Operational strategy for fuel agnostic prechamber enabled mixing-controlled combustion

Abstract

A method of mixing controlled combustion in a cylinder with a prechamber includes actively fueling the prechamber with a first fuel injector. The prechamber is sparked with a spark plug operatively connected to the prechamber. Fuel is injected into a main chamber from a second fuel injector. Jet flames are produced that that exit at least one orifice of a nozzle of the prechamber. The jet flames extend into the main chamber. The fuel injected from the second fuel injector into the main chamber is ignited by the jet flames.

Description

OPERATTONAL STRATEGY FOR FUEL AGNOSTIC PRECHAMBER ENABLED MIXING-CONTROLLED COMBUSTIONSTATEMENT REGARDING GOVERNMENT SPONSORED RESEARCH OR DEVELOPMENT

[0001] This invention was made with government support under the U.S. Department of Energy’s Office of Energy Efficiency and Renewable Energy under Award Numbers DE- EE0009872 and DE-AR0001528. The U.S. Government has certain rights to this invention.BACKGROUND

[0002] Internal combustion engines (ICE) are fundamental technologies used in the day- to-day life of people and industries. Spark ignition (SI) engines are widespread for comparatively light-duty applications. SI engines are well adapted to bum low carbon intensity fuels with low reactivity / high octane, for example gasoline, E10 and E15 fuel blends, and ethanol (E100). SI engines lack the performance characteristics desired for robust heavy-duty operation. SI engines are not able to meet the transient and high torque at low- speed requirements of heavy-duty applications. Knock is a common limiting factor to obtain high load at reasonable efficiency in heavy-duty SI engines. Alternative fuels like ethanol and methanol exhibit high latent heat of vaporization and excellent knock resistance but exhibit large thermal sensitivity and excessive cyclic variability near the lean-limit. SI engines further lack good transient snap-torque response in fear of knock and pre-ignition. The transient nature and peak load demands of heavy-duty engine applications limit the applicability of SI engines in heavy-duty applications.

[0003] Therefore, heavy-duty engines are predominantly compression ignition (CI) diesel fuel engines. The mixing-controlled combustion (MCC) process is well adapted to lower reactivity and high compression ratios exhibited by diesel fuel. The MCC process enables high fuel conversion efficiency and robust power delivery over a wide operational range. The MCC process of fossil diesel fuel however comes at the cost of high emissions of carbon monoxide (CO), carbon dioxide (CO2), nitrogen oxides (NOx), and soot.

[0004] A natural solution to combat high emissions is to replace diesel fuel with high octane, low cetane, low carbon intensity fuels in the MCC process. However, in order to burn low- cetane fuels in MCC, some form of ignition assistance must be provided to initiate combustion. Heightened in-cylinder temperatures during engine operation will increase the propensity of thefuel to auto-ignite. Numerous solutions have been attempted, however, none have produced the desired performance to provide a suitable alternative to diesel MCC. One such attempt included the use of glow plugs, typically used for startup diesel ignition assistance, operationally to continuously increase chamber temperature. Mueller et al. “Glow Plug Assisted Ignition and Combustion of Methanol in an optical di diesel engine” SAE Tech Pap. DOI: 10.4271 / 2001-01- 2004, found this to combust low-cetane fuels in an MCC engine, however, glow plug technology exhibits much greater power requirements for operational glow plug use with expected significant reduction in glow plug life when used under such power requirements.

[0005] Dual-fuel systems seek to combine diesel and a low cetane fuel to achieve the properties of both. In non-premixed systems, a diesel pilot is used to initiate combustion of the low cetane fuel. However, these systems exhibited a lack of combustion stability over the operational rand and increased HC emissions at middle to low loads with a high methanol substitution ratio. Additionally, a dual-fuel system requires additional complexity to design, operate, and maintain two fuel systems.

[0006] US Patent No. 10,458,311, entitled “Internal Combustion Engine” and US Patent No. 11,840,954, entitled “Spark Ignited Engine with a Pre-Chamber, A Prechamber and an Adapter Insert for the Engine” both disclose prechambers for use with an internal combustion engine and are incorporated by reference herein.BRIEF DISCLOSURE

[0007] Examples of high octane, low cetane, low carbon intensity and renewal fuel sources include ethanol, methanol, hydrogen, natural gas, ammonia, gasoline, and E10 and El 5 fuel blends.

[0008] An example of a method of mixing controlled combustion in a cylinder with a prechamber includes actively fueling the prechamber with a first fuel injector. The prechamber is sparked with a spark plug operatively connected to the prechamber. Fuel is injected into a main chamber from a second fuel injector. Jet flames are produced that that exit at least one orifice of a nozzle of the prechamber. The jet flames extend into the main chamber. The fuel injected from the second fuel injector into the main chamber is ignited by the jet flames.

[0009] Further examples of the method include the second fuel injector provided on a central axis of the main chamber and the prechamber provided at a side of the main chamber. Theprechamber is at least partially radially external of an internal wall of a crown ring of a piston. The first fuel injector and the second fuel injector arc connected to a common fuel source. A same fuel is injected by the first fuel injector and the second fuel injector. The fuel has low reactivity and high octane. The fuel is selected from E10, E15, E85 and E100 fuel blends. The fuel is selected from ethanol, methanol, hydrogen, natural gas, ammonia, gasoline, and E10 and E15. The first fuel injector fuels the prechamber with between 1% and 7% of a total combustion cycle fuel, and the second fuel injector injects the balance 93% - 99% of the total combustion cycle fuel. The first fuel injector provides a fuel at an equivalence ratio <([>) greater than 1.0. The fuel is provided by the first fuel injector at an equivalence ratio between 1.1 -1.4. The first fuel injector injects fuel at a pressure less than 300bar and the second fuel injector injects fuel at a pressure greater than 500bar.

[0010] In additional examples of the method include the prechamber is actively fueled by the first fuel injector by -40° ATDC. The spark plug sparks the prechamber between -25° ATDC and -5° ATDC, and optionally sparks the prechamber between -20° ATDC and -10° ATDC. The second fuel injector begins injection of fuel into the main chamber between -15° and -1° ATDC, optionally less than or equal to -10° ATDC, less than or equal to -5° ATDC, or less than or equal to -3° ATDC. The spark plug sparks the prechamber at a prechamber pressure below 50 bar. Injecting fuel into the main chamber from the second fuel injector is a first injection of fuel from the second fuel injector. The first injection of fuel includes 5-20% of a total combustion cycle fuel. Subsequently, a second injection of fuel is injected from the second fuel injector into the main chamber. The second injection includes a balance of the total combustion cycle fuel. The first fuel injector fuels the prechamber with between 1% and 7% of the total combustion cycle fuel. The second injection of fuel includes 73%-94% of the total combustion cycle fuel. The spark plug sparks the prechamber between -40° and -20° ATDC and the first injection of fuel from the second fuel injector occurs within 1 to 2 CAD from the spark plug spark. The first injection of fuel from the second fuel injector occurs after the spark plug spark and between -39° and -18° ATDC, and optionally between -35° and -20° ATDC, between -30° and -20° ATDC, between -25° and -20° ATDC, and / or between -25° and -20° ATDC. The first injection of fuel from the second fuel injector is ignited by the jet flames. Combustion of the first injection of fuel from the second fuel injector ignites the second injection of fuel from the second fuel injector.BRTEF DESCRIPTION OF THE DRAWINGS

[0011] Figure 1 depicts an example of a mixing controlled combustion cylinder with a side mounted prechamber as presently disclosed.

[0012] Figure 2 is a perspective view of a prechamber body.

[0013] Figure 3 is a cross-sectional view of a prechamber system.

[0014] Figure 4 is an image of a modeled prechamber jet flames in a main chamber.

[0015] Figure 5 is graph that depicts a basic fueling and spark strategy for ignition assistedMCC.

[0016] Figure 6A is a graph of simulated results for pressure and heat release rate in the prechamber and the main chamber by crank angle for a side mounted prechamber.

[0017] Figure 6B is a graph of simulated results for various prechamber equivalency ratios.

[0018] Figure 7 A is a top perspective view of an exemplary prechamber and main chamber.

[0019] Figure 7B is a side- sectional view of the exemplary depicts prechamber and main chamber of Fig. 7A.

[0020] Figure 8 presents graphs summarizing ignition assistance across various orifice swept angles.

[0021] Figure 9 presents graphs summarizing ignition assistance performance for various total orifice cross-sectional areas.

[0022] Figure 10 presents graphs summarizing ignition assistance performance based for various prechamber volumes.

[0023] Figure 13 depicts a further example of a prechamber with a single orifice.

[0024] Figure 11 is an image of a modeled prechamber combustion.

[0025] Figure 12 depicts an example of a modified direct injector with asymmetric fuel plumes.

[0026] Figure 13 depicts an example a mixing controlled combustion cylinder with two prechambers.

[0027] Figure 14 presents graphs that demonstrate exemplary combustions processes for E10 and El 00 fuel in PC-MCC and diesel fuel in CDC.

[0028] Figure 15 is a flow chart of an exemplary method of prechamber ignition assistance of mixing controlled combustion.

[0029] Figure 16 is a graph that presents the crank angle at the start of combustion.

[0030] Figure 17 presents graphs summarizing the spark timing sweep for tested equivalence ratios across various engine performance metrics.

[0031] Figure 18 illustrates the crank angle timing diagram of a PC-MCC operating strategy with a direct injection pilot.

[0032] Figure 19 presents the pressure progression and heat release rate in both the prechamber and the main chamber for E10 and El 00 fuels.

[0033] Figure 20 is a flow chart of an exemplary method of piloted prechamber ignition assistance of mixing controlled combustion.

[0034] Figure 21 is a timing diagram for unpiloted PC-MCC across exemplary engine operational ranges.

[0035] Figure 22 is a timing diagram for piloted PC-MCC across exemplary engine operational ranges.

[0036] Figure 23 presents graphs representing various combustion performance parameters for E10 and E100 fuels in piloted and unpiloted PC-MCC combustion.DETAILED DISCLOSURE

[0037] Investigation into improved mixing-controlled combustion (MCC) processes for low cetane fuels have been reported by the inventors in Dempsey et al. “A System to Enable Mixing Controlled Combustion with High Octane Fuels Using a Prechamber and High-Pressure Direct Injector” Front. Meeh. Eng. 7:637665 (2021); Dempsey et al. “Prechamber Enabled Mixing Controlled Combustion - A Fuel Agnostic Technology For Future Low Carbon Heavy-Duty Engines,” SAE Technical Paper 2022-01-0449; Zeman et al. “Assessment of Design and Location of an Active prechamber igniter to enable mixing-controlled combustion of ethanol in heavy-duty engines,” International Journal of Engine Research (2023); Zeman et al. “Characterization of Flex- Fuel Prechamber Enabled Mixing-Controlled Combustion with Gasoline / Ethanol Blends at High Load,” ASME Journal of Engineering for Gas Turbines and Power (2024); and Nsaif et al. “Reducing Methane Emissions from Lean Burn Natural Gas engines with Prechamber Ignited Mixing-Controlled Combustion,” ASME Journal of Engineering for Gas Turbines and Power (2024) each of which are incorporated by reference in their entireties.

[0038] Disclosed herein is an active prechamber ignition assistance for a mixing-controlled combustion (PC-MCC) engine and process which enables the MCC of low cctanc / high octane fuels over a wide range of heavy-duty ICE powered operating conditions. MCC engines require fuels with cetane number of ~40 or higher. This corresponds to an approximate research octane number of ~45. Thus, the proposed PC-MCC concept is aimed at fuels with cetane number of less than 40 and octane number greater than 45. Examples of high octane, low cetane, low carbon intensity and renewal fuel sources include ethanol, methanol, hydrogen, natural gas, propane, ammonia, gasoline, gasoline / ethanol fuel blends such as E10, E15, and E85. Examples of the prechamber and combustion design as described herein were tested using the Caterpillar C9.3B off-road engine platform, although it will be recognized that the disclosure herein may be used across a wide variety of ICE platforms and across a variety of low cetane fuels.

[0039] Figure 1 depicts an example of an MCC system 100 having a cylinder 102 and a prechamber system 104. A piston 106 reciprocates within the cylinder 102. A main chamber 114 is defined between the crown 108, the cylinder 102 and a cylinder head 116. The piston 106 top has a crown 108. The crown 108 includes a wall 110 extending annularly about the perimeter of the crown 108 and defines a bowl 112 radially interior of the wall 110. A wall face 111 represents the vertical extent of the piston 106 and is closest to the cylinder head 116. A main chamber 114 is defined between the crown 108, the cylinder 102 and a cylinder head 116. The cylinder head 116 includes an intake valve 118 which controls the flow of fresh air into the main chamber 114. An exhaust valve 120 in the cylinder head 116 controls the flow of exhaust from the main chamber 114. A direct injector 122 extends axially through the cylinder head 116. While an example of a four-stroke engine is shown and described herein, it is recognized that a similar MCC system 100 may be implemented in a two-stroke engine arrangement based upon the present disclosure. Furthermore, in the context of a four-stroke engine, the present disclosure primarily focuses on the compression and power cycles as these pertain to engine combustion.

[0040] The prechamber system 104 includes a prechamber body 124 which is shown in greater detail in perspective view in Figure 2 and in a cross-sectional view in Figure 3. The prechamber system 104 is positioned through the cylinder head 116. The prechamber body 124 defines a prechamber 126. The prechamber system 104 has a nozzle 125 with at least one orifice 128 that opens from the prechamber 126 to the main chamber 114. A cove 148 is formed into the wall 110 of the piston crown 108. The cove 148 extends into the crown 108 from the wall 110.The cove 148 accommodates the nozzle 125 and / or any associated portion of the prechamber body 124 that extends through the cylinder head 116 to expose the at least one orifice 128 to the main chamber 114. Because this portion of the prechamber body 124 is proud of the cylinder head 116, collision between the prechamber body 124 and the piston crown 108 must be avoided. Increasing the space between the crown 108 and the cylinder head 116 would impact engine compression ratio and squish. Instead, the cove 148 accommodates the nozzle 125 of the prechamber system 104 while maintaining operational performance of the engine. The cove 148 extends radially outward from the bowl 112 through the wall 110, but terminates radially interior of an outer circumference of the piston 106 to maintain a continuous exterior perimeter of the piston and the wall 110. The cove 148 may extend into the crown 108 to a maximum depth of the bowl 112, although in other examples, the cove 148 may extend to a level below the wall face 111 and above the depth of the bowl 112. In the context of the disclosure herein, it will be recognized that the nozzle 125 is positioned to promote unobstructed emission of flames from the at least one orifice 128 into the main chamber 114, and that the cove 148 is sized and dimensioned to accommodate the nozzle 125 within a corresponding portion of the wall 110 of the crown 108 of the piston 106.

[0041] Table 1 provides exemplary dimensions and characteristics for configurations of the prechamber.

[0042] It will be recognized that the values presented above are merely exemplary and non-limiting on the scope of prechamber designs as may be used within the present disclosure. As described herein the nozzle may include at least one, but also more than one orifice, including but not limited to 1 or 2 or 3 orifices. Figure 7A is an exemplary top perspective view of a prechamber 126 and main chamber 114. Figure 7 A depicts a non-limiting example of a nozzle 125 with twoorifices 128. The orifice(s) 128 are centered along a line 0 between the nozzle 125 and the direct injector 122. The orifice swept angle (a) generally refers to the angle between the center points of the orifices 128. The line 0 bisects the swept angle (a). In an example of a single orifice, the center point of the orifice would be in alignment with the line 0 and the swept angle (a) is 0°. In addition to swept angle (a), the orifice(s) may also be characterized by an extent angle (co) representing the angle between the outer extents of the orifice(s). For two or more orifices, the extent angle (co) will be slightly larger (related to orifice diameter) compared to the swept angle (a). For a single orifice, the extent angle (co) will be representative of the width (e.g. diameter) of the orifice. While the images may generally depict circular orifice(s), it will be recognized that the shape of the orifice need not be limited to circles and that orifices may exemplarily, but not limited to ovals, ellipse, slots, or rectangles. In examples, the vertical dimension of the orifice(s) may be limited by the prominence of the nozzle 125 from the cylinder head, and therefore, orifice(s) shape may be elongated in a horizontal dimension, which would primarily be reflected in extent angle (co) compared to a circular orifice. Total cross-sectional area is based upon the number and size of the orifices, the cross-sectional area may be between 6-12.5 mm2, between 6.5-12.5 mm2, between 7.6-12.5 mm2, or 9-12.5 mm2an example of 9.1mm2was tested for the present disclosure.

[0043] Figure 7B is a side sectional view of the prechamber 126 and main chamber 114 of Fig. 7A. The included angle is based upon the need to direct the prechamber jet flames within the main chamber 114 radially towards the direct injector 122, but angled below the nozzle of the direct injector but also out of the piston bowl 112. The included angle has a comparatively tighter range or acceptable angles, exemplarily between -10° and -30° below horizontal (represented by the cylinder head), or in another example, between 12.5°-22.5° below horizontal.

[0044] It will be recognized that the size and configuration of the prechamber orifices may be dependent upon the size of the engine cylinder bore for which the prechamber is used. As an example, Table 2 presents example upper and lower ranges for orifice diameter and total orifice cross sectional area for various orifice configurations used with a 115 mm engine cylinder bore.

[0045] As the engine bore size changes, the prechamber jet flame penetration should change with it to adequately provide ignition assistance across the main chamber. The prechamber jet flames penetration is primarily controlled by the prechamber orifice diameter. A turbulent gaseous jet’s penetration scales with the square root of the orifice diameter, as represented by the equation:

[0046] Table 3 below presents an example of how orifice diameter may scale with engine bore size, using the above disclosed example of 115 mm cylinder bore as a baseline.

[0047] The prechamber 126 includes a fuel aperture 130 and a spark aperture 132 at locations in the prechamber 126 at an end generally opposite from the at least one orifice 128. The prechamber body 124 includes an injector cavity 134 which extends away from the fuel aperture 130 and is configured to receive a direct injector 136 therein. The prechamber body 124 includes a spark plug cavity 138 that extends away from the spark aperture 132 and is configured to receive a spark plug 140 therein. The fuel injector 136 and direct injector 122 are both connected to acommon fuel system 142, that common fuel system including fuel which is of high octane, low cetane, and low carbon intensity, including, but not limited to ethanol, methanol, hydrogen, natural gas, propane, ammonia, and gasoline / ethanol fuel blends, such as E10, E15, and E85. The fuel injector 136 is exemplarily an automotive gasoline fuel injector that operates within an exemplary operating range of 50 to 300bar, one non limiting example of such a fuel injector is a 0261500298 GDI fuel injector, available from Robert Bosch Gmbh. The direct injector 122 is exemplarily a diesel fuel injector having an exemplary operational range between 500-2500bar. An example of a direct injector 122 has 6 equally spaced (through 360°) holes that are ~0.2mm in diameter and an umbrella angle of 135°. It will be recognized that direct injectors may include more or fewer holes and exemplary umbrella angles may be between 120° to 160°. A non-limiting example of a direct injector 122 is a 538-7297: C9.3B HRC-A common rail fuel injector, available from Caterpillar, Inc.

[0048] In operation, as will be described in further detail herein, the prechamber system 104 provides ignition assistance to the MCC, resulting in a PC-MCC combustion of low cetane, high octane fuel. The fuel in the prechamber 126 is ignited during the compression stroke. Hot, premixed jet flames 144 of the ignited prechamber fuel exit the prechamber 126 through the one or more orifices 128 into the main chamber 114. These jet flames 144 impinge upon the incoming high pressure direct injected fuel 146 from the direct injector 122 into the main chamber 114 during the compression stroke. Figure 4 is an image of a modeled prechamber combustion and jet flames 144 extending therefrom into the main chamber 114. High pressure (e.g. between 500-2500bar) plumes of direct injected fuel 146 radiate within the main chamber 114 from the direct injector 122. Meanwhile, the jet flames 144 extend from the orifices 128 of the prechamber system 104. The jet flames 144 impinge upon the fuel plumes 146. Depending upon engine and fuel conditions, this impingement either directly ignites the fuel plumes 146 initiating MCC or adds sufficient heat to the main chamber 114 to initiate MCC as the crank advances.

[0049] Figure 11 presents an image of a modeled prechamber combustion and jet flames, similar- to that of Fig. 4, but as a top-down view for a prechamber with a single orifice. The jet flames 144 from the single orifice are centered along line 0 at the direct injector 122, and provides a narrower angle of extent (co). High pressure plumes of direct injected fuel 146 radiate within the main chamber 114 from the direct injector (not depicted). The heat of the jet flame 144 from the prechamber 126 engulfs the high pressure plumes of fuel 146 to provide ignition assistance.

[0050] Figure 5 is graph 150 that depicts a basic fueling strategy of the prechamber and main chamber fuel injectors and initiation of the spark by the prechamber spark plug 140, under a relatively high speed and low load of 5 bar IMEPg at 1800rpm in a Caterpillar C9.3B heavy-duty engine. Further details and variations of this fueling and spark operation will be described herein. In an example, the fueling strategy begins at with fueling the prechamber 126. The prechamber 126 receives between 1-7% of the total stroke fuel amount, with the balance of the amount (93- 99%) being delivered by the direct injector 122 to the main chamber 114 for the main combustion. The total stroke fuel is a function of the energy density of the fuel and the required torque output of the engine. As depicted in Fig. 5, the prechamber is fueled at 152 at a lower injection pressure (e.g. between 10-20MPa) over a longer time (e.g. about 30 crank angle degrees (CAD), between - 50° to -20° after top dead center (ATDC)) compared to the high pressure direct injection of the main combustion fuel at 154, which reaches an exemplary peak of 70MPa during an injection duration between -10° to 10° ATDC crank angle. The prechamber spark is initiated after fueling of the prechamber and during an in an exemplary sparking window 156 of -20° to -5° ATDC crank angle, while other spark windows are contemplated, for example, but not limited to between -30° to -5°. In relation between the timing events, the prechamber fueling 152 occurs before the prechamber is sparked, while the spark 156 occurs between 10° before to 5° after the start of the direct fuel injection 154 into the main chamber. It will be recognized that prechamber and main chamber fueling and spark timing may occur within different windows primarily based upon engine output, with higher output generally requiring advancement of these timing events.

[0051] Figure 6A is a graph 160 of simulated results for pressure and heat release rate (HRR) in the prechamber and the main chamber by crank angle for a side mounted prechamber (SMPC) PC-MCC of E100 fuel and MCC of diesel fuel. As will be described in further detail herein, PC-MCC performance has a noticeable dependency on equivalence ratio at spark timing. Laminar flame speeds and adiabatic flame temperatures peak slightly fuel-rich of stoichiometric and dramatically decrease lean of stoichiometric hence an increase in peak prechamber pressure, increase in peak heat release rate, and slight decrease in burn duration is observed as the prechamber traverses from lean to fuel-rich operation.

[0052] Close inspection of the main chamber HRR for a rich prechamber fueling c[) = 1.1 — 1.4 reveals a non-zero chemical heat release in the main chamber at reference 162 prior to start of injection (SOI) that is sustained until start of combustion (SOC). The same phenomenon occursfor lean and stoichiometric prechamber operation but to a lesser extent. This is better seen in Figure 6B which is a graph 164 similar to the graph 160 of Fig. 6A, but representing lean (e.g. <|) = 0.8), stoichiometric (e.g. <) = 1.0), and rich (e.g. c|) = 1.2) prechamber fuel equivalency ratios. The inset 166 of Fig. 6B shows a magnified version of the chemical heat release at reference 162, with the rich fueling producing the largest heat release, and subsequently earliest SOC at 168. This indicates that premixed prechamber fuel is entrained within the jets and subsequently burns as the jets penetrate the main chamber. The presence of unbumt fuel from the prechamber within the jets is due to incomplete oxidation of the air-fuel mixture and / or premixed fuel being expelled into the main chamber ahead of the flame front due to rapid pressure rise.

[0053] In addition to improved ignition assistance performance, operating the prechamber at fuel-rich conditions may also have benefits in reducing the production of thermal NOx. By having the fuel burn in the prechamber under fuel-rich conditions, this dramatically reduces the thermal NOx production in the prechamber as compared to operating the prechamber fuel-lean or stoichiometric. An investigation of total NOx formation for prechamber-assisted MCC, at high speed and low load at a few selected spark timings (e.g. -7.5° ATDC, -12.5° ATDC, and -17.5° ATDC) and against a diesel baseline, found that NOx emissions can be reduced 30% with an advanced spark timing and fuel-rich prechamber. The 30% reduction in NOx emissions all stems from a NOx reduction in the prechamber. An exemplary prechamber volume of 3 cc and prechamber shot mass of 6 mg was used across all spark timings. By varying the spark timing from -7.5° to - 17.5°ATDC at fixed prechamber fueling, the equivalence ratio at spark timing varies from (|) ~ 0.9-1.2, respectively.

[0054] The PC-MCC configurations produce a distinct, two-stage NOx formation progression attributed to the NOx formation in the prechamber and subsequent NOx formation in the main chamber after the prechamber has ignited the main injection. As spark timing is advanced, the equivalence ratio in the prechamber becomes fuel-rich and results in lower NOx emissions. NOx formation in the main chamber combustion was however found to be similar to that as exhibited during diesel compression ignition (CI) with an abrupt premixed burn spike followed by mixing-controlled combustion. A more sedate and elongated bum within the main chamber would be expected to lower NOx, as is seen from lower NOx formation with decreased injection pressure for diesel CI. A slower bum rates produce less mass at high temperature, which ultimately lowers NOx formation. Thus, the overall NOx formation for PC-MCC has a strong dependence onequivalence ratio within the prechamber at spark timing and the resultant MCC process within the main chamber. PC-MCC operation can thus minimize NOx formation within the prechamber and provide additional ignition assistance by having fuel-rich (i.e. reacting) jet flames with a slightly fuel-rich condition for prechamber combustion (<|) ~ 1.1-1.4). However, it will be recognized that at high loads and / or lower engine speeds, the relatively unchanged main chamber NOx formation may overwhelm the reduction in prechamber NOx formation. However, the benefits of lower emissions at high speed and low load remain.

[0055] In the disclosed PC-MCC combustion systems, the hot jet flames produced from the prechamber igniter interact with the direct injected fuel in the main chamber. In other words, the strength or robustness of a prechamber igniter as an ignition assistant is dependent on how well its heating contributions are utilized to ignite direct injected fuel. Therefore, the prechambers designs are configured to locate the heating contributions of the jet plumes to where the direct injected fuel is going to be. The parameters of orifice total cross-sectional area (e.g. area across all prechamber orifices, along with swept angle a and orifice included angle p, have been found to be important considerations for configuration of a prechamber to locate the jet plumes to the direct injected fuel. As previously mentioned, the orifice swept angle (a) generally refers to the angle between the center points of the orifices, while the orifices may also be characterized by an extent angle (co) representing the angle between the outer extents of the orifice(s). The orifice included angle ( ) represents the angle of the orifices into the main chamber from the horizontal of the cylinder head. Various swept angles (a) were parametrically studied for a two-orifice, 2.08 mm orifice diameter (total cross-sectional area (CSA) = 6.8mm2), 3 cc, SMPC was performed across swept angles between 20° to 50° in 10° increments and over prechamber spark timings of -7.5° to -22.5° ATDC in 2.5° increments. The results of this investigation are reported in the graphs of Figure 8.

[0056] The narrowest two angles of a = 20° and a = 30° were found to produce SOC that compared comparably to the diesel baseline and also to exhibit combustion efficiencies that met or exceeded that of diesel across the tested spark timing range. Comparatively, the a = 40° and a = 50° geometries resulted in misfires across many of the tested spark times. The a = 20° geometry was found to cause the hot jet plumes to impinge upon the piston pip the most and as a result located more of the heat at the direct injector tip. As a result, ignition delays that exceed the diesel baseline were achieved for many of the spark timings investigated. From an ignition assistancestandpoint, this is positive and could be important strategy for cold weather operation and cold starting. The a = 30° geometry was found to locate the prechamber jets to straddle the piston pip in close proximity to the direct injected E100.

[0057] The ignition assistance robustness observed in the a = 20° and a = 30° geometries were observed to dissipate as the included angle is expanded to a = 40° and a = 50°. As the orifice swept angle a increases, the ignition delay begins to increase significantly as the spark timing is advanced. For the a = 50° geometry, the two most advanced spark timings result in near complete misfire. These findings indicate that there is a strong dependence on localized heating near the injector nozzle for sufficient ignition assistance. In other words, the further the prechamber jets are pointed away from the direct injector, the further the direct injected fuel must travel to find the heating contributions from the prechamber jets. This allows more time for the heat from the prechamber jets to dissipate with the surrounding bulk gas and main chamber surfaces, resulting in lessened ignition assistance.

[0058] With reference to Fig. 8, while diesel-like combustion efficiency was achieved for all swept angle geometries at late spark timings, combustion efficiency fell off as spark timing advanced for the wider swept angle geometries. It is observed that for the most retarded spark timing of -7.5°ATDC, all the geometries produce ignition delays similar to diesel combustion (within 3 CAD) and combustion efficiencies over 99%. The larger swept angle geometries begin to produce longer ignition delays for the more advanced spark timings and combustion efficiency begins to drop rapidly. Based on the performance of the narrow angle two-orifice SMPC’s in this investigation, in SMPC embodiments with more orifices, for example with three or four orifices, swept angles of 50° and 60° across the orifices, respectively, which are the narrowest orifice swept angles that were geometrically and physically possible. The narrow angle two-orifice SMPC’s provide better ignition assistance than the three and four orifice SMPCs with wider alpha angles, as the heat form the outer prechamber jet flames is not utilized well, as the heat needs to be direct towards the direct injector.

[0059] In addition to the swept angle of multiple orifices, the prechamber orifice diameter is a key design parameter as it dictates various performance aspects including prechamber jet velocity, jet penetration, pressure rise rate within the prechamber during combustion, and scavenging effects. An investigation of the impact of orifice diameter on PC-MCC performance was conducted by varying the total cross-sectional area of the orifices in SMPC prechamber igniterconfigurations from 4.5 to 9.1 mm2while maintaining the prechamber volume at 3cc against a diesel baseline. For the investigation, spark timing of each prechamber was swept from -7.5° to - 20°ATDC in 2.5° increments. A summary of ignition assistance performance for the total orifice cross-sectional areas in SMPC prechamber designs are presented in the graphs of Figure 9.

[0060] As shown in Fig. 9, the orifice diameter has a noticeable impact on prechamber pressure during compression and combustion. The differences between main chamber and prechamber pressure during compression and expansion indicate that smaller orifices are more constrictive and thus have higher dissipative gas exchange losses across each chamber. The constrictive nature of smaller orifices will also inherently scavenge poorly and promote higher concentrations of combustion residuals within the prechamber from cycle-to-cycle. During combustion, the constrictive smaller orifices increase the backpressure on the prechamber and thus elongate the rise and fall of prechamber pressure relative to the main chamber pressure. These differences in prechamber pressure history will alter the rate at which combustion products evacuate into the main chamber. In other words, the duration or lifetime of the prechamber jet flames will increase as orifice diameter decreases. The larger 9.1mm2area orifice prechambers were observed to be the first (e.g. prior to -8°ATDC) to produce a jet flame penetrating the main chamber (with a spark timing of -12.5° ATDC) but was almost completely exhausted near - 2°ATDC. The smaller orifice cross section of 4.5mm2(with a spark timing of -12.5° ATDC) was observed to not produce an appreciable jet as quickly (e.g. -6°ATDC) but is driven by expanding gases within the prechamber up to and beyond TDC. Despite the larger orifice jets exhausting quickly, the jet diameter was observed to be larger and thus containing more mass at high temperature. This leads to more localized heating and requires more time to mix-out with the surrounding in-cylinder gases. Larger orifice cross sectional area was also observed to result in higher jet core temperatures and sustain higher jet temperatures and velocities at a further distance away from the orifice exit. These observed jet characteristics help explain the superior ignition assistance performance of the larger orifice diameter prechambers with close-coupled spark timings, again as illustrated in the graphs of Figure 9.

[0061] For the SMPC igniter configuration, the start of combustion (SOC) generally continually advanced with spark timing for all orifice sizes considered. The direction of the SMPC jets toward the direct injector increases the residence time of localized heating near- the directinjector. The SMPC igniter configuration produced localized heating contributions near the main injector nozzle before, during, and after start of injection (SOI), regardless of orifice diameter.

[0062] Referring to the summary graphs of Fig. 9, the disclosed prechamber enabled MCC produces diesel-like GIE and gross indicated specific NOx emissions. Despite PC-MCC cases having slightly higher GIE relative to diesel in this investigation, it is thought that this is primarily due to the advanced CASO for PC-MCC. It is expected that the diesel baseline would have the same or slightly higher indicated efficiency relative to PC-MCC at an equivalent CA50 timing. This is due to the expected heat transfer and dissipate gas exchange losses associated with addition of the pre- chamber.

[0063] NOx emissions appear to peak near a spark timing of -10° ATDC, which corresponds to the prechamber being fired with a near stoichiometric mixture in the prechamber volume. As spark timing advances, the equivalence ratio in the prechamber increases to fuel-rich conditions and total engine-out NOx emissions simultaneously decrease. Similarly, the reduction in NOx from a spark timing of -10° ATDC to -7.5° ATDC suggests that lean operation of the prechamber will also subsequently reduce NOx emissions. It is well understood that thermal NOx production peaks at slightly fuel-lean e.g. (|) ~ 0.9 conditions, thus targeting a prechamber equivalence ratio that is fuel-rich with <]) ~ 1.1 to 1.4 is expected to reduce NOx emissions in the prechamber, while also providing the most robust ignition assistance in the main chamber.

[0064] In the PC-MCC system as described herein, the size of the prechamber volume dictates how much fuel energy is able to be provided for ignition assistance. In the PC-MCC system prechamber volume would be one that is able to provide robust ignition assistance over the desired operating range but not be oversized, because as prechamber volume increases, heat transfer losses will increase and scavenging efficiency will decrease. Prechamber volumes for the PC-MCC system were investigated to minimize these penalties. For both mounting configurations, the prechamber volume is evaluated at 2, 3, and 4 cc while maintaining the total orifice cross sectional area at 6.8 mm2with a main chamber clearance volume of 103.2 cc while maintaining an effective compression ratio of 16.0:1. Each prechamber configuration is evaluated at spark timings of -7.5 to -20°ATDC in 2.5° increments. Figure 10 presents graphs summarizing the ignition assistance performance based upon prechamber volume resulting from this investigation.

[0065] The prechamber volume impacts prechamber pressure during compression and combustion. As the volume increases, the observed pressure difference between main chamber andprechamber during compression and expansion increases as well. This is due to the increase in mass that is transferred between each chamber, hence it is expected that there arc higher dissipative gas exchange losses and higher concentrations of residuals within the prechamber from cycle-to- cycle for larger volumes. During combustion, the larger volumes have more charge mass to bum and therefore more combustion residual mass to evacuate. At a fixed orifice cross-sectional area across all volumes considered, the evacuation of combustion residuals is elongated for the larger volumes, resulting in a slower decay in prechamber pressure after the peak.

[0066] The lesser fuel energy in the smaller volume is observed to retard the development of an appreciable jet and the resulting jet is completely exhausted prior to complete exhaustion of the larger combustion volume. The additional prechamber fuel energy in the larger volume enables higher internal jet core velocities that are sustained at further distances away from the orifice exit. Increase in prechamber volume also sustains higher jet temperatures at further distances from the orifice exit due to the continual supply of hot combustion products from the prechamber.

[0067] As summarized in Fig. 10, the largest prechamber volume investigated (4cc) produced the best ignition assistance, as it produces the most advanced SOC’s and CA50 timings for nearly every spark timing considered whilst maintaining combustion efficiencies above 99%. However, the gross thermal efficiency decreases as prechamber volume increases due to the expected heat transfer and dissipative gas exchange losses across both chambers. However, the SMPC demonstrates less sensitivity to prechamber volume, with minor incremental improvements in CA50 and SOC achieved with increases in volume. Because of this, an advantage tradeoff between the ignition assistance from the larger volume (4cc) to instead use a small volume (~2cc) SMPC, which produced the highest indicated efficiency, favorable CA50 timings, robust combustion that is not sensitive to prechamber spark timing, and NOx emissions that are lower than conventional diesel combustion. In comparison to Figure 9, a similar' response in NOx emissions levels to the equivalence ratio of the prechamber at spark timing is observed. This is indicative that operation of the prechamber at fuel-rich conditions may be favorable for reducing NOx emissions but may introduce heightened CO and UHC emissions as prechamber volume increases possibly due to the crevices present within the prechamber geometry and increased surface area of the prechamber walls.

[0068] Figures 12 and 13 provide additional examples of further optional implementations of the PC-MCC system as described herein. Fig. 12 depicts an example with a modified direct fuelinjector 122A with the orifices of the direct fuel injector 122A limited to an angle of arc biased towards the side mounted prechamber. In examples, this may be 180° degrees, or between 135°- 225° of arc. Typical direct fuel injectors evenly distribute the high pressure plumes of direct injected fuel 146 about the axis of the direct injector. This improves fuel distribution and mixing, however, with the ignition assistance heat from the prechamber biased to one side, focusing the fuel density in the main chamber to the region of maximized heat from the prechamber jet flames 144. This increases the high pressure plumes of fuel that directly interact with the prechamber jet flames 144. Figure 13 presents an embodiment that achieves similar improved jet flame - fuel plume interaction by locating two opposed prechambers to direct jet flames from either opposing sides towards the high pressure plumes of direct injected fuel. These examples may promote ignition assistance and improve combustion at advanced spark timings.

[0069] It will be recognized that a side mounted prechamber (SMPC) embodiment as shown and described with respect to Fig. 1 has been disclosed herein, while it will be recognized, based upon the present disclosure, that features as described herein may be implemented in other prechamber arrangements including a centrally mounted prechamber (CMPC).

[0070] As previously noted, examples of the PC-MCC system as described herein may be used with other low cetane, high octane fuels. One such additional fuel may include natural gas (NG). Natural gas is an attractive as a lower CO2 emission fuel. One challenge of natural gas ICE operation is methane slip. Methane slip refers to the unbumed methane emissions that escape from the combustion cylinder. However, because methane is a more potent greenhouse gas, unless methane slip is reduced to near zero levels, natural gas does little to address net effective GHG emissions. Incomplete combustion is a common cause of methane slip in NG SI engines, therefore, existing NG SI engines run globally lean. However, even with lean operation, methane slip is still a limiting consideration for adoption and use of NG SI engines. However, it has been determined that the PC-MCC systems as disclosed herein provide improved combustion efficiency of NG over NG SI combustion, resulting reduced methane slip. Testing results showed an approximate lOx reduction in methane slip (~0.1 g / kW-hr versus ~2 g / kW-hr) with the PC-MCC system. It is also noted that, like the investigations described above, the size of the prechamber orifices and spark timing held the greatest measured influence on combustion efficiency and methane slip reduction. Advanced spark timings (e.g. -20° ATDC or -15° ATDC compared to -10° ATDC) and greater overall cross-sectional area (e.g. 9.1 mm2) resulted in less methane slip.

[0071] When an engine is operated globally stoichiometrically or fuel-rich, there is no oxygen left over in the exhaust. The consumption of the available oxygen during combustion is associated with lower overall GHG formation in exhaust. This enables the use of low cost 3-way catalyst after treatment devices to meet emissions standards. However, diesel engines today are operated globally lean, which leaves excess oxygen in the exhaust stream, requiring complicated and expensive after treatment devices because of this excess oxygen. Stoichiometric operation of diesel engines is currently not possible because at stoichiometric operation, diesel engines emit significant amounts of soot, at levels where diesel engines cannot operate stoichiometrically within emissions standards. However, as a further advantage of the PC-MCC system as described, diesellike performance is achieved with the use of low sooting fuels and can be operated globally stoichiometrically, enabling the use of the low cost 3-way catalysis after treatments commonly available to SI engines while exhibiting diesel-like performance.

[0072] The above-described prechamber-enabled mixing-controlled combustion (PC- MCC) system has been found to present a commercially viable solution to providing diesel-like MCC performance with high octane / low cetane fuels, for example, but not limited to: ethanol (E100) and ethanol-gasoline blends (e.g. E10 or E15). In addition to the system described above, refinements to the combustion process as described herein have been found to further improve performance of the PC-MCC system, performance range of the PC-MCC system, and / or reduce emissions thereof.

[0073] Figure 14 presents graphs demonstrating the rate-limited combustion process of E10 and E100 fuel in a PC-MCC system with diesel fuel in a conventional diesel combustion (CDC) system, at a comparatively high load of ~18 bar brake mean effective pressure (BMEP) and 2200 rmp in a Caterpillar C9.3B heavy-duty engine, as compared to the low load example with respect to the description above with respect to Fig. 5. With the higher load, the fuel injection to the prechamber was advanced to between -80° and -40° ATDC and spark timing was advanced to an exemplary window between -34° to -14°, with a specific spark timing of -14° ATDC used for the graph (for E10 and E100). As shown in these graphs, the air-fuel mixture in the prechamber is quickly consumed after sparking and the small volume of the prechamber experiences a sudden large rise in pressure. The pressure difference between the main chamber (MC) and the prechamber (PC) produces hot jet flames that emanate from the PC towards the fuel direct injected into the main chamber. The jets impinge upon the fuel and the fuel is readily ignited with E10 and E100having nearly identical start of combustion (SOC) timing that is within 1 crank angle degree (CAD) of the diesel baseline. Despite the very similar ignition delays, the premixed spike portion of the E10 and E100 bum is notably larger than with diesel. This is due to the higher volatility of E10 and E100 which promotes more fuel to bum in a premixed phase at SOC, especially in the vicinity of the large hot jets provided near the injector nozzle by the prechamber. Despite this different HRR in the PC-MCC system, the overall pressure curves hare highly similar compared to the comparative diesel CDC system. The post-spike local HRR maxima for each fuel corresponds to the lower graph of the differences in fuel energy delivery. The example shown in Fig. 14 maintained the same PC-MCC system, therefore, the differences in energy density between the fuels (diesel fuel being more energy dense than E100) resulted in varying CAD durations for the delivery of the same energy (e.g. -6800 J).

[0074] Figure 15 is a flow chart of a method 200 of prechamber ignition assistance of mixing controlled combustion. It will be recognized that the method 200 can be carried out using any of the systems 100 as described above with respect to Figs. 1-13 or any other suitable system as may be recognized in view of the present disclosure. As disclosed elsewhere, it will be recognized that the timing of fuel injection and spark will vary based upon engine output with higher loads requiring comparatively advanced timings compared to lighter loads. The following description is exemplarily made with the high load of - 18 bar- BMEP and 2200 rmp in a Caterpillar C9.3B heavy-duty engine conditions as discussed above.

[0075] At 202 the prechamber 124 is actively fueled with the fuel injector 136. The prechamber is provided with between about 2-10% of the total cycle fuel, with the remainder being delivered to the main chamber 114 as will be described in further detail herein. While highest flame temperatures are achieved with a stoichiometric mixture, it has been found that a fuel-rich prechamber with an equivalence ratio of <|) > 1.0, and exemplarily an equivalence ratio of > = 1.1 - 1.4 produces more favorable combustion within the PC-MCC. Fueling of the prechamber exemplarily is completed by about -40° ATDC, although in other examples may extend until about -40° ATDC. As noted above, the delivery of fuel containing a specified amount of energy is dependent upon the energy density of the fuel being delivered, the pressure from the fuel injector, and the duration of the fuel injection. Under otherwise fixed conditions a less energy dense fuel will require a longer injection than a more energy dense fuel to deliver the same energy capacityto the prechamber. Therefore, the fuel injection may exemplarily take between 20-40 CAD to complete.

[0076] After the prechamber is fueled, the spark plug 140 is activated at 204 to spark the prechamber. Experimental results have found that under some conditions, spark timings between -30° ATDC and -0° ATDC, -25° ATDC and -5° ATDC, or between -20° ATDC and -10° ATDC may be used. In still other examples, spark timings of -8°, -10°, -12°, -14°, -16°, -18°, or -20° may also be used, or any other value within the ranges disclosed.

[0077] After the prechamber is sparked at 204, then at 206, the direct injector 122 stalls the direct injection of the remaining balance of the cycle fuel into the main chamber 114. This produces high pