Internal combustion engine and operating method therefor

Abstract

The present disclosure provides an internal combustion engine having an engine housing with at least one cylinder that has diameter less than about 3 inches. A fuel injector is provided and disposed at least partially within the at least one cylinder, and includes a plurality of outlet orifices having a diameter between about 50 microns and about 125 microns, or about 0.05 millimeters and about 0.125 millimeters. The injector may include more than one set of separately controllable fuel outlet orifices, at least one of which could have an average diameter between about 0.05 millimeters and about 0.125 millimeters. The disclosure further provides a method of operating an internal combustion engine. The method includes the steps of rotating an engine crank shaft of the engine at a speed greater than about 5000 revolutions per minute, injecting a quantity of fuel into each of the cylinders, and burning at least every fourth piston stroke a sufficient quantity of the injected fuel to yield a brake mean effective pressure of at least about 200 lbs. per square inch.

Description

TECHNICAL FIELD

The present disclosure relates generally to internal combustion engines, and relates more particularly to a direct injection compression ignition engine and method utilizing fuel injectors having tiny outlet orifices.

BACKGROUND

Internal combustion engines have long been used as power sources in a broad range of applications. Internal combustion engines may range in size from relatively small, hand held power tools to very large diesel engines used in marine vessels and electrical power stations. In general terms, larger engines are more powerful, whereas smaller engines are less powerful. Engine power can be calculated with the following equation, where “BMEP” is brake mean effective pressure, the average cylinder pressure during the power stroke of a conventional four-stroke piston engine:Power=(BMEP)×(Engine Displacement)×(RPM)×(1/792,000).

(English units)

While larger engines may be more powerful, their power-to-weight or size ratio or “power-density” will be typically less than in smaller engines. Power varies with the square of a given scale factor whereas weight and volume vary with the cube of the scale factor. Scaling engine size up by a factor of two, for example, by doubling the cylinder bore size and doubling the piston stroke of a typical engine will, with everything else being equal, increase power about four times. The size and weight, however, will increase by about eight times. The “power density” may thus decrease by one half. The same principles are generally applicable when attempting to scale down an engine. Where bore size of a typical engine is decreased by a factor of two, engine power will decrease by a factor of four, but size and weight of the engine will decrease by a factor of eight. Thus, while smaller engines will have comparatively less available power output, their theoretical power density will in many cases be greater than similar larger engines.

Another related factor bearing on power density is the stroke distance of pistons in a particular engine. In many engines, there is a trade-off between stroke distance and RPM. Relatively longer stroke engines tend to have more torque and lower RPM, whereas relatively shorter stroke engines tend to have lower torque and greater RPM. Even where a short stroke engine and a long stroke engine have the same horsepower, the shorter stroke engine may have a greater power density since it may be a shorter, smaller engine.

For many applications, smaller, more power dense engines may be desirable. In many aircraft, for example, it is desirable to employ relatively small, lightweight, power dense engines with a relatively large number of cylinders rather than large engines having relatively fewer cylinders. However, attempts to scale down many internal combustion engines below certain limits have met with little success, particularly with regard to direct injection compression ignition engines. Many smaller, theoretically more power dense engines may be incapable of fully burning sufficient fuel per each power stroke in their comparatively small cylinders to meet higher power demands.

For example, if a conventional engine is running at a lower temperature and boost, where relatively small fuel quantities are injected for each cycle, and more power is demanded of the engine, an inability to burn the higher demanded fuel quantities may limit the engine's power output. As more fuel is injected over longer injection times, the liquid fuel spray can contact the piston surfaces and any other combustion chamber surfaces, known in the art as “wall wetting,” before it has a chance to adequately mix with the cylinder's fresh charge of air. This problem is particularly acute in smaller bore engines. Wall wetting can thus limit small bore engines to lower power and worse emissions than what intuitively could be their inherent capabilities, as wall wetting tends to cause poor combustion and high hydrocarbon and particulate emissions.

At relatively higher temperatures and in-cylinder pressures, wall wetting is less of a problem. Inadequate mixing of the fuel and air, however, can cause excessive smoke before combustion, limiting the engine's power long before its theoretical power limit is reached. One reason for these limitations is that at higher RPMs, there is only a relatively small amount of time within which to inject and ignite fuel in each cylinder.

As a result of the above limitations, two very general classes of small diesel engines have arisen, those that operate at relatively higher BMEP and lower RPM, and those that operate at relatively lower BMEP and higher RPM. However, neither type of engine is typically capable of providing an attractive power density commensurate with their size and weight. One example of a small bore diesel engine is the TKDI 600, designed by the Dr. Schrick company of Remscheid, Germany. The TKDI 600 claims a 34 KW output at 6000 RPM, or about 46 hp. The bore size of the TKDI 600 may be about 76 mm or about 3 inches, and the piston stroke may be about 66 mm or 2.6 inches. Although the TKDI 600 is claimed to have certain applications, such as in a small unmanned aircraft, the available BMEP is relatively low, about 169 PSI and the engine is therefore somewhat limited in its total available power output and hence, power density.

The present disclosure is directed to one or more of the problems or shortcomings set forth above.

