thereby reducing the rate at which fluid may flow from bore 328. This reduction in fluid flow rate translates to a reduction in velocity of piston 322 and to a reduction in the closing or seating velocity of intake valve 52. In this manner, snubbing valve 338 may control the velocity at which intake valve 52 closes to prevent intake valve 52 from being damaged.
It should be appreciated that other alternatives exist for reducing the closing speed of intake valve 52. For example, an impact absorber (not shown) may be placed between piston 322 and rocker arm 58. The impact absorber may include a spring/damper element, for example, a self-contained hydraulic, pneumatic, or elastomeric element. As another example, a cam (not shown) may be used to reduce the closing speed of intake valve 52. Such a cam may be referred to as a “decelerating” or “handoff” cam because it reduces the closing speed of intake valve 52 at the handoff or impact point.
An exemplary late intake closing of intake valve 52 is illustrated in FIG. 7. As shown, intake valve 52 may be extended past a conventional closing 314 into a portion of the compression stroke of piston 24 during a late closing 316. This allows some of the air in combustion chamber 36 to escape, thereby changing a compression ration of internal combustion engine 12. The amount of air allowed to escape the combustion chamber 36 may determine the compression ratio of internal combustion engine 12 and may be controlled by adjusting the crank angle at which control valve 336 is opened. Control valve 336 may be opened at an earlier crank angle to decrease the amount of escaping air or at a later crank angle to increase the amount of escaping air.
The disclosed engine valve actuation system may selectively alter the timing of the intake and/or exhaust valve actuation of an internal combustion engine. The actuation of the engine valves may be based on sensed operating conditions of the engine. For example, the engine valve actuation system may implement a late intake Miller cycle when the engine is operating under normal operating conditions, and the late intake Miller cycle may be disengaged when the engine is operating under other conditions such as, for example during starting or otherwise operating under cold conditions. Thus, the present invention allows for selective disengagement of the late Miller cycle.
Controller 104 may disengage the late intake Miller cycle by leaving control valve 336 in the open position. If control valve 336 is continuously open, no fluid will be trapped within bore 328. Accordingly, piston 322 will be free to move within housing 320 and will not prevent intake valves 52 from returning to the closed position. Thus, the actuation of intake valves 52 will be driven by the shape of cam 76.
Thus, when control valve 336 is continuously open, intake valves 52 will follow a conventional diesel cycle as governed by cam 76. As illustrated in FIG. 7, intake valve actuation will follow a conventional closing 314. In the conventional closing 314, the closing of intake valves 52 may substantially coincide with the end of the intake stroke of piston 24. When intake valves 52 close at the end of the intake stroke, no air will be forced from cylinder 22 during the compression stroke. This results in piston 24 compressing the fuel and air mixture to a higher pressure, which will facilitate diesel fuel combustion. This is particularly beneficial when operating in cold conditions.
As will be apparent from the foregoing description, the describe system provides an engine valve actuation system that may selectively alter the timing of the intake and/or exhaust valve actuation of an internal combustion engine. The actuation of the engine valves may be based on sensed operating conditions of the engine. For example, the engine valve actuation system may implement a late intake Miller cycle when the engine is operating under normal operating conditions. The late intake Miller cycle may be disengaged when the engine is operating under adverse operating conditions, such as when the engine is cold. Thus, the disclosed system and method provide a flexible engine valve actuation system that provides for both enhanced cold starting capability and fuel efficiency gains.
The disclosed system and method also provides a variable valve closing mechanism that is self-contained in a single housing. All essential elements of the variable valve closing mechanism are contained in the housing, including the fluid supply reservoir. As the variable valve closing mechanism does not have to share fluid with another system in the engine, the possibility of operating fluid contamination is reduced. Also, the variable valve closing mechanism may use any type of operating fluid, including a fluid that is not affected by a change in temperature. Thus, the disclosed variable valve closing mechanism may not experience performance problems when the operating fluid is cold.
In addition, the described variable valve closing mechanism does not rely upon oil from the engine lubrication system. Accordingly, any contamination of the lubricating oil will not affect the operation of the variable valve closing mechanism. Also, as the amount of fluid stored in the variable valve closing mechanism is substantially less than the amount of oil included in the engine lubrication system, the fluid in the described variable valve closing mechanism may reach a normal operating temperature faster than the oil in the engine lubricating system. Thus, the described variable valve closing mechanism may provide for reliable and timely operation, even under undesirable conditions.
The high pressure air provided by the exemplary air supply systems 14, 212, 238, and 280 may provide extra boost on the intake stroke of piston 24 to enable intake valve 52 to be closed even later than in a conventional Miller cycle engine. In the present invention, intake valve 52 may remain open until the second half of the compression stroke of piston 24, for example, as late as about 80° to 70° before top dead center (BTDC). While intake valve 52 is open, air may flow between combustion chamber 36 and intake manifold 16. Thus, cylinder 22 experiences less of a temperature rise in combustion chamber 36 during the compression stroke of piston 24.
Since the closing of intake valve 52 may be delayed, the timing of fuel injection may also be retarded. For example, controller 104 may controllably operate fuel injector assembly 116 to supply fuel to combustion chamber 36 after intake valve 52 is closed. For example, fuel injector assembly 116 may be controlled to supply a pilot injection of fuel contemporaneous with or slightly after intake valve 52 is closed and to supply a main injection of fuel contemporaneous with or slightly before combustion temperature is reached within combustion chamber 36. As a result, a significant amount of exhaust energy may be available for recirculation by a air supply systems 14, 212, 238, and 280, which may efficiently extract additional work from the exhaust energy.
Referring to the exemplary air supply system 14 of FIG. 1, exhaust gas from internal combustion engine 12 may be directed from exhaust manifold 18 through exhaust duct 178 to impinge on and causes rotation of turbine wheel 180 of first turbocharger 170. Because turbine wheel 180 is coupled with shaft 182, which in turn carries compressor wheel 186, the rotational speed of compressor wheel 186 may correspond to the rotational speed of turbine wheel 180. Second turbocharger 172 may extract otherwise wasted energy from the exhaust stream of first turbocharger 170 to turn compressor wheel 202 of second turbocharger 172, which is in series with compressor wheel 186 of first turbocharger 170. The extra restriction in the exhaust path resulting from the addition of second turbocharger 172 may raise the back pressure on piston 24. However, the energy recovery accomplished through the use of second turbocharger 172 may offset the work consumed by the higher back pressure. For example, the additional pressure achieved by the series turbochargers 170,172 may do work on piston 24 during the induction stroke of the combustion cycle. Further, the added pressure on cylinder 22 resulting from second turbocharger 172 may be controlled and/or relieved by using the late intake valve closing. Thus, the series turbochargers 170, 172 may provide fuel efficiency via air supply system 14, and not simply more power.
It should be appreciated that air coolers 208, 236, 272, and 296 (referring to FIGS. 1 and 7-9) of air supply systems 14, 212, 238, and 280, preceding the intake manifold 16 may extract heat from the air to lower the inlet manifold temperature, while maintaining the denseness of the pressurized air. The optional additional air coolers 210, 274 between compressors 176 and 196, and 244 and 246 or after air coolers 208, 236, 272, and 296 may further reduce the inlet manifold temperature, but may lower the work potential of the pressurized air. The lower inlet manifold temperature may further reduce NOx emissions.
Referring to FIG. 12, a change in pressure of exhaust gases passing through PM filter 300 may result from an accumulation of particulate matter, thus indicating a need to regenerate PM filter 300 (i.e., burn away the accumulation of particulate matter). For example, as particulate matter accumulates, pressure in PM filter 300 may increase.
PM filter 300 may be a catalyzed diesel particulate filter (CDPF) or an active diesel particulate filter (ADPF). A CDPF allows soot to burn at much lower temperatures. An ADPF is defined by raising the PM filter internal energy by means other than internal combustion engine 12, for example electrical heating, burner, fuel injection, and the like.
One method to increase the exhaust temperature and initiate PM filter regeneration is to use throttle valve 294 to restrict inlet air, thus increasing exhaust temperature. Other methods to increase exhaust temperature may include variable geometry turbochargers, smart wastegates, variable valve actuation, and the like. Yet another method to increase exhaust temperature and initiate PM filter regeneration may include the use of a post injection of fuel ( i.e., a fuel injection timed after delivery of a main injection).
Throttle valve 294 may be coupled to EGR valve 304 so that they are both actuated together. Alternatively, throttle valve 294 and EGR valve 304 may be actuated independently of each other. Both valves may operate together or independently to modulate the rate of EGR being delivered to intake manifold 16.
CDPFs regenerate more effectively when the ratio of NOx, to particulate matter (i.e., soot) is within a certain range, for example, from about 20:1 to about 30:1. It has been found, however, that an EGR system combined with the above described methods of multiple fuel injections and variable valve timing may result in a NOx to soot ratio of about 10:1. Thus, it may be desirable to periodically adjust the levels of emissions to change the NOx, to soot ratio to a more desired range and then initiate regeneration. Examples of methods which may be used include adjusting the EGR rate and adjusting the timing of main fuel injection.
A venturi (not shown) may be used at the EGR entrance to the fresh air inlet. The venturi would depress or lower the pressure of the fresh air at the inlet, thus allowing EGR to flow from the exhaust to the intake side. The venturi may include a diffuser portion which would restore the fresh air to near original velocity and pressure prior to entry into compressor 292. The use of a venturi and diffuser may increase engine efficiency.
It will be apparent to those skilled in the art that various modifications and variations can be made in the disclosed engine valve actuation system without departing from the scope of the invention. Other embodiments of the invention will be apparent to those skilled in the art from consideration of the specification and practice of the invention disclosed herein. It is intended that the specification and examples be considered as exemplary only, with a true scope of the invention being indicated by the following claims and their equivalents.

Claims

1. A method of operating an internal combustion engine including at least one cylinder and a piston slidable in the cylinder, the method comprising:

supplying pressurized air from an intake manifold to an air intake port of a combustion chamber in the cylinder;

operating an air intake valve to open the air intake port to allow pressurized air to flow between the combustion chamber and the intake manifold substantially during a majority portion of a compression stroke of the piston; and

controlling a hydraulic actuator to close the air intake port to prohibit pressurized air from flowing between the combustion chamber and the intake manifold.

2. The method of claim 1, wherein the controlling is based on at least one engine condition.

3. The method of claim 1, further including controlling a fuel supply system to inject fuel into the combustion chamber.

4. The method of claim 3, further including injecting at least a portion of the fuel during a portion of the compression stroke.

5. The method of claim 4, wherein injecting at least a portion of the fuel includes supplying a pilot injection at a predetermined crank angle before a main injection.

6. The method of claim 5, wherein said main injection begins during the compression stroke.

7. The method of claim 1, further including cooling the pressurized air prior to supplying the pressurized air to the air intake port.

8. The method of claim 1, wherein said supplying includes supplying a mixture of pressurized air and recirculated exhaust gas from the intake manifold to the air intake port, and wherein said operating of the air intake valve includes operating the air intake valve to open the air intake port to allow the pressurized air and exhaust gas mixture to flow between the combustion chamber and the intake manifold substantially during a majority portion of the compression stroke of the piston.

9. The method of claim 8, wherein said supplying a mixture of pressurized air and recirculated exhaust gas includes providing a quantity of exhaust gas from an exhaust gas recirculation (EGR) system.

10. The method of claim 1, further including controlling the hydraulic actuator to maintain the air intake port open to allow pressurized air to flow between the combustion chamber and the intake manifold.

11. The method of claim 1, wherein the controlling includes moving a piston of the hydraulic actuator between a first position and a second position.

12. The method of claim 11, further including locking the piston in the second position.

13. The method of claim 12, wherein the piston engages the air intake valve in the second position and locking includes:

directing a flow of fluid from a tank disposed within the actuator housing to a bore in the actuator housing, the bore being associated with the piston; and

selectively preventing the fluid from flowing from the bore to the tank to trap fluid in the bore and prevent the piston from moving with respect to the actuator housing, the piston preventing the air intake valve from returning to the first position.

14. The method of claim 13, further including selectively allowing the fluid to flow from the bore to the tank to release the piston and thereby allow the air intake valve to return to the second position.

15. The method of claim 14, further including directing a portion of the flow of fluid between the tank and the bore to an accumulator.

16. An internal combustion engine, comprising:

an engine block defining at least one cylinder;

a head connected with said engine block, the head including an air intake port, and an exhaust port;

a piston slidable in the cylinder;

a combustion chamber being defined by said head, said piston, and said cylinder;

an air intake valve controllably movable to open and close the air intake port;

an air supply system including at least one turbocharger fluidly connected to the air intake port;

a fuel supply system operable to inject fuel into the combustion chamber;

a cam assembly configured to move the air intake valve; and

a hydraulic actuator configured to selectively control movement of the air intake valve.

17. The engine of claim 16, wherein the hydraulic actuator is configured to keep the intake valve open during at least a portion of a compression stroke of the piston.

18. The engine of claim 17, wherein the hydraulic actuator is configured to keep the intake valve open for a portion of a second half of the compression stroke.

19. The engine of claim 16, wherein the hydraulic actuator is configured to close the intake valve before bottom dead center of an intake stroke of the piston.

20. The engine of claim 16, wherein the at least one turbocharger includes a first turbine coupled with a first compressor, the first turbine being in fluid communication with the exhaust port, the first compressor being in fluid communication with the air intake port; and wherein the air supply system further includes a second compressor being in fluid communication with atmosphere and the first compressor.

21. The engine of claim 16, wherein the at least one turbocharger includes a first turbocharger and a second turbocharger, the first turbocharger including a first turbine coupled with a first compressor, the first turbine being in fluid communication with the exhaust port and an exhaust duct, the first compressor being in fluid communication with the air intake port, the second turbocharger including a second turbine coupled with a second compressor, the second turbine being in fluid communication with the exhaust duct of the first turbocharger and atmosphere, and the second compressor being in fluid communication with atmosphere and the first compressor.

22. The engine of claim 16, further including an exhaust gas recirculation (EGR) system operable to provide a portion of exhaust gas from the exhaust port to the air supply system.

23. The engine of claim 16, wherein hydraulic actuator includes:

an actuator housing including a tank;

a piston slidably disposed in the actuator housing and configured to engage the air intake valve; and

a control valve disposed between the tank and the piston, the control valve configured to selectively fluidly communicate the tank and the piston.

24. The engine of claim 23, wherein the tank includes a spring-loaded piston.

25. The engine of claim 23, wherein the actuator housing defines a chamber between the tank and the piston and wherein the chamber includes a spring-loaded piston.

26. The engine of claim 23, further including a snubbing valve adapted to slow a movement of the piston.

27. The engine of claim 23, further including a pivotable rocker arm operably coupling a cam assembly with the air intake valve, wherein the piston includes an end configured to selectively engage the rocker arm.

28. A method of operating an internal combustion engine including at least one cylinder and a piston slidable in the cylinder, the method comprising:

imparting rotational movement to a first turbine and a first compressor of a first turbocharger with exhaust air flowing from an exhaust port of the cylinder;

imparting rotational movement to a second turbine and a second compressor of a second turbocharger with exhaust air flowing from an exhaust duct of the first turbocharger;

compressing air drawn from atmosphere with the second compressor;

compressing air received from the second compressor with the first compressor;

supplying pressurized air from the first compressor to an air intake port of a combustion chamber in the cylinder via an intake manifold;

operating a fuel supply system to inject fuel directly into the combustion chamber;

operating a cam assembly to move the air intake valve; and

operating a hydraulic actuator to control movement of the air intake valve.

29. The method of claim 28, wherein fuel is injected during a combustion stroke of the piston.

30. The method of claim 29, wherein fuel injection begins during a compression stroke of the piston.

31. The method of claim 28, wherein said operating a hydraulic actuator includes operating the hydraulic actuator to keep open the air intake valve to allow pressurized air to flow between the combustion chamber and the intake manifold during a portion of a compression stroke of the piston.

32. The method of claim 31, wherein said operating a hydraulic actuator includes operating the hydraulic actuator to keep open the air intake valve for a portion of a second half of a compression stroke of the piston.

33. The method of claim 28, wherein said operating a hydraulic actuator includes closing the air intake valve before bottom dead center of an intake stroke of the piston.

34. The method of claim 28, further including operating the cam to move the air intake valve and cyclically open and close the air intake port, wherein said operating the hydraulic actuator includes interrupting the cyclical opening and closing of the air intake port.

35. The method of claim 28, wherein operation of the hydraulic actuator is based on at least one engine condition.

36. The method of claim 28, wherein said first and second compressors compress a mixture of air and recirculated exhaust gas, and wherein said supplying includes supplying the compressed mixture of pressurized air and recirculated exhaust gas to said intake port via said intake manifold.

37. The method of claim 28, wherein operating a hydraulic actuator includes extending a piston to engage a rocker arm operably coupled with the air intake valve.

38. The method of claim 37, further including allowing fluid to flow from a tank in an actuator housing to a bore in the actuator housing and trapping the fluid in the bore of the actuator housing to block retraction of the piston.

39. The method of claim 38, further including selectively allowing fluid to drain from the bore to the tank to release the piston and thereby allow the air intake valve to close before bottom dead center of an intake stroke of the piston.

40. The method of claim 28, further including operating the cam to move an air intake valve and cyclically open and close the air intake port, wherein said operating a hydraulic actuator includes interrupting the cyclical opening and closing of the air intake port.

41. The method of claim 28, wherein operation of the hydraulic actuator is based on at least one engine condition.

42. The method of claim 28, wherein said first and second compressors compress a mixture of air and recirculated exhaust gas, and wherein said supplying includes supplying the compressed mixture of pressurized air and recirculated exhaust gas to said intake port via said intake manifold.

43. A method of controlling an internal combustion engine having a variable compression ratio, said engine including a block defining a cylinder, a piston slidable in said cylinder, a head connected with said block, said piston, said cylinder, and said head defining a combustion chamber, the method comprising:

pressurizing air;

supplying said air to an intake manifold of the engine;

maintaining fluid communication between said combustion chamber and the intake manifold during a portion of an intake stroke and through a portion of a compression stroke;

injecting fuel directly into the combustion chamber; and

controlling the communication between said combustion chamber and the intake manifold at least in part by hydraulic actuator.

44. The method of claim 43, wherein said injecting fuel includes injecting fuel directly to the combustion chamber during a portion of a combustion stroke of the piston.

45. The method of claim 43, wherein said injecting fuel includes injecting fuel directly to the combustion chamber during a portion of the compression stroke.

46. The method of claim 43, wherein said injecting includes supplying a pilot injection at a predetermined crank angle before a main injection.

47. The method of claim 46, wherein said portion of the compression stroke is at least a majority of the compression stroke.

48. The method of claim 43, wherein said pressurizing includes a first stage of pressurization and a second stage of pressurization.

49. The method of claim 48, further including cooling air between said first stage of pressurization and said second stage of pressurization.

50. The method of claim 43, further including cooling the pressurized air.

51. The method of claim 43, wherein the pressurizing includes pressurizing a mixture of air and recirculated exhaust gas, and wherein the supplying includes supplying the pressurized air and exhaust gas mixture to the intake manifold.

52. The method of claim 51, further including cooling the pressurized air and exhaust gas mixture.

53. The method of claim 43, further including operating the hydraulic actuator to varying a closing time of an air intake valve controlling a flow to the combustion chamber.

54. The method of claim 43, further including operating the hydraulic actuator to move a piston of the hydraulic actuator between a first position and a second position and selectively blocking the piston in the second position with fluid from a tank disposed within the hydraulic actuator.

55. The method of claim 43, further including controlling the communication between said combustion chamber and the intake manifold at least in part by a cam assembly.

56. A method of operating an internal combustion engine including at least one cylinder and a piston slidable in the cylinder, the method comprising:

operating an air intake valve to open the air intake port to allow pressurized air to flow between the combustion chamber and the intake manifold substantially during a portion of a compression stroke of the piston;

injecting fuel into the combustion chamber after the intake valve is closed, wherein the injecting includes supplying a pilot injection of fuel at a crank angle before a main injection of fuel; and

operating a hydraulic actuator to control movement of the air intake valve.

57. The method of claim 56, wherein at least a portion of the main injection occurs during a combustion stroke of the piston.

58. The method of claim 56, further including cooling the pressurized air prior to supplying the pressurized air to the air intake port.

59. The method of claim 56, wherein said supplying includes supplying a mixture of pressurized air and recirculated exhaust gas from the intake manifold to the air intake port, and wherein said operating the air intake valve includes operating the air intake valve to open the air intake port to allow the pressurized air and exhaust gas mixture to flow between the combustion chamber and the intake manifold substantially during a majority portion of the compression stroke of the piston.

60. The method of claim 56, wherein said supplying a mixture of pressurized air and recirculated exhaust gas includes controllably providing a quantity of exhaust gas from an exhaust gas recirculation (EGR) system.

61. The method of claim 56, wherein operating a hydraulic actuator includes moving a piston of the hydraulic actuator between a first position and a second position, engaging a rocker arm operatively coupled to the air intake valve, and selectively restricting movement of the piston when in the second position.

62. The method of claim 61, wherein the moving includes directing a fluid from a tank within the hydraulic actuator to a bore housing the piston, and selectively trapping the fluid within the bore.

63. The method of claim 62, wherein the moving further includes allowing the fluid to drain from the bore to the tank to allow the piston to move to the first position.

64. The method of claim 61, wherein the moving further includes moving the piston to slow a closing of the air intake valve.

65. The method of claim 56, further including selectively mechanically linking a cam to the air intake valve to move the air intake valve between a first position at which the pressurized air flows through the air intake port and a second position at which the air intake valve blocks the flow of pressurized air through the air intake port; and

controlling the hydraulic actuator to decouple the mechanical link between the cam and the air intake valve.