Low emissions compression ignited engine technology

A method and apparatus for operating a compression ignition engine having a cylinder wall, a piston, and a head defining a combustion chamber. The method and apparatus includes delivering fuel substantially uniformly into the combustion chamber, the fuel being dispersed throughout the combustion chamber and spaced from the cylinder wall, delivering an oxidant into the combustion chamber sufficient to support combustion at a first predetermined combustion duration, and delivering a diluent into the combustion chamber sufficient to change the first predetermined combustion duration to a second predetermined combustion duration different from the first predetermined combustion duration.

TECHNICAL FIELD

This invention relates generally to a method and apparatus for operating a compression ignition engine and, more particularly, to a method and apparatus for operating an engine in a homogeneous charge compression ignition mode to achieve low emissions during normal operating load conditions.

BACKGROUND

Internal combustion engines are used extensively for a variety of purposes. The transportation infrastructure relies almost exclusively on the use of engines to provide power for mobility. Electrical power generation also relies heavily on internal combustion engines.

The prolific use of engines in our society has created a number of issues, one of which is the ever-increasing amounts of combustion by-products being emitted. Although today's engines operate with much lower emission levels than previous generations of engines, the rapidly increasing numbers of engines being used creates the need to reduce emission levels even more.

Governments around the world recognize this problem and are taking regulatory steps to address the emission levels of engines. For example, levels of oxides of nitrogen (NOx), hydrocarbons (HC), carbon monoxide (CO), and smoke, among others, must be reduced drastically to meet evolving government standards.

Spark ignition engines, by the nature of their operation and the types of fuel used, tend to produce low levels of NOx and particulate emissions. Compression ignition engines, for example diesel engines, generally produce high levels of NOx and particulate emissions. Diesel engines, however, are still popular in use because they provide higher thermal efficiency than their spark-ignition counterparts, and thus offer higher power output for work applications.

Engines that operate in homogeneous charge compression ignition (HCCI) mode have generated much interest due to the potential to operate at high fuel efficiency while generating low combustion emissions. HCCI engines differ from conventional diesel compression ignition engines in that diesel engines ignite fuel that is rich, i.e., highly concentrated in an area in a combustion chamber, while HCCI techniques create a dispersed homogeneous fuel/air mixture by the time of combustion. Combustion of a homogeneous fuel/air mixture allows an engine to operate such that emission by-products are significantly reduced.

The theory of HCCI mode operation has not been met by the reality, however. It has proven to be extremely difficult to create a desired homogeneous mixture of fuel and air and still control operation of the engine. For example, it is very difficult to control the timing of combustion when introducing a homogeneous mixture into a combustion chamber. Past attempts by others has only resulted in partial success under low load, e.g., one half load or less, conditions. In U.S. Pat. No. 6,286,482, Flynn et al. recognize this issue and only operate an engine in PCCI mode (which is equivalent to HCCI) under low to medium load conditions. Operation switches to spark ignition mode at high loads. Yanagihara, in a paper entitled “Ignition Timing Control at Toyota ‘UNIBUS’ Combustion System”, limits engine operation to one half load to enable operation in HCCI mode.

SUMMARY OF THE INVENTION

In one aspect of the present invention a method for operating a compression ignition engine having a cylinder wall, a piston, and a head defining a combustion chamber is disclosed. The method includes the steps of delivering fuel substantially uniformly into the combustion chamber, the fuel being dispersed throughout the combustion chamber and spaced from the cylinder wall, delivering an oxidant into the combustion chamber sufficient to support combustion at a first predetermined combustion duration, and delivering a diluent into the combustion chamber sufficient to change the first predetermined combustion duration to a second predetermined combustion duration different from the first predetermined combustion duration.

In another aspect of the present invention a method for operating a compression ignition engine having a cylinder wall, a piston, and a head defining a combustion chamber is disclosed. The method includes the steps of delivering fuel substantially uniformly into the combustion chamber, the fuel being dispersed throughout the combustion chamber and spaced from the cylinder wall, delivering an oxidant into the combustion chamber sufficient to support combustion at a first predetermined pressure rise rate, and delivering a diluent into the combustion chamber sufficient to change the first predetermined pressure rise rate to a second predetermined pressure rise rate different from the first predetermined pressure rise rate.

In yet another aspect of the present invention a method for delivering fuel into a combustion chamber of a compression ignition engine, the combustion chamber being defined by a cylinder wall, a piston, and a head, is disclosed. The method includes the steps of delivering the fuel to a nozzle of an injector, the nozzle having a plurality of holes distributed in a desired pattern, and injecting the fuel through the nozzle holes into the combustion chamber in a predetermined spray pattern so that the fuel is dispersed throughout the combustion chamber and spaced from the cylinder wall.

In still another aspect of the present invention an apparatus for operating a compression ignition engine having a cylinder wall, a piston, and a head defining a combustion chamber is disclosed. The apparatus includes a fuel injector having a nozzle positioned to inject fuel in a dispersed pattern throughout the combustion chamber and spaced from the cylinder wall, and an air supply system for delivering at least one of an oxidant and a diluent into the combustion chamber.

DETAILED DESCRIPTION

Referring to the drawings and the specification, a method and apparatus100for operating a compression ignition engine102is disclosed.

Referring toFIG. 1, there is shown an engine assembly104. The engine assembly104depicts fundamental operation of a compression ignition engine102. Additional features of the engine assembly104ofFIG. 1, for example an exhaust gas recirculation assembly, are described below with reference to additional figures.

The engine assembly104includes a plenum member106, and an air source108. The plenum member106has an inlet opening112, and an exit opening110defined therein. The air source108supplies air to the inlet opening112. Air from the air source108advances into a plenum chamber114defined in the plenum member106via the inlet opening112. It is noted that the description pertaining toFIG. 1refers to air as being the medium being provided to the engine assembly104. However, as described below, any suitable fluid medium may be used, for example, recirculated exhaust gases combined with air, and the like.

The engine assembly104further includes a cylinder assembly116. The cylinder assembly116includes a block118having a cylinder119defined therein. The cylinder119is defined by a cylinder wall120. An engine head122is secured to the block118. The engine head122has an intake port124, an exhaust port126, and a fuel injector opening154defined therein. An intake conduit128places the intake port124in fluid communication with the exit opening110of the plenum member106. An exhaust passage146places the exhaust port126in fluid communication with an exhaust manifold148.

The engine assembly104further includes a piston130which translates in the cylinder119in the general direction of arrows132and136. As the piston130moves downwardly in the general direction of arrow136to the position shown inFIG. 1, a connecting rod134urges a crankshaft142to rotate in the general direction of arrow144. Subsequently, as the crankshaft142continues to rotate in the general direction of arrow144, the crankshaft142urges the connecting rod134and the piston130in the general direction of arrow132to return the piston130to the uppermost position (not shown).