SUMMARY OF THE DISCLOSURE

In one aspect, the present disclosure provides a method of operating an internal combustion engine, including the steps of injecting a liquid fuel into a combustion chamber of the engine in an engine cycle via a first set of outlet orifices but not a second set of outlet orifices, and injecting a liquid fuel into the combustion chamber via a second set of outlet orifices but not the first set in an engine cycle. The second set of outlet orifices include an average minimum cross sectional flow area less than an average minimum cross sectional flow area of the first set, the average minimum cross-sectional flow area of the second set being between about 0.002 square millimeters and about 0.01 square millimeters.

In another aspect, the present disclosure provides an engine having an engine housing with at least one combustion chamber therein, a piston movable within the at least one combustion chamber and configured to compress air therein to a compression ignition condition and a fuel injection apparatus disposed at least partially within the at least one combustion chamber. The fuel injection apparatus includes a first set of outlet orifices and a second set of outlet orifices, the fuel injection apparatus being configured to selectively spray liquid fuel into the combustion chamber via either of the first set of outlet orifices and the second set of outlet orifices, the second set of outlet orifices having an average minimum cross-sectional flow area less than an average minimum cross-sectional flow area of the first set, the average minimum cross-sectional flow area of the second set being between about 0.002 square millimeters and about 0.01 square millimeters.

In still another aspect, the present disclosure provides a fuel injection apparatus for an internal combustion engine, including at least one injector body having at least one fuel supply passage therein, a first set of fuel outlet orifices having a first average minimum cross sectional flow area and a second set of fuel outlet orifices having a second average minimum cross-sectional flow area less than the first average minimum cross-sectional flow area. The second average minimum cross-sectional flow area is between about 0.002 square millimeters and about 0.01 square millimeters. A first check is provided which is configured to control fluid communication between the first set of outlet orifices and the at least one fuel supply passage to control spraying of a liquid fuel from the first set of outlet orifices into a combustion chamber of an engine. A second check is provided which is operable separately from the first check and configured to control fluid communication between the at least one fuel supply passage and the second set of outlet orifices to control spraying of a liquid fuel from the second set of outlet orifices into a combustion chamber of an engine.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic illustration of an engine according to the present disclosure;

FIG. 2 is an enlarged sectioned side diagrammatic view of a portion of an engine cylinder that includes a fuel injector, according to the present disclosure;

FIG. 3 is a graph illustrating plots of various compression ignition engine types relating BMEP and RPM;

FIG. 4 is a schematic view of a portion of an engine system according to another embodiment; and

FIG. 5 is a graph illustrating fuel injection rate shaping according to one embodiment.

DETAILED DESCRIPTION

Referring to FIG. 1, there is shown a schematic illustration of an engine 10 according to one embodiment of the present disclosure. Engine 10 includes an engine housing 12 having a plurality of cylinders 14 therein. A fuel injector 16 is disposed at least partially within each of cylinders 14 and operable to direct inject a liquid fuel therein. Each of fuel injectors 16 may include a fuel injector tip 20 extending into the associated cylinder, and each tip 20 has a plurality of outlet orifices 22. Engine 10 further includes a plurality of pistons 21, each disposed at least partially within one of cylinders 14 and movable therein, and each piston is coupled with a crankshaft 30 via a piston rod 23. Engine 10 may further include a pressurized fuel source 17, which may include a high pressure pump or cam-actuated fuel pressurizer, for example. Pressurized fuel source 17 may be fluidly connected with each of fuel injectors 16 via a high pressure feed line or common rail 19 and a plurality of supply passages 26. It is contemplated that source 17 will pressurize fuel to at least about 150 MPa, although the present disclosure is not thereby limited. Relatively higher pressures have in some instances been shown to facilitate atomization of injected fuel, however, the actual pressure may be selected based upon various desired operating characteristics of the particular engine, and feasibility. It is contemplated that engine 10 may be either a compression ignition engine, for example a diesel engine, or a spark ignited engine using, for instance, gasoline. Engine 10, or any of the other engines contemplated herein, may include at least one sensor 27 configured to sense values indicative of engine speed and/or engine load, and output corresponding signals to an electronic controller 15.

Referring also to FIG. 2, there is shown a close-up view of a portion of engine 10 of FIG. 1, including a cylinder 14 with a piston 21 movably positioned therein. Each cylinder 14 of engine 10 will have a diameter D₁, that is less than about 3 inches, and may be between about 2 inches and about 3 inches. About 3 means between 2.5 and 3.5. About 2.5 means between 2.45 and 2.55. These examples will allow one to determine precisely what is meant by the phrase about X, in the context of the present disclosure. In certain embodiments, D₁ will be between about 2.5 and about 2.8 inches, and may also be about 2.7 inches in one practical embodiment. Although it is contemplated that engine 10 might be constructed having only a single cylinder and single piston, most embodiments will include a plurality of cylinders and pistons, typically at least eight, and embodiments are contemplated wherein engine 10 includes 12 cylinders, or even up to 16 or more cylinders depending upon the application. The arrangement of cylinders in engine 10 may comprise any known configuration, such as a V-pattern, in-line, radial, opposed, etc. In many embodiments, size and space will be at a premium and thus a V-pattern engine, for example, may be a practical design.

Engine 10 may be either of a two-stroke or four-stroke engine, although it is contemplated that a four-stroke cycle will be a practical implementation strategy. To this end, fuel will be injected via fuel injectors 16 at least about once every fourth piston stroke. Each piston 21 will typically have a stroke distance “L” that is between about 2 inches and about 3 inches, and embodiments are contemplated wherein the stroke distance of each piston 21 will be about 2.5 inches. Given the typical stroke distance of each piston 21, the total displacement of each cylinder 14 of engine 10 will typically be less than about 25 cubic inches and may be between about 6 cubic inches and about 25 cubic inches. Embodiments are contemplated wherein the total displacement of each cylinder 14 will be between about 7 cubic inches and about 25 cubic inches, and may be about 14 cubic inches, for example.

At least a portion of outlet orifices 22 of each fuel injector 16 will be between about 50 microns and about 125 microns in diameter, D₂ in FIG. 2. References herein to microns should be understood as corresponding to metric units, thus 50 microns equals 0.05 millimeters, 60 microns equals 0.06 millimeters, 85 microns equals 0.085 millimeters, 90 microns equals 0.09 millimeters, 110 microns equals 0.1 millimeters, and 125 microns equals 0.125 millimeters. In certain embodiments, some or all of orifices 22 will be between about 0.06 millimeters and about 0.09 millimeters and some or all may be about 0.085 millimeters. Orifices 22 may be formed by laser drilling holes in injector tip 20 connecting an exterior of injector tip 20 with a nozzle chamber 24 of each fuel injector 16. One suitable laser drilling process is taught in commonly owned U.S. Pat. No. 6,070,813 to Durheim. Although it is contemplated that laser drilling of orifices 22 will be a workable strategy, other methods of forming ultra small injector orifices may be used. For instance, orifices 22 may be formed via known methods of coating or plating larger holes down to the desired diameter, or casting ceramic injector nozzles with small wires therein, and burning the wires away during curing of the nozzles, or any other currently known or to be discovered injector orifice making technique.

The number of orifices 22 may vary, in most embodiments the ultra-small orifices of orifices 22 will number greater than about 8 and typically between about 10 and about 30. Flow area will vary with the square of a scale factor in orifice diameter. Thus, designing an engine having fuel injector orifices with approximately one half the diameter of conventional, 160 micron orifices, for example, will yield a flow area per each 80 micron orifice that is ¼ that of a 160 micron orifice. Thus, in this example, at least 4 smaller holes are necessary to equal the flow area capability of one larger orifice.

It is contemplated that orifices 22 may have a variety of shapes. Conventional fuel outlet orifices are generally cylindrical, however, recent advances in orifice forming techniques have opened the door to the use of more complex shapes, tailored specifically to certain applications. Thus, in some embodiments, orifices 22 might be tapered, trumpet-shaped, oval in cross section, or still some other shape. It is contemplated, however, that orifices 22 will in most embodiments have an average minimum cross sectional flow area that is between about 0.002 square millimeters and about 0.01 square millimeters. Thus, those skilled in the art will appreciate that many different orifice configurations, number, size, pattern, etc. may be implemented in a fuel injector and/or engine which will fall within the scope of the present disclosure.

Depth of penetration of the fuel spray will be generally linearly related with orifice size. The likelihood and degree of wall wetting and spraying of the injected fuel onto a piston face in a given cylinder will typically be related to depth of penetration of the fuel spray. Accordingly, because smaller cylinder bores tend to experience wall wetting more easily than larger bores, it may be generally desirable to utilize relatively smaller orifices with relatively smaller cylinder bore sizes. For example, in an embodiment wherein D₁ is relatively closer to 2 inches, orifices having a diameter D₂, relatively closer to 0.05 millimeters may be appropriate. The converse may be applicable to larger size cylinders, e.g. closer to 3 inches and having fuel injector orifices closer to 0.125 millimeters.

In one specific example, it is contemplated that engine 10 will utilize a fuel system capable of delivering a fuel injection pressure of at least about 150 MPa, and in some instances at least about 240 MPa. Increased fuel injection pressures have been found to enhance mixing of the fuel and air without substantially affecting the depth of penetration of atomized fuel into the cylinder. Fuel flow rate scales with the square root of the scale factor, thus doubling injection pressure will yield an increase in flow rate for a given orifice size that is about the square root of two times the original flow rate.

The present disclosure further provides a method of operating an internal combustion engine. The method may include the step of rotating crankshaft 20 of engine 10 at greater than about 5000 RPM, and in certain embodiments or under certain operating conditions at greater than about 6000 RPM, or even greater than about 6500 RPM. The method may further include burning a sufficient quantity of injected fuel in each of cylinders 14 to yield a brake mean effective pressure (BMEP) of at least about 200 pounds per square inch (PSI), and in certain embodiments or under certain operating conditions burning sufficient fuel to yield a BMEP of at least about 250 PSI, or even at least about 350 PSI.