The piston130, the cylinder wall120, and the engine head122cooperate so as to define a combustion chamber138. In particular, when the piston130is advanced in the general direction of arrow132, the volume of the combustion chamber138is decreased. On the other hand, when the piston130is advanced in the general direction of arrow136, the volume of the combustion chamber138is increased as shown inFIG. 1.

The engine assembly104further includes a fuel reservoir158. A fuel pump160draws low pressure fuel from the fuel reservoir158and advances high pressure fuel to a fuel injector156via a fuel line162. The fuel injector156is positioned in the injector opening154and is operable to inject a quantity of fuel into the combustion chamber138through the injector opening154. In particular, the fuel injector156injects fuel into the combustion chamber138upon receipt of an injector control signal on a signal line166. Furthermore, the fuel can be any one of the following group of fuels: diesel fuel, crude oil, lubricating oil, or an emulsion of water and diesel fuel. More generally, the fuel may be any type of fuel which has a high cetane number, thus having the property of combusting readily.

It should be appreciated that the amount of fuel injected by the fuel injector156controls the ratio of air to fuel, or air/fuel ratio, advanced to the combustion chamber138. Specifically, if it is desired to advance a leaner mixture to the combustion chamber138, a fuel control signal received via signal line166causes the fuel injector156to operate so as to inject less fuel to the combustion chamber138. On the other hand, if it is desired to advance a richer mixture of air and fuel to the combustion chamber138, a fuel control signal received via signal line166causes the fuel injector156to operate so as to advance more fuel to the combustion chamber138.

It is noted that other methods of introducing the fuel and air mixture to the combustion chamber138may be used without deviating from the spirit and scope of the present invention. For example, the fuel may be mixed with air at any point from the air source108through the intake conduit128, including upstream of a turbocharger (not shown).

An intake valve140selectively places the plenum chamber114in fluid communication with the combustion chamber138. The intake valve140may be actuated in a known manner by a camshaft (not shown), a pushrod (not shown), and a rocker arm (not shown) driven by rotation of the crankshaft142. Alternatively, the intake valve140may be actuated by other means, such as hydraulically, electronically, a combination of electro-hydraulically, and the like. When the intake valve140is placed in the open position (shown inFIG. 1), air is advanced from the intake conduit128to the combustion chamber138via the intake port124. When the intake valve140is placed in the closed position (not shown), air is prevented from advancing from the intake conduit128to the combustion chamber138since the intake valve140blocks fluid flow through the intake port124.

An exhaust valve152selectively places the exhaust manifold148in fluid communication with the combustion chamber138. The exhaust valve152may be actuated in a known manner by a camshaft (not shown), a pushrod (not shown), and a rocker arm (not shown) each of which are driven by the rotation of the crankshaft142. Alternatively, the exhaust valve152may be actuated by other means, such as hydraulically, electronically, a combination of electro-hydraulically, and the like. When the exhaust valve152is placed in the open position (not shown), exhaust gases are advanced from the combustion chamber138to the exhaust manifold148via a fluid path that includes the exhaust port126and the exhaust passage146. From the exhaust manifold148, exhaust gases are advanced to an exhaust conduit150. When the exhaust valve152is placed in the closed position (shown inFIG. 1), exhaust gases are prevented from advancing from the combustion chamber138to the exhaust manifold148since the exhaust valve152blocks fluid flow through the exhaust port126.

Combustion of the mixture of fuel and air in the combustion chamber138produces a number of exhaust gases. After the mixture of fuel and air is combusted in the combustion chamber138, exhaust gases are advanced through the exhaust conduit150. Included among the exhaust gases are quantities of oxides of nitrogen (NOx), hydrocarbons (HC), carbon monoxide (CO), smoke, and the like.

The engine assembly104further includes a controller164. The controller164is preferably a microprocessor-based engine control unit (ECU). The controller164may perform a variety of functions, including, as described above, controlling actuation of the fuel injector156.

Referring toFIG. 2, a schematic representation of an engine102having an intake conduit128and an exhaust passage146is shown. An engine block230provides housing for at least one cylinder119.FIG. 2depicts six cylinders119. However, any number of cylinders119could be used, for example, one, three, six, eight, ten, twelve, or any other number. The intake conduit128provides an intake path for each cylinder119for air, recirculated exhaust gases, or a combination thereof. The exhaust passage146provides an exhaust path for each cylinder119for exhaust gases.

In the embodiment shown inFIG. 2, a two-stage turbocharger system208is illustrated. The turbocharger system208includes a first turbocharger stage210having a low pressure turbine216and a first stage compressor218. The turbocharger system208also includes a second turbocharger stage212having a high pressure turbine214and a second stage compressor220. The two-stage turbocharger system208operates to increase the pressure of the air and exhaust gases being delivered to the cylinders119via the intake conduit128, and to maintain a desired air to fuel ratio during an extended open duration of an intake valve, which is described in more detail below. It is noted that a two-stage turbocharger system208is not required for operation of the present invention. Other types of turbocharger systems, such as a high pressure ratio single-stage turbocharger system, a variable geometry turbocharger system, and the like, may be used instead.

The engine assembly includes an exhaust system202, which in turn includes an exhaust gas recirculation (EGR) system204. The EGR system204shown inFIG. 2is typical of a low pressure EGR system in an internal combustion engine. Variations of the EGR system204shown may also be used with the present invention. Furthermore, other types of EGR systems, for example, by-pass, venturi, piston-pumped, peak clipping, and back pressure, could be used as well.

An oxidation catalyst222receives exhaust gases from the low pressure turbine216. The oxidation catalyst222may also be coupled with a De-NOxcatalyst to further reduce NOxemissions. A particulate matter (PM) filter206receives exhaust gases from the oxidation catalyst222. Although the oxidation catalyst222and the PM filter206are shown as separate items, they may alternatively be combined into one package.

Some of the exhaust gases are delivered out the exhaust from the PM filter206. However, a portion of exhaust gases are rerouted to the intake conduit128through an EGR cooler224, through an EGR valve226, and through the turbocharger system208.

FIG. 3shows a variation of the EGR system204ofFIG. 2. InFIG. 3, some of the exhaust gases are routed from the low pressure turbine216, through the oxidation catalyst222, and through the PM filter206. However, a portion of exhaust gases are rerouted to the intake conduit128from the low pressure turbine216, i.e., before entering the oxidation catalyst222, through an additional PM filter302, then through the EGR cooler224, EGR valve226, and the turbocharger system208. The additional PM filter302may be smaller in size than the PM filter206in the main exhaust stream since only a portion of the exhaust gases need be filtered. In addition, by installing the additional PM filter302in the return path of the EGR system204, the packaging and routing of the filter302and the associated input and output ductwork becomes more compact and manageable around the vicinity of the engine102.

Referring toFIGS. 4aand4b, operation of a fuel injector156suited for use with the present invention is shown. InFIG. 4a, the fuel injector nozzle154, i.e., the tip of the injector156, is shown in more detail. The fuel injector nozzle154includes a plurality of micro-sized holes401., e.g., 10, 16, 24, 32 and the like, arranged in a pattern such that a desired fuel spray402is achieved. The exemplary fuel injector opening154ofFIGS. 4aand4breflects a24hole “showerhead” design, arranged such that a first set of holes injects fuel spray at a first angle of dispersion α and a second set of holes injects fuel spray at a second angle of dispersion β. For example, a first set of 8 holes injects fuel spray at an angle α equal to about 50 degrees and a second set of 16 holes injects fuel spray at an angle β equal to about 90 degrees. It is noted that any number and combination of holes, sets of holes, and angles of dispersion may be used as well without deviating from the scope of the present invention.

The design of the fuel injector nozzle154ofFIGS. 4aand4boffers the advantage of distributing the fuel spray402uniformly throughout desired portions of the combustion chamber138, in particular with respect to a particular geometry of the piston130. This control over the fuel spray402allows for fuel injection in advance of normal injection timing to allow sufficient time for the fuel and air, i.e., fluid medium, to mix homogeneously without fuel being allowed to deposit on the cylinder wall120prior to combustion. Preferably, the fuel spray402is configured to inject the fuel such that the fuel is dispersed substantially uniformly into the combustion chamber138and spaced from the cylinder wall120. More specifically, the fuel spray402is intended to disperse throughout the combustion chamber138without any fuel contacting the cylinder wall120, thus preventing fuel from quenching on the cylinder wall120, which may be at a lower temperature than the remainder of the combustion chamber138, and thus may result in increased levels of HC and CO during combustion.

Alternative fuel injection techniques may be used with the present invention. For example,FIGS. 5a–5cillustrate the function of a fuel injector156suited for use in mixed-mode operations. More specifically, the fuel injector opening154includes at least one HCCI nozzle outlet504and at least one conventional nozzle outlet506. The HCCI nozzle outlet504is configured at an angle θ from a longitudinal axis502of the fuel injector156to inject fuel spray504in a pattern represented byFIG. 5b. The conventional nozzle outlet506is configured at an angle λ from the longitudinal axis502to inject fuel spray504in a pattern represented byFIG. 5c.

During HCCI mode operations, the fuel spray402is directed downward toward the piston130. Injection takes place more in advance of top dead center, as can be seen by the relative position of the piston130inFIG. 5bcompared toFIG. 5c, which allows more time for the fuel and fluid medium, e.g. air, to combine into a homogeneous mixture.

During conventional mode operations, e.g., diesel compression mode, the fuel spray402is directed more toward the sides of the cylinder119and injection takes place closer to top dead center, as evidenced by the position of the piston130inFIG. 5c.

It is noted that variations of the injector configuration ofFIG. 5amay be used without deviating from the scope of the present invention. For example, a showerhead type of output nozzle may be used in place of the HCCI nozzle outlet504for HCCI operations, while the conventional nozzle outlet506may be employed during conventional diesel operations. Furthermore, in the fuel injector156ofFIG. 5a, the angles θ and λ of the nozzle outlets504,506may be varied over a wide range to suit particular applications. Alternatively, the HCCI injection may be accomplished with port injection, i.e., the fuel is injected for example in the intake conduit128to provide for a homogeneous mixture of fuel and air in the combustion chamber138. This method, however, may result in fuel condensing on the cylinder wall120, thus contributing to oil degradation.

The timing of fuel injection may be varied to improve performance during HCCI operation. A timing range from about 50 degrees before top dead center (TDC) to about 180 degrees before TDC is typically used to insure a near complete homogeneous mixture of fuel and fluid medium. However, it is preferred to inject fuel as late as possible, i.e., closer to TDC, since excessive time for fuel presence in the cylinder119results in fuel condensing on the cylinder walls120, which in turn contaminates and degrades the engine oil. It has been shown, as depicted inFIGS. 6aand6b, that with a 24 hole showerhead fuel injector and no EGR an optimal injection timing of about 70 degrees before TDC may be achieved. More specifically, at about 70 degrees before TDC, levels of NOx and smoke are minimal and levels of HC and CO are greatly reduced. It has been further found that the addition of EGR may retard the optimal injection timing to about 60 degrees before TDC, thus alleviating fuel condensation even more. Further refinements in operating conditions, such as injector tip geometry, fuel dispersion patterns, EGR quantity, air intake, and the like, have enabled fuel injection in the range of from about 30 degrees before TDC to about 90 degrees before TDC, with optimal emissions reported when fuel injection occurs at about 40 degrees before TDC.

It is desired during HCCI operations to maintain a low combustion temperature. One reason is that levels of NOx are reduced at low combustion temperatures. One method for achieving low combustion temperatures is to introduce a high level of excess mass, i.e., large amounts of a fluid medium such as air, EGR, water, inert gas and the like, into the combustion chamber138. Using air, i.e., fresh air, as the excess mass medium requires very large amounts of air to be delivered to the combustion chamber138to achieve desired excess mass levels. For example, an air to fuel ratio of about 36 to 1 or greater may be desired, corresponding to an equivalence ratio of 0.4 or less.

Alternatively, some other type of fluid medium may be used to achieve excess mass. For example, the use of EGR in place of at least a portion of fresh air may enable operation of the engine102at a near stoichiometric equivalence ratio, i.e., with an air to fuel ratio of about 14.5 to 1.

EGR may also be used to control a heat release rate and a pressure rise rate within the combustion chamber138. For example, as the graph702shown inFIG. 7depicts, a first plot704is indicative of a pressure rise rate during combustion in HCCI mode without the addition of EGR. The plot704illustrates a sharp rise in pressure in the combustion chamber138. This sharp rise in pressure creates stresses in components such as the engine head122. A second plot706is indicative of the pressure rise rate with EGR added. First, it is noted that the duration of combustion, i.e., the time for combustion to take place, has changed. More specifically, the combustion duration is extended. Second, it is noted that the peak pressure has changed. More specifically, the peak pressure is reduced. It has been found that the addition of EGR enables brake mean effective pressure (BMEP) levels to approach 1600 kPa. Without EGR, BMEP is limited to about 1100 kPa, i.e., about one half load.

The fluid added does not necessarily have to be EGR. More generally, the addition of a diluent such as EGR, water, carbon dioxide, nitrogen, and the like performs the function of lowering combustion temperature, limiting peak combustion pressure, and extending the duration of combustion. The diluent affects combustion by lowering the heat release rate in the combustion chamber138and creating a number of interim chemical reactions during combustion which serves to extend the combustion event. It is noted that the mass of the diluent contributes to the total fluid mass in the combustion chamber138, the other portion of fluid mass being the oxidant, e.g., air, introduced to support combustion.

Referring toFIG. 8, the amount of EGR added is preferably quantified as a volumetric percentage, as exemplified by the following equation:

%⁢⁢EGR=CO2⁢⁢(in)CO2⁢⁢(ex)×100(Eq.⁢1)
where CO2(in) is an amount of carbon dioxide being returned to the engine by way of the EGR system204, and CO2(ex) is an amount of carbon dioxide exhausted from the engine102. The amount of EGR may be a significant percentage, for example 40% to 60%, under certain operating conditions. It is noted that the percentage of EGR may be quantified by some other method such as, for example, the mass of the EGR divided by the total mass in the combustion chamber138.

Referring toFIG. 9a, a graph902of cylinder pressure vs. crank angle degrees (CAD) is shown. The plot indicates a first pressure rise portion906having a rise slope which levels off, then increases in slope to a second pressure rise portion908. The “double-humped” curve is indicative of a homogeneous mixture during combustion, and thus defines an HCCI mode. In like manner, inFIG. 9b, a graph904of heat release rate vs. CAD is shown. The plot includes a first heat release peak portion910, followed by a second heat release peak portion912. As noted, the second heat release peak portion912is much larger in value than the first heat release peak portion910. The curve serves to define an HCCI mode as well.

The excess mass may be provided by the use of high boost pressure at the intake conduit128, i.e., intake manifold, of the engine102. Exemplary techniques for providing high boost pressure are described below.

Although the introduction of excess mass serves to control the pressure rise rate in the combustion chamber138, it is also desired to control a peak pressure during combustion. AsFIG. 7illustrates, the first plot704has a peak pressure that is higher in value than the peak pressure of the second plot706. One method of controlling the peak pressure is by use of a variable compression ratio (VCR).

There are many techniques in use which provide VCR of an engine. One common strategy is to employ variable valve timing, in particular variable intake valve timing. For example, an intake valve may be kept open for a period of time into a compression cycle, for example from about 20 to about 50 degrees into compression. Variable valve timing may be accomplished by several means. Exemplary techniques may include mechanical, e.g., control of cam actuation, hydraulic, electric, electro-hydraulic, and the like.

Another common strategy, and one that may be more effective than variable valve timing, is to vary the geometric characteristics of a cylinder. For example, as depicted inFIG. 10, a secondary cylinder1002may be used in cooperation with a secondary piston1004to vary the effective volume of the cylinder119. A rod1006connected to the secondary piston1004is also connected to an actuator1008, such as a cam actuator, a hydraulic actuator, a solenoid actuator, or other actuation device. As the position of the secondary piston1004is varied in the secondary cylinder1002, the effective compression ratio of the piston130and cylinder119is varied. It is noted that the example ofFIG. 10is but one of many methods by which the compression ratio of a cylinder may be varied using geometric techniques.

Preferably, to enable combustion to occur at a desired time, the VCR is varied as a function of engine speed and engine load. Typically, as speed and load increases, more fuel is delivered to the combustion chamber138. This additional fuel causes an increase in pressure. The VCR may be lowered as speed and load increases to help compensate for this pressure increase. An exemplary range for compression ratio may be from about 8:1 to about 16:1. For example, a compression ratio of 10:1 was used in a test engine running at about 75% load. Preferably, compression ignition rather than spark ignition is maintained during the above-referenced lower compression ratios.

Although the engine102may be operating in HCCI mode and may be using a fuel such as diesel, the addition of EGR as described above, for example about 40% to about 60% EGR, enables operation at near stoichiometric. Under these conditions, it is possible to use a 3-way catalyst for further reductions in HC, CO and NOx. For example, referring toFIG. 11, an exemplary 3-way catalyst suited for use with the present invention is shown.

FIG. 11shows a series combination catalytic converter1110in which three different catalytic substrates1133,1134and1135are mounted in series within an individual tubular housing1122. The inner structure includes mounting each of the catalytic substrates in its own sub-can1130,1131and1132, respectively. Tubular housing1122may be formed from thin stainless steel and may be formed on the outlet end1124with an annular retaining lip1125that prevents the individual sub-cans from escaping through the outlet. In addition, the curvature of the bend which creates retaining lip1125may be useful as a guide when mounting the converter1110in an opening having a diameter very close to that of the housing1122. Each of the sub-cans1130,1131and1132may be held within the tubular housing1122by a peripheral seam weld at corners1128.

The substrate1133may be coated with a typical deNOx catalyst, such as a combination precious metal and zeolite catalyst. The substrate1134may be coated with a catalyst appropriate to target secondary undesirable nitrogen compounds existing in the exhaust after exiting the substrate1133. After emerging from the substrate1134, the exhaust contains very low levels of NOx compounds and even less undesirable secondary nitrogen compounds which would otherwise become NOx compounds after proceeding through an oxidation catalyst. The substrate1135may be coated with an oxidation catalyst to promote the conversion of any existing HC and CO into carbon dioxide and water. Only small amounts of the exhaust are turned back into undesirable NOx compounds after passing through oxidation catalyst substrate1135. Upon exiting the converter1110at the outlet1124, the exhaust has acceptable levels of both HC and NOx.

The sub-cans1130,1131and1132are preferably made from relatively thin stainless steel that is rolled on both ends to create an annular retaining lip that traps the individual ceramic substrates1133,1134and1135within the sub-cans. A matting material1129may be mounted between the inner surface of each sub-can and the outer surface of each substrate. The edges of the individual strips of matting1129may be shielded from the corrosive effects of the exhaust by end rings1127. Each of the sub-cans may be fixed within the tubular housing1122via a peripheral seam weld at the annular corners1128.

It is noted that the above example of a 3-way catalyst is for exemplary purposes only, and that variations of the above catalyst may be used as well. Furthermore, other types of catalysts, e.g., deNOx catalysts only, oxidation catalysts only, and the like, may be used as well.

The large amounts of excess mass, e.g., EGR, will require a significant level of boost pressure, i.e., intake manifold pressure, to deliver the excess mass into the combustion chamber138. For example, a boost pressure value of about 4.5 to 1 or higher may be required under full load operating conditions. That is, the pressure at the intake manifold will need to be at least 456 kPa. The achievement of this high boost pressure requires an air system capable of generating sufficient pressure. For example, the 2-stage turbocharger system208ofFIGS. 2 and 3illustrates one possible air system capable of generating sufficient boost pressure.

Referring toFIGS. 12–14, exemplary variations of the turbocharger system208are shown. Discussion of the components ofFIGS. 12–14is provided below, with new element labeling to provide further clarification of various air systems.

Referring toFIG. 12, an exemplary air supply system1202for an internal combustion engine1204, for example, a four-stroke, diesel engine, is provided. The internal combustion engine1204includes an engine block1206defining a plurality of combustion cylinders1208, the number of which depends upon the particular application. For example, a 4-cylinder engine would include four combustion cylinders, a 6-cylinder engine would include six combustion cylinders, etc. In the exemplary embodiment ofFIG. 12, six combustion cylinders1208are shown.

The internal combustion engine1204also includes an intake manifold1210and an exhaust manifold1212. The intake manifold1210provides fluid, for example, air or a fuel/air mixture, to the combustion cylinders1208. The exhaust manifold1212receives exhaust fluid, for example, exhaust gas, from the combustion cylinders1208. The intake manifold1210and the exhaust manifold1212are shown as a single-part construction for simplicity in the drawing. However, it should be appreciated that the intake manifold1210and/or the exhaust manifold1212may be constructed as multi-part manifolds, depending upon the particular application.

The air supply system1202includes a first turbocharger1214and may include a second turbocharger1216. The first and second turbochargers1214,1216may be arranged in series with one another such that the second turbocharger1216provides a first stage of pressurization and the first turbocharger1214provides a second stage of pressurization. For example, the second turbocharger1216may be a low pressure turbocharger and the first turbocharger1214may be a high pressure turbocharger. The first turbocharger1214includes a turbine1218and a compressor1220. The turbine1218is fluidly connected to the exhaust manifold1212via an exhaust duct1222. The turbine1218includes a turbine wheel1224carried by a shaft1226, which in turn may be rotatably carried by a housing1228, for example, a single-part or multi-part housing. The fluid flow path from the exhaust manifold1212to the turbine1218may include a variable nozzle (not shown) or other variable geometry arrangement adapted to control the velocity of exhaust fluid impinging on the turbine wheel1224.

The compressor1220includes a compressor wheel1230carried by the shaft1226. Thus, rotation of the shaft1226by the turbine wheel1224in turn may cause rotation of the compressor wheel1230.

The first turbocharger1214may include a compressed air duct1232for receiving compressed air from the second turbocharger1216and an air outlet line1234for receiving compressed air from the compressor1220and supplying the compressed air to the intake manifold1210of the engine1204. The first turbocharger1214may also include an exhaust duct1236for receiving exhaust fluid from the turbine1218and supplying the exhaust fluid to the second turbocharger1216.

The second turbocharger1216may include a turbine1238and a compressor1240. The turbine1238may be fluidly connected to the exhaust duct1236. The turbine1238may include a turbine wheel1242carried by a shaft1244, which in turn may be rotatably carried by the housing1228. The compressor1240may include a compressor wheel1246carried by the shaft1244. Thus, rotation of the shaft1244by the turbine wheel1242may in turn cause rotation of the compressor wheel1246.

The second turbocharger1216may include an air intake line1248providing fluid communication between the atmosphere and the compressor1240. The second turbocharger1216may also supply compressed air to the first turbocharger1214via the compressed air duct1232. The second turbocharger1216may include an exhaust outlet1250for receiving exhaust fluid from the turbine1238and providing fluid communication with the atmosphere. In an embodiment, the first turbocharger1214and second turbocharger1216may be sized to provide substantially similar compression ratios. For example, the first turbocharger1214and second turbocharger1216may both provide compression ratios of between 2 to 1 and 3 to 1, resulting in a system compression ratio of at least 4:1 with respect to atmospheric pressure. Alternatively, the second turbocharger1216may provide a compression ratio of 3 to 1 and the first turbocharger1214may provide a compression ratio of 1.5 to 1, resulting in a system compression ratio of 4.5 to 1 with respect to atmospheric pressure.

The air supply system1202may include an air cooler1252, for example, an aftercooler, between the compressor1220and the intake manifold1210. The air cooler1252may extract heat from the air to lower the intake manifold temperature and increase the air density. Optionally, the air supply system1202may include an additional air cooler1254, for example, an intercooler, between the compressor1240of the second turbocharger1216and the compressor1220of the first turbocharger1214. Alternatively, the air supply system1202may optionally include an additional air cooler (not shown) between the air cooler1252and the intake manifold1210. The optional additional air cooler may further reduce the intake manifold temperature.

FIG. 13is a block diagram illustrating another exemplary air supply system1302for the internal combustion engine1204. The air supply system1302may include a turbocharger1304, for example, a high-efficiency turbocharger capable of producing at least about a 4.5 to 1 compression ratio with respect to atmospheric pressure. The turbocharger1304may include a turbine1306and a compressor1308. The turbine1306may be fluidly connected to the exhaust manifold1212via an exhaust duct1310. The turbine1306may include a turbine wheel1312carried by a shaft1314, which in turn may be rotatably carried by a housing1316, for example, a single-part or multi-part housing. The fluid flow path from the exhaust manifold1212to the turbine1306may include a variable nozzle (not shown), which may control the velocity of exhaust fluid impinging on the turbine wheel1312.

The compressor1308may include a compressor wheel1318carried by the shaft1314. Thus, rotation of the shaft1314by the turbine wheel1312in turn may cause rotation of the compressor wheel1318. The turbocharger1304may include an air inlet1320providing fluid communication between the atmosphere and the compressor1308and an air outlet1322for supplying compressed air to the intake manifold1210of the engine1204. The turbocharger1304may also include an exhaust outlet1324for receiving exhaust fluid from the turbine1306and providing fluid communication with the atmosphere.

The air supply system1302may include an air cooler1326between the compressor1308and the intake manifold1210. Optionally, the air supply system1302may include an additional air cooler (not shown) between the air cooler1326and the intake manifold1210.

FIG. 14is a block diagram illustrating another exemplary air supply system1402for the internal combustion engine1204. The air supply system1402may include a turbocharger1404, for example, a turbocharger1404having a turbine1406and two compressors1408,1410. The turbine1406may be fluidly connected to the exhaust manifold1212via an inlet duct1412. The turbine1406may include a turbine wheel1414carried by a shaft1416, which in turn may be rotatably carried by a housing1418, for example, a single-part or multi-part housing. The fluid flow path from the exhaust manifold1212to the turbine1406may include a variable nozzle (not shown), which may control the velocity of exhaust fluid impinging on the turbine wheel1414.

The first compressor1408may include a compressor wheel1420carried by the shaft1416, and the second compressor1410may include a compressor wheel1422carried by the shaft1416. Thus, rotation of the shaft1416by the turbine wheel1414in turn may cause rotation of the first and second compressor wheels1420,1422. The first and second compressors1408,1410may provide first and second stages of pressurization, respectively.

The turbocharger1404may include an air intake line1424providing fluid communication between the atmosphere and the first compressor1408and a compressed air duct1426for receiving compressed air from the first compressor1408and supplying the compressed air to the second compressor1410. The turbocharger1404may include an air outlet line1428for supplying compressed air from the second compressor1410to the intake manifold1210of the engine1204. The turbocharger1404may also include an exhaust outlet1430for receiving exhaust fluid from the turbine1406and providing fluid communication with the atmosphere.

For example, the first compressor1408and second compressor1410may both provide compression ratios of between 2 to 1 and 3 to 1, resulting in a system compression ratio of at least 4:1 with respect to atmospheric pressure. Alternatively, the second compressor1410may provide a compression ratio of 3 to 1 and the first compressor1408may provide a compression ratio of 1.5 to 1, resulting in a system compression ratio of 4.5 to 1 with respect to atmospheric pressure.

The air supply system1402may include an air cooler1432between the second compressor1410and the intake manifold1210. Optionally, the air supply system1402may include an additional air cooler1434between the first compressor1408and the second compressor1410of the turbocharger1404. Alternatively, the air supply system1402may optionally include an additional air cooler (not shown) between the air cooler1432and the intake manifold1210.

It is noted that other types of air supply systems could be used as well. For example, an air-to-EGR cooler, a blower and turbocharger arrangement, and an electric turbocharger assist are a few of the types of air supply systems which may provide the needed boost pressure for the present invention.

Referring toFIG. 15, a block diagram illustrating an embodiment of a control system for the present invention is shown. The engine102is monitored and controlled by the controller164, e.g., an electronic control module (ECM) typically used for engine monitoring and control.

A signal indicative of cylinder pressure feedback is delivered to the controller164by way of signal line1502and may be used to determine an event such as a start of combustion. The cylinder pressure feedback may be sensed directly, for example by a cylinder pressure sensor (not shown), or may be derived from other sensed parameters. For example, engine speed and load parameters may be monitored and used to determine a start of combustion event.

The controller164may, upon receipt of the cylinder pressure feedback signal, determine that some control of engine operations is needed. For example, it may be determined that the timing of the start of combustion should be changed. The controller164may have several options to use for controlling engine operations. For example, the controller164nay deliver a control signal via signal line1504to modulate an intake manifold temperature, the controller164may deliver a control signal via signal line1506to modulate a timing of actuation of an intake valve, a control signal may be delivered via signal line1508to modulate a rate at which EGR is being delivered, a control signal may be delivered via signal line1510to modulate a timing of injection of fuel, or a control signal may be delivered, via signal line1512to modulate a boost pressure value. It is understood that any combination of the above control strategies may be employed. Furthermore, other control strategies may be incorporated as well.

The complexities of engine operation due to the interactions of many variables indicates that it may be desired to configure the controller164to use advanced techniques for data analysis and engine control. For example, it may be desired to incorporate a neural network (not shown) into the controller164to make control decisions based on an historical database of engine operations.

Referring toFIG. 16, the block diagram ofFIG. 2is reproduced with the addition of an oxygen sensor (O2)1602and a mass airflow sensor (MAF)1604. The O2 sensor1602may be located at some position suitable for sensing an amount of oxygen in the exhaust gases after combustion, for example at the exhaust passage146. The MAF sensor1604may be located at some position suitable for sensing the mass of EGR gases, for example prior to the EGR valve226. Alternatively, the MAF sensor1604may be located elsewhere, for example after the EGR valve226to sense a total flow of mass, e.g., EGR plus fresh air, being delivered to the engine102.

The O2 and MAF sensors1602,1604may be used separately or in combination, and may deliver sensed values to the controller164for processing to further determine and control a rate of EGR being delivered to the engine102.

In an alternate embodiment, it may be desired to incorporate membrane technology to use nitrogen as an inert gas in place of, or in combination with, EGR as the excess mass used to control heat release rates in the combustion chamber138. For example,FIG. 17depicts an exemplary intake air separation system1702suited for use with the present invention.

Referring toFIG. 17, a diagrammatic illustration of an intake air separation system1702for an engine1704is shown. The intake side of the engine1704includes an intake air conduit1706, an intake manifold1708, intake air pressurizing device1710, e.g., a turbocharger, and an intercooler or an air-to-air aftercooler1716. The intake air pressurizing device1710may include an exhaust gas driven turbine1714, which in turn drives a compressor1712. The engine1704also includes a main combustion section1720, and an exhaust system1724. Although not shown in great detail, the typical main combustion section1720includes, among other elements, an engine block and a cylinder head forming a plurality of combustion cylinders1722therein. Associated with each of the cylinders1722is a fuel injector, a cylinder liner, at least one air intake port and corresponding intake valves, at least one exhaust gas port and corresponding exhaust valves, and a reciprocating piston moveable within each cylinder to define, in conjunction with the cylinder liner and cylinder head, the combustion chamber. The exhaust system1724of the engine1704includes an exhaust manifold1726or split exhaust manifolds, one or more exhaust conduits1728, and the turbine1714. Optionally, the exhaust system1724may include one or more aftertreatment devices (not shown) such as particulate traps, NOx adsorbers, oxidation and/or lean NOx catalysts, or other recent advances in exhaust gas aftertreatment. Finally, the engine1704includes an electronic control module (ECM)1730, i.e., a controller, for operatively controlling the fuel injection timing and air system valve operations in response to one or more measured or sensed engine operating parameters, used as inputs to the ECM1730.

The intake air conduit1706is in flow communication with intake air input1732, the compressor1712of the intake air pressurizing device1710, and the aftercooler1716. Although the intake air separation system1702is shown and described in conjunction with a conventional turbocharged diesel engine, the disclosed system1702is equally useful on engines with a variable geometry turbocharger (VGT) or other supercharged engines, including engines with pressure wave supercharging devices. The intake manifold1708is connected to an end of the intake air conduit1706. An inlet pressure sensor1718is shown located somewhere in the intake air separation system1702, e.g., shown proximate the intake manifold1708, and provides intake air pressure data to the ECM1730. Other sensors such as temperature sensors, oxygen sensors (not shown), and the like may also be incorporated within the intake air separation system1702and likewise coupled as inputs to the ECM1730. In addition, various other devices such as filters, valves, actuators, bypass conduits, etc., although not shown, may also be incorporated within the intake air separation system1702. Any such electronically operative components such as valves and/or actuators are preferably operatively coupled to the ECM1730and operate in response to selected engine operating parameters or conditions, including engine speed, engine load, boost pressure conditions, etc.

The illustrated intake air separation system1702includes an intake air separation device1734disposed within the intake air separation system1702of the engine1704. The intake air separation device1734may be adapted for receiving substantially all of the engine combustion air at an air separation device inlet1736, i.e., an intake air inlet, and separating the same into a flow1738of oxygen enriched air, i.e., a permeate flow, and a flow1740of nitrogen enriched air, i.e., a retentate flow. The illustrated intake air separation device1734includes two inlets and two outlets. The first inlet is the intake air inlet1736that receives the air to be separated into an oxygen rich stream and a nitrogen rich stream. The second inlet is a purge air inlet1742that is adapted to receive a flow1744of sweep air or purge air which enhances the permeation effectiveness of the intake air separation device1734. The purge air1744maybe taken from a flow of intake air1758from the compressor1712and the aftercooler1716. Alternatively, the flow of purge air1744may be a separate flow of filtered ambient air. The first outlet, or permeate outlet1746of the intake air separation device1734is adapted to receive the permeate flow1738of oxygen enriched air combined with the flow of purge air1744.

The second outlet, or retentate outlet1748is adapted to receive the retentate flow1740of nitrogen enriched air. Preferably, the intake air separation device1734is a full flow separation unit and thus there is no need for subsequent mixing of the nitrogen enriched air flow1740exiting the retentate outlet1748with more intake air. The retentate outlet1748is further in flow communication with the intake manifold1708of the engine1704. A permeate flow valve1750may be disposed proximate the permeate outlet1746. The permeate flow valve1750is preferably actuated in response to signals received from ECM1730which controls the permeate flow1738away from the intake air separation device1734, and thereby controls the flow1740from the retentate outlet1748to the intake manifold1708. More specifically, the permeate flow valve1750located proximate the permeate outlet1746controls both the permeate flow1738and the flow of purge air1744away from intake air separation device1734and thus controls the relative concentrations of nitrogen and oxygen in the air directed to the intake manifold1708and to the combustion cylinders1722.

The location of the permeate flow valve1750is preferably at or proximate to the permeate outlet1746. Such an arrangement aids the responsiveness of the engine1704based on a relatively fast change in oxygen and nitrogen content of the air exiting the retentate outlet1748into the intake manifold1708when the permeate flow valve1750is actuated, e.g., opened or closed, during transient operating conditions. Selective operation of the permeate flow valve1750allows the engine1704to operate in essentially three different charge air modes, namely nitrogen enriched mode, i.e., valve partially or fully open, standard intake air mode, i.e., valve closed for selected length of time,.and transient oxygen enriched mode, which occurs for a short period or duration as the permeate flow valve1750is first closed. The exact location of the permeate flow valve1750is preferably optimized to take advantage of the different modes of charge air, and in particular the transient charge of oxygen enriched air that occurs when the permeate flow valve1750is first closed.

The intake air separation device1734preferably uses a plurality of selectively permeable separation membranes1754that separates ambient intake air into streams of oxygen enriched air and nitrogen enriched air. Such membranes1754are well known in the art.

The intake air separation device1734may include a housing or shell1756, having the intake air inlet1736, the purge air inlet1742, the permeate outlet1746, and the retentate outlet1748. A plurality of selectively permeable membrane elements or fibers are disposed in a general longitudinal or helical, i.e., spiral, orientation within the housing1756and potted or sealed at each end. The air separation membranes1754are preferably hollow, porous, coated tubes through which selected gases such as hydrogen, helium, water vapors, carbon dioxide, and oxygen tend to permeate outwardly through the membranes at a relatively fast rate while other gases such as carbon monoxide, argon and nitrogen permeate less rapidly and are mostly retained and transported along the membrane tubes. Different gases present in the flow1758of intake air tend to permeate through the membrane1754at different relative permeation rates and generally through the sidewalls of the membrane1754. The rate of permeation is also dependent in part on the membrane temperature, and therefore altering or controlling the temperatures of gases entering the intake air separation device1734ultimately controls permeability.

The intake air is introduced into the housing1756and membranes1754of the intake air separation device1734in an orientation or direction that is generally along the length of the membranes1754. In this manner the flow1758of intake air is transported or flows generally along the length of the intake air separation device1734. Conversely, the flow1744of purge air is introduced into the housing1756and membranes1754in a cross flow orientation or direction such that the flow1744of purge air flows generally across outer surfaces of the membranes1754. The flow1744of puree air then exits the housing1756via the permeate outlet1746as part of the permeate flow1738and together with the permeated oxygen rich air. The retentate flow1740of nitrogen rich air exits from the housing1756via retentate outlet1748.

The above description of an intake air separation device1734illustrate only one example of sweep or purge air flow configurations that produce good separation results. Various other flow configurations can also be employed. The various purge flow configurations offer differences in separation performance and packaging issues and can be tailored to the specific application in which the air separation device is used.

The compressor1712of the intake air pressurizing device1710is used to forcibly move intake air through the membrane-based intake air separation device1734in what is often referred to as the pressure mode. In like manner, the flow1744of purge air is received or diverted from the flow1758of boosted, cooled intake air and delivered to the purge air inlet1742. A purge air valve1752operatively coupled to the ECM1730may be used to control the flow1744of purge air under various operating conditions. Thus, the flow1744of purge air and the flow of intake air1758are typically pressurized while the permeate flow1738of oxygen enriched air and purge air exiting the intake air separation device1734is preferably at a somewhat lower pressure, due to pressure losses incurred by flowing through the intake air separation device1734. This pressure gradient across the membranes1754enables air separation to occur. As illustrated, the permeate flow1738is preferably vented to the atmosphere or otherwise fed to other parts of the engine1704, including, but not limited to the exhaust system1724. However, the permeate flow1738may also be delivered to the combustion cylinders1722to provide at least a portion of a supply of oxidant to support combustion. The retentate flow1740of nitrogen enriched air is fed to the intake manifold1708in a generally pressurized condition, albeit at a lower pressure than the feed or intake air pressure due to losses caused by the membrane-based intake air separation device1734.

Referring briefly toFIG. 1, it may be desired to utilize variable valve timing to aid in performance of the present invention. For example, the temperature in the cylinder119may be increased, thus assisting in control of combustion, by varying the timing of the exhaust valve152. More specifically, by varying the timing of closing of the exhaust valve152, some of the hot residual gases from combustion are trapped in the combustion chamber138and the start of combustion for the next cycle is advanced.

It may also be desired to vary the timing of the intake valve140to modulate the air to fuel ratio during acceleration of the engine102, thus further controlling combustion. Varying the timing of closing of the intake valve140serves to operate the engine102in a Miller cycle which lowers the effective compression ratio which in turn retards the start of combustion. Varying the timing of opening of the intake valve140allows hot exhaust gases to flow into the intake port124, which advances the start of combustion.

FIGS. 18 and 19illustrate operation of an exemplary technique for achieving variable valve timing. Although the description below andFIGS. 18and19depict variable valve operation of the intake valve, similar principles apply to varying the timing of an exhaust valve.

Referring toFIG. 18, a diagrammatic and cross-sectional illustration of a portion of an internal combustion engine1802is shown. A cylinder head1804is connected to an engine block1806. The cylinder head1804houses one or more cylinders1808. For purposes of illustration,FIG. 18is described below with reference to one cylinder1808.

The cylinder1808contains a piston1810slidably movable in the cylinder1808. A crankshaft1812is rotatably disposed within the engine block1806. A connecting rod1814couples the piston1810to the crankshaft1812so that sliding motion of the piston1810within the cylinder1808results in rotation of the crankshaft1812. Similarly, rotation of the crankshaft1812results in a sliding motion of the piston1810. For example, an uppermost position of the piston1810in the cylinder1808corresponds to a top dead center position of the crankshaft1812, and a lowermost position of the piston1810in the cylinder1808corresponds to a bottom dead center position of the crankshaft1812.

As one skilled in the art will recognize, the piston1810in a conventional, four-stroke engine cycle reciprocates between the uppermost position and the lowermost position during a combustion (or expansion) stroke, an exhaust stroke, an intake stroke, and a compression stroke. Meanwhile, the crankshaft1812rotates 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. Then, the four-stroke cycle begins again. Each piston stroke 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°.

The cylinder1808includes at least one intake port1816and at least one exhaust port1818, each opening to a combustion chamber1820. The intake port1816is coupled to an intake passageway1822and the exhaust port1818is coupled to an exhaust passageway1824. The intake port1816is opened and closed by an intake valve assembly1826, and the exhaust port1818is opened and closed by an exhaust valve assembly1828. The intake valve assembly1826includes, for example, an intake valve1830having ahead1832at a first end1834, with the head1832being sized and arranged to selectively close the intake port1816. A second end1836of the intake valve1830is connected to a rocker arm1838or any other conventional valve-actuating mechanism. The intake valve1830is movable between a first position permitting flow from the intake port1816to enter the cylinder1808and a second position substantially blocking flow from the intake port1816to the cylinder1808. Preferably, a spring1840is disposed about the intake valve1830to bias the intake valve1830to the second, closed position.

A camshaft1842carrying a cam1844with one or more lobes1846is arranged to operate the intake valve assembly1826cyclically based on the configuration of the cam1844, the lobes1846, and the rotation of the camshaft1842to achieve a desired intake valve timing. The exhaust valve assembly1828is configured in a manner similar to the intake valve assembly1826and is preferably operated by one of the lobes1846of the cam1844. In one embodiment, the intake lobe1846is configured to operate the intake valve1830in a conventional Otto or diesel cycle, whereby the intake valve1830moves to the second, 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. Alternatively, the intake valve assembly1826and/or the exhaust valve assembly1828may be operated hydraulically, pneumatically, electronically, or by any combination of mechanics, hydraulics, pneumatics, and/or electronics.

In the preferred embodiment, the intake valve assembly1826includes a variable intake valve closing mechanism1848structured and arranged to selectively interrupt cyclical movement of and extend the closing timing of the intake valve1830. The variable intake valve closing mechanism1848may be operated hydraulically, pneumatically, electronically, mechanically, or any combination thereof. For example, the variable intake valve closing mechanism1848may be selectively operated to supply hydraulic fluid, for example, at a low pressure or a high pressure, in a manner to resist closing of the intake valve1830by the bias of the spring1840. That is, after the intake valve1830is lifted, i.e., opened, by the cam1844, and when the cam1844is no longer holding the intake valve1830open, the hydraulic fluid may hold the intake valve1830open for a desired period. The desired period may change depending on the desired performance of the engine1802. Thus, the variable intake valve closing mechanism1848enables the engine1802to operate under a conventional Otto or diesel cycle or under a variable late-closing Miller cycle. In alternative embodiments, the intake valve1830may be controlled by a camless system (not shown), such as an electrohydraulic system, as is well known in the art.

As shown inFIG. 19, the intake valve1830may begin to open at about 360° crank angle, that is, when the crankshaft1812is at or near a top dead center position of an intake stroke1906. The closing of the intake valve1830may be selectively varied from about 540° crank angle, that is, when the crankshaft1812is at or near a bottom dead center position of a compression stroke1907, to about 650° crank angle, that is, about 70° before top center of the combustion stroke. Thus, the intake valve1830may be held open for a majority portion of the compression stroke1907, that is, for the first half of the compression stroke1907and a portion of the second half of the compression stroke1907.

A controller1850, e.g., an electronic control module (ECM), may be electrically connected to the variable intake valve closing mechanism1848. Preferably, the controller1850is configured to control operation of the variable intake valve closing mechanism1848based on one or more engine conditions, for example, engine speed, load, pressure, and/or temperature in order to achieve a desired engine performance. It should be appreciated that the functions of the controller1850may be performed by a single controller or by a plurality of controllers.

Referring back toFIG. 1, it is noted that, under some operating conditions such as engine start-up and light load operation, it may be desired to operate the engine102using a spark ignition system (not shown), as is well known in the art.

INDUSTRIAL APPLICABILITY

As an example of an application of the present invention, reference is made toFIG. 20in which a flow diagram illustrating a method for operating a compression ignition engine102having a cylinder wall120, a piston130, and a head122defining a combustion chamber138is shown.

In a first control block2002, fuel is delivered to the combustion chamber138so that the fuel is dispersed substantially uniformly throughout the combustion chamber138and is spaced from the cylinder wall120. More particularly, the fuel is dispersed throughout the combustion chamber138to provide a substantially homogeneous distribution, yet the fuel dispersion is controlled such that fuel does not impinge on the cylinder wall120, which would result in fuel condensation and subsequent degradation of the lubricating oil in the engine102.

In a second control block2004, sufficient oxidant is delivered to the combustion chamber138to support combustion at a first predetermined combustion duration. Typically, the oxidant includes a supply of fresh air, as is well known in the art. However, the oxidant could be at least in part a supply of oxygen obtained from such means as use of membrane technology, as described above.

In a third control block2006, a supply of diluent is delivered to the combustion chamber138sufficient to change the first predetermined combustion duration to a second predetermined combustion duration. Preferably, the second predetermined combustion duration differs from the first predetermined combustion duration. For example, the second predetermined combustion duration may be greater than the first predetermined combustion duration so that combustion is controlled over a longer period of time.

The diluent may be EGR, air, an inert gas such as nitrogen, and the like. For example, as described above, the diluent may be a gas which includes up to 40–60% EGR. As another example, the diluent may include a quantity of nitrogen obtained by means such as membrane technology, as described above. The diluent may also include a combination of gases.

In combination with the combustion duration being changed by the addition of diluent, the diluent may also serve to change a first predetermined pressure rise rate in the combustion chamber138to a second predetermined pressure rise rate. For example, the pressure rise rate during combustion may decrease from the addition of diluent. As described above,FIG. 7serves to illustrate the change in both combustion duration and combustion pressure rise rate (and peak pressure) by the addition of a diluent.

Other aspects can be obtained from a study of the drawings, the disclosure, and the appended claims.