Referring also to FIG. 3, three specific embodiments of engines according to the present disclosure W, V and F are represented therein, all located within an operating zone Z of engines according to the present disclosure, described hereinbelow. Certain specifications of engines W, V and F are set forth in the following table, in comparison to conventional engines M and U. All of engines W, V and F will include a plurality of injector orifices 22 having a diameter D₂ within the described predetermined ranges of about 0.05 millimeters to about 0.125 millimeters. As described herein, power density is the ratio of power to mass/volume. Those skilled in the art will appreciate that bore size of a particular engine will be related to engine mass/volume. Thus, in general terms, the 6 inch bore of engine M is scaled by a factor of 2 with regard to the 3 inch bore of engine F. With a scale factor of 2, power of engine M will be about 4 times that of engine F per cylinder, as power varies with the square of the scale factor. Mass and volume of engine M, however, will be about 8 times the mass and volume of engine F per cylinder, as mass and volume vary with the cube of the scale factor. Engine F will thus be more power dense than engine M.

	M	U	W	V	F
bore size	6	in.	4	in.	2	in.	2.7	in.	3	in.
stroke	6	in.	4	in.	2	in.	2.5	in.	3	in.
distance
cylinders	4	4	16	12	16
bmep	400	psi	400	psi	400	psi	400	psi	400	psi
rpm	2667	4000	8000	5926	5334
power	914	hp	406	hp	406	hp	514	hp	914	hp
displacement	678.6	in3	201	in3	100.5	in3	171.8	in3	339.3	in3
hp/in3	1.35	2.02	4.04	2.99	2.69

Turning to FIG. 4, there is shown schematically a portion of an engine system 110 according to another embodiment. Engine system 110 includes at least one cylinder 114 having a piston 121 reciprocable therein. Engine system 110 may also comprise a direct injected compression ignition engine, having certain similarities with the foregoing embodiments, but also differing in that rather than a single fuel injection orifice set, a fuel injection apparatus 116 is provided which includes a first set of outlet orifices 124 and a second set of outlet orifices 122, separate from the first set. Outlet orifices 124 and 122 may be disposed in an injector body 119 extending at least partially into cylinder 114. Fuel injection apparatus 116 may also be coupled with a common rail 19 and include a control valve assembly 131 configured to control fuel injection into cylinder 114 via apparatus 116. Control valve assembly 131 may include separate control valves 132a and 132b, each including an electrical actuator for example, configured to control fluid communication between common rail 19 and orifices 122 and 124 via at least one fuel supply passage 135. Passage 135 may be disposed at least partially within injector body 119.

Fuel injection apparatus 116 may comprise separate, side-by-side sets of outlet orifices, or it might alternatively include one of the various dual concentric check injectors which are known in the art. In either case, however, fuel injection apparatus 116 will typically be capable of separately controlling fuel spray out of the respective sets of outlet orifices 124 and 122. In one embodiment, separate, direct control of fuel spray may be achieved via a first needle check 118a and a second needle check 118b configured to separately control fuel spray out of orifices 124 and 122, respectively, needle checks 118a and 118b being operably coupled with control valves 132a and 132b, respectively. As used herein, the term “direct control” should be understood as referring to a system wherein the application of fluid pressure or some other closing force to a control surface of a valve member such as needle valve members 118a and 118b is used to control the closing and/or opening of the respective sets of orifices. In other words, direct control will utilize some means other than fluid pressure acting on opening hydraulic surfaces to enable fuel injection. To this end, control valve assembly 131 may comprise any of a variety of direct control systems.

In the embodiment shown in FIG. 4, hydraulic pressure is controllably applied to and relieved from a first pressure surface 117a and a second pressure surface 117b of needle checks 118a and 118b, the respective pressure surfaces being exposed to a fluid pressure in first and second needle control chambers 133a and 133b. Control valves 132a and 132b may be independently operable to permit different hydraulic pressures to be applied to pressure surfaces 117a and 117b. In a typical embodiment, one or both of control valves 132a and 132b will provide for supplying of rail pressure to control chambers 133a and 133b. Control valves 132a and 132b may be actuated to connect one or both of chambers 133a and 133b to a low pressure drain passage 137, relieving pressure in control chambers 133a and/or 133b and allowing rail pressure to lift the corresponding needle check 118a and/or 118b to permit the spraying of fuel from the associated orifices 124, 122.

It is further contemplated that in the FIG. 4 embodiment, at least one of the sets of outlet orifices 124 and 122 may comprise tiny outlet orifices having sizes and/or flow rates similar to outlet orifices 22, described with regard to the FIGS. 1 and 2 embodiments. The other set of orifices may be a conventional set, for example, including orifices having relatively larger sizes at or close to what would be considered suitable for a given engine in view of the present state of the art, for example between about 0.15 millimeters and about 0.20 millimeters. In still other embodiments, each of the sets of orifices 124 and 122 could include orifices having sizes and/or flow rates similar to orifices 22. In such instances, different numbers of orifices in the respective sets could be used to achieve different net flow rates, or flow areas.

Each of the sets of orifices 124 and 122 may be disposed in an annular pattern about an axis A₁ and an axis A₂, respectively, extending through the corresponding needle checks 118a and 118b. Orifices 124 and orifices 122 may also be disposed at different average spray angles relative to axes A₁ and A₂. In particular, orifices 122, the relatively smaller set in one embodiment, may be disposed at a relatively narrower average spray angle, whereas orifices 124 may be disposed at a relatively larger average spray angle. It should be appreciated that the embodiment of FIG. 4 is calculated to be applicable to both relatively small bore engines such as that described with regard to FIGS. 1 and 2, as well as other engines in relatively larger size ranges.

INDUSTRIAL APPLICABILITY

During a typical four-stroke cycle, a main fuel injection will take place when each of pistons 21 is at or close to a top dead center position, every fourth piston stroke and in a conventional manner. Additionally, smaller pilot and/or post injections may accompany each main injection. In a compression ignition version of engine 10, compressed air and the injected, atomized fuel will ignite and combust to drive each of the respective pistons 21 and rotate crankshaft 30. Spark ignited designs will typically use a spark plug in a well known fashion to effect ignition.

Directly injecting fuel into cylinder 14 via orifices 22 having the predetermined diameter ranges described herein can allow ignition and better or more efficient combustion of a greater quantity and proportion of the injected fuel than in designs utilizing conventional fuel injection orifices. Several advantages result from this ability. First, the potential BMEP is higher. Higher BMEP in each cylinder means that an overall greater average pressure can act on each piston 21, providing more force to drive each piston 21 in its respective cylinder 14 and rotate crankshaft 30. The relatively smaller size of atomized fuel droplets from orifices 22 than from conventional sized orifices is believed to enhance ignition and overall combustion as compared to the larger fuel droplets in a conventional design. The spray pattern from each injector orifice may have such a spread angle and internal fuel/air ratio that the mixing with the charge air may be much faster. Accordingly, this may allow both a greater absolute quantity of fuel to be burned, and may allow the fuel to be burned faster and more easily ignite. It may also allow a greater proportion of the fuel injected to burn than in earlier designs. The higher injection pressure expected to be used in conjunction with the smaller orifices will help compensate for the lower flow rates of the smaller orifices and also will help fuel/air mixing without substantially affecting the depth of fuel penetration. In general, the combination of smaller orifices and higher pressure can thus allow better combustion before reaching wall-wetting and its associated degradation of combustion.

Secondly, given the inherently limited time within which to burn the injected fuel, the relatively smaller fuel droplets and a lower fuel/air ratio within the fuel spray plume available in engine 10 can allow fuel ignition and combustion to take place more quickly, allowing relatively faster piston stroke speeds and correspondingly greater engine RPMs. The combination of relatively greater BMEP and higher RPM allows engine 10 to operate with a relatively higher power, and hence with a higher power density than many heretofore available small cylinder bore engine designs.

Certain earlier small cylinder bore engines were able to approach the BMEP possible in engine 10, but not without shortcomings in other operating parameters. In order to burn sufficient fuel during each power stroke to achieve higher BMEP, many earlier engines typically operated at lower RPM than engine 10. In an attempt to cram more fuel into each cylinder for every ignition stroke, and increase the BMEP, in some known operating schemes an excess of fuel is delivered to each cylinder. Where an excess of fuel is made available, however, the quantities of unburned hydrocarbons, soot and other pollutants may be so high as to make operation undesirable and inefficient in many environments. For instance, a visible “smoke signature” may be undesirable in certain military applications.

Similarly, certain earlier small bore engine designs are known that operate at an RPM approaching that of engine 10, but not without their own set of tradeoffs. In such relatively higher RPM engines, BMEP tends to be lower as smaller fuel injection quantities are injected to avoid excessive smoke and wasting of fuel. As a result, such engines may operate at relatively high RPM, but insufficient fuel can be burned during each power stroke to reach higher BMEP. In either previous design/scheme the available power of the engine is relatively lower than in similar engines of larger size, and the power density of such smaller engines tends to be lower than what it might in theory be given their relatively smaller size.

Engine horsepower is directly proportional to both RPM and BMEP, hence the capability of engine 10 to operate at both relatively high RPM and BMEP allows the total available power of engine 10 to be significantly greater than in previously known designs. Given the relatively small size of engine 10, its power density can be more commensurate with its actual size, and engine 10 can take fuller advantage of its small scale design than previous engines.

Engine 10 provides still further advantages over known designs which relate to the enhanced ease of ignition of the fuel injected through orifices 22. During cold starting conditions, many known compression ignition engines utilize external heat sources or the addition of combustible compounds such as ether to initially begin operating. In a compression ignition version of engine 10, the need for these and similar starting aids may be reduced over earlier designs or eliminated, as the smaller fuel droplets and lower fuel/air ratio in the fuel spray plume tend to make ignition occur more readily.

Further advantages of engine 10 relate to its ability to quiescently mix fuel and air in certain contemplated embodiments. This approach contrasts with most if not all earlier small cylinder bore designs wherein “swirl” mixing was necessary to mix the charge of fresh air with injected fuel. Swirl mixing requires a swirling of the charge of air delivered to the cylinder, primarily via appropriate geometry of the air intake system or turbochargers and cylinder ports. In contrast, quiescent mixing is commonly used in larger engine designs, wherein simply spraying the fuel into un-swirled air will provide sufficient mixing. Quiescent mixing may have the advantage of transferring less heat from the combustion space to the cylinder walls, head and piston during combustion and, accordingly, will allow more heat energy to be converted to shaft horsepower rather than transferred to the coolant through the cylinder walls, head and piston.

Still further advantages relate to the fuel economy of engine 10, as well as its relatively lower emissions. Burning more of the injected fuel allows the relative quantity of unburned hydrocarbons emitted from engine 10 to be reduced, improving its use of the fuel made available. In some contemplated embodiments, such as in certain aircraft, weight may be at a premium. Thus, in engine 10 the mass and size of the engine itself are not only relatively smaller, but the quantity of fuel that must be carried for a given travel range is reduced. In addition, the relatively higher proportion of fuel burned can reduce the smoke emitted during operation. There has been a perception that diesel engines often emit relatively large quantities of visible smoke. Aesthetics, environmental and in some instances tactical concerns, such as in military vehicles, can make minimizing visible smoke desirable or imperative. Engine 10 will typically be capable of substantially smokeless operation, for example, having a Bosch Smoke Number of 3 or less for transient operation and 2 or less for steady state operation. One means for quantifying the smoke content of engine exhaust is an exhaust opacity “smoke meter” such as the Bosch ESA 110-Computer Controlled Smoke Meter, available from Equipment Supplies Biddulph of Biddulph, Staffs, United Kingdom and other commercial suppliers.

Turning to FIG. 3, there is shown a plot of the operating zone of several different sets of conventional diesel engines in comparison to the operating zone Z of engine 10, and approximate locations of engines M and U of the foregoing table. The Y axis represents BMEP whereas the X axis represents RPM. In FIG. 1, set P represents a group of relatively heavy duty diesel engines having a BMEP between about 250 PSI and about 325 PSI. The engines of set P may include relatively smaller diesel engines, such as small scale power generators, mid-size engines such as might be found in trucks or off-highway work machines, and large diesel marine or power generation engines. The range of RPM in engines of set P tends to be between about 1000 RPM and about 2500 RPM. Set Q includes engines such as are known from common pick-up trucks, having a relatively higher RPM but lower BMEP than those of set P. Set R includes engines such as certain military vehicles having BMEP between about 350 PSI and about 400 PSI, and RPM between about 3000 and about 4000. Set S in turn includes such engines as may be used in many European passenger cars. Set T includes engines such as certain military motorcycle engines and engines proposed for unmanned aerial vehicles, with BMEP between about 150 PSI and 175 PSI and RPM between about 5500 and about 6000. As illustrated in FIG. 3, the operating zone of engine 10 includes higher BMEP and RPM in combination than any of the other, known engine types or groups. Pushing the engine RPM limits above that of known engines, particularly diesels, and elevating the attainable BMEP as described herein can thus provide a relatively small, lightweight and powerful engine. Point V of FIG. 3 represents one possible embodiment of the present disclosure, capable of a BMEP of about 400 PSI or greater, and an RPM between about 6000 and about 6500.

While much of the foregoing description focuses on the use of tiny fuel outlet orifices in a relatively small, power dense engine, the present disclosure is not thereby limited. In other embodiments, the use of tiny orifices may confer advantages in relatively larger engines, particularly direct injection diesel engines. In one specific embodiment, using both tiny outlet orifices and conventional outlet orifices similar to that shown in FIG. 4, the respective orifice sets can be used to inject fuel separately based on particular engine operating conditions such as speed and/or load. A sensor such as sensor 27 shown in FIG. 1 may also be used in engine 110 in determining the relative engine speed and/or load for purposes of selecting a desired injection strategy. Signals from sensor 27 may be input to an electronic controller similar to controller 15 shown in FIG. 1, and appropriate commands output to control valves 132a and 132b to inject fuel from the desired set of orifices based on the speed and/or load of engine 110.

During relatively lower speed and/or load conditions, it may be desirable to utilize the relatively smaller outlet orifices, for example, tiny orifices of set 122 in the FIG. 4 embodiment. Where engine 110 is operating in a lower portion of a speed and/or load range, injected liquid fuel may have a relatively greater tendency to impinge upon the piston surfaces and/or walls of the engine's combustion chamber. Accordingly, the relatively lesser depth of penetration associated with fuel spray from orifices 122, having an average minimum cross sectional flow area between about 0.002 square millimeters and about 0.01 square millimeters, can enable operation with little or no wall wetting. Reduced or no wall wetting is associated with various advantages, as described above. At relatively higher speeds and/or loads, for instance in an upper half of a speed and/or load range, injection of relatively larger quantities of fuel, at relatively higher flow rates, for example, may be appropriate. In such instances, orifices 124, having conventional average size, may be used. Inputs from sensor 27 may be used to indicate speed and/or load range to determine that operation in one or more engine cycles using orifices 124 but not orifices 122 is appropriate, or that operation in one or more engine cycles using orifices 122 but not orifices 124 is appropriate.

It should further be appreciated that the present disclosure is applicable to different operating strategies relating to injection timing, size and injection rate shaping. In one example, the relatively smaller orifices 122 might advantageously be used for one or more pilot injections, or one or more post injections, whereas orifices 124 could be used for one or more relatively large, main injections. The same set of orifices might also be used for each of a plurality of injections in a given engine cycle. Orifices 122 might also be used for injections relatively early in an engine cycle in such operating regimes as are generally known as homogeneous charge compression ignition or HCCI. In addition to or instead of HCCI-style injections, pilot injections, post injections, etc., either of orifices 122 and 124 might be used to inject fuel for conventional diffusion burning. As piston 121 reciprocates, it may compress air to a compression ignition condition in cylinder 114, before, during and/or after which injection out of one of orifices 122 and 124 may be initiated to achieve a diffusion burn of fuel in combustion chamber 114.

Still another feature of the present disclosure relates to the relatively greater ability to control fuel injection rate, particularly at the start of injection and end of injection, through the use of the multiple, separately controlled sets of outlet orifices disclosed herein. Referring to FIG. 5, there is shown a graph wherein the y-axis represents injection rate and the x-axis represents time. In FIG. 5, “G” denotes a curve representing fuel injection rate over time, the profile of the curve G illustrating a fuel injection rate shape. It may be noted that curve G includes an initial portion “B” corresponding to an initial period of fuel injection known to those skilled in the art as a “boot.” It has heretofore been difficult, if not impossible, to control the relative shape of the boot in a fuel injection rate curve. The use of a conventional single check generally results in the boot portion of a fuel injection curve being essentially an all or nothing phenomenon, given challenges in achieving the extremely precise control over the position of the outlet check that would be required to tailor the boot.

The use of dual sets of orifices 122 and 124 is contemplated to provide relatively more precise control over fuel injection rate in the boot portion of an injection rate curve than that available in conventional strategies. In other words, rather than the initial portion, i.e. the boot, of an injection rate curve being all or nothing, the present disclosure may allow the boot shape to be controlled cycle to cycle. One specific aspect of the boot which may be controlled is its relative length. In FIG. 5, a portion of the boot shown via range R represents an approximate plateau which typically exists between initial opening of fuel injection orifices and a relatively sharper increase in fuel injection rate subsequent to range R. Separate control over fuel injection orifices 124, 122, is contemplated to provide sufficiently precise control in some instances that the relative size of range R may be varied, as shown by the different available initial profiles of curve G in the boot portion B. The profile of curve G during the main portion of fuel injection can also be varied, as represented by broken line G₁ in FIG. 5. Further, rather than a boot contiguous with the rest of the injection curve, the boot might instead be a tiny injection followed by, but separate from, a main injection having a relatively shorter or even negligible boot portion, as illustrated via broken lines G₂ in FIG. 5. Use of the strategy described herein may also provide for improved injection rate control toward the end of fuel injection, as injection rate drops toward zero.

The present description is for illustrative purposes only, and should not be construed to narrow the breadth of the present disclosure in any way. Thus, those skilled in the art will appreciate that various modifications might be made to the presently disclosed embodiments without departing from the intended spirit and scope of the present disclosure. For example, while many of the embodiments described herein are discussed in the context of both elevated BMEP and elevated RPM, those skilled in the art will appreciate that in certain applications it may be desirable to operate an engine with only one of RPM or BMEP significantly elevated as compared to conventional engines. It may be noted that set Z of FIG. 3 encompasses a relatively broad operating range of both BMEP and RPM. Small cylinder bore engines might be designed according to the present disclosure capable of operating at relatively high RPM of at least about 7500, but with BMEP no greater than about 200 PSI. Similarly, higher BMEP engines, but with relatively lower RPM may be desirable for other applications. The directly proportional relationship of both RPM and BMEP with power thus allows substantial flexibility in designing relatively high power density, small cylinder bore direct injected engines according to the present disclosure. Still further embodiments are contemplated wherein orifice size, shape, orientation, etc. varies, and can vary orifice to orifice on a given injector tip. This includes, for example, using a plurality of ultra-small orifices, a plurality of larger, conventional sized orifices, with individual geometric shape and orientation varying to create a simple or complex array of orifices to provide the best overall spray pattern. Thus, there need be no particular sizing or any particular number or arrangement of ultra-small hole orifices so long as a sufficient number are provided to impart the desired operating characteristics, as described herein. Other aspects, features and advantages will be apparent upon an examination of the attached drawing Figures and appended claims.

Claims

1. A method of operating an internal combustion engine comprising the steps of: injecting a liquid fuel into a combustion chamber of the engine, which is an engine cyclinder with a diameter less than about 3 inches, in an engine cycle via first set of outlet orifices but not a second set of outlet orifices;injecting a liquid fuel into the combustion chamber via a second set of outlet orifices but not the first set in an engine cycle, the second set of outlet orifices having an average minimum cross sectional flow area less than an average minimum cross sectional flow area of the first set, the average minimum cross sectional flow area of the second set being between about 0.002 square millimeters and about 0.01 square millimeters;compression igniting the injected fuel; andburning the injected fuel to yield a brake mean effective pressure of at least about 200 pounds per square inch.

2. The method of claim 1 further comprising a step of compressing air in the combustion chamber with a piston to an autoignition condition in a plurality of engine cycles; wherein the step of injecting liquid fuel via the first set of outlet orifices includes injecting liquid fuel after air in the combustion chamber is compressed to an autoignition condition in an engine cycle; andwherein the step of injecting liquid fuel via the second set of outlet orifices also includes injecting liquid fuel after air in the combustion chamber is compressed to an autoignition condition in an engine cycle.

3. The method of claim 2 wherein the step of injecting liquid fuel via the first set of outlet orifices includes injecting liquid fuel in a first engine cycle, and wherein the step of injecting liquid fuel via the second set of outlet orifices includes injecting liquid fuel in a second engine cycle that is different from the first engine cycle.

4. The method of claim 3 wherein: the step of injecting liquid fuel via the first set of outlet orifices includes injecting liquid fuel at least in part by controlling fluid communication between a fuel supply passage and the first set of outlet orifices with a first direct operated check; andthe step of injecting liquid fuel via the second set of outlet orifices includes injecting liquid fuel at least in part by controlling fluid communication between a fuel supply passage and the second set of outlet orifices with a second direct operated check, the second set of outlet orifices including at least about ten outlet orifices having an average diameter in the range of about 0.06 millimeters to about 0.09 millimeters.

5. The method of claim 4 wherein each of the injecting steps includes injecting fuel into the combustion chamber via a fuel injection apparatus that includes at least one injector body disposed at least partially within the combustion chamber, the fuel injection apparatus being fluidly connected via the fuel supply passage with a high-pressure rail.

6. The method of claim 5 wherein the piston defines a displacement between about 6 cubic inches and about 25 cubic inches, and wherein the step of compressing air in the combustion chamber includes compressing between about 6 cubic inches and about 25 cubic inches of air in one of every four piston strokes.

7. The method of claim 2 further comprising a step of monitoring at least one of engine speed and engine load, wherein the step of injecting liquid fuel via the first set of outlet orifices comprises injecting fuel via the first set of outlet orifices but not the second set where the engine is at a relatively higher speed and load, and wherein the step of injecting liquid fuel via the second set of outlet orifices comprises injecting fuel via the second set of outlet orifices but not the first set where the engine is at a relatively lower speed and load.

8. The method of claim 7 wherein: the engine comprises a plurality of cylinders, a plurality of pistons reciprocable one within each of the cylinders and a plurality of fuel injection apparatuses each disposed at least partially within one of the cylinders and having a first set of outlet orifices with an average diameter between about 0.15 millimeters and about 0.20 millimeters and a second set of outlet orifices with an average diameter between about 0.06 millimeters and about 0.09 millimeters; andthe method further comprises injecting fuel into each of the cylinders via the respective first sets of outlet orifices of each of the fuel injection apparatuses at a first average spray angle, and injecting fuel into each of the cylinders via the respective second sets of outlet orifices at a second, narrower average spray angle.

9. An engine comprising: an engine housing having at least one combustion chamber therein, which is an engine cylinder with a diameter less than about 3 inches;a piston movable within said at least one combustion chamber and configured to compress air therein to a compression ignition condition;a fuel injection apparatus disposed at least partially within said at least one combustion chamber and having a first set of outlet orifices and a second set of outlet orifices, said fuel injection apparatus being configured to selectively spray liquid fuel into said combustion chamber via either of the first set of outlet orifices and the second set of outlet orifices, said second set of outlet orifices having an average minimum cross sectional flow area less than an average minimum cross sectional flow area of the first set, the average minimum cross sectional flow area of the second set being between about 0.002 square millimeters and about 0.01 square millimeters; andwherein a quantity of fuel is burned in each engine cylinder to yield a brake mean effective pressure of at least about 200 pounds per square inch.

10. The engine of claim 9 wherein said fuel injection apparatus comprises an injector body wherein said first and second sets of outlet orifices are disposed, said injector body being positioned at least partially within said at least one combustion chamber.

11. The engine of claim 10 wherein an average diameter of said second set of outlet orifices is between about 0.05 millimeters and about 0.125 millimeters.

12. The engine of claim 11 wherein the average diameter of said second set of outlet orifices is between about 0.06 millimeters and about 0.09 millimeters.

13. The engine of claim 12 wherein said at least one combustion chamber comprises a plurality of engine cylinders, said engine further comprising a plurality of pistons reciprocable one within each of said cylinders and a plurality of fuel injection apparatuses each including an injector body disposed at least partially within one of said cylinders and having a first set of outlet orifices and a second set of outlet orifices, the respective second sets of outlet orifices each having an average diameter between about 0.06 millimeters and about 0.09 millimeters.

14. The engine of claim 13 wherein each of said plurality of pistons has a displacement in the range of about 6 cubic inches to about 25 cubic inches.

15. The engine of claim 13 further comprising a common rail connected to a source of pressurized fuel, wherein the respective first and second sets of outlet orifices are disposed side by side in the nozzle body of the corresponding fuel injection apparatus, each fuel injection apparatus further comprising: a first electrically actuated control valve operably coupled with a first needle check configured to control fluid communication between said common rail and the first set of outlet orifices of the corresponding fuel injection apparatus;a second electrically actuated control valve operably coupled with a second needle check configured to control fluid communication between said common rail and the second set of outlet orifices of the corresponding fuel injection apparatus;at least one sensor configured to monitor at least one of engine speed and engine load and output signals corresponding with the at least one of engine speed and engine load; andan electronic controller coupled with said at least one sensor and in control communication with each of said control valves, said electronic controller being configured to output control commands to each of the control valves responsive to signals from said at least one sensor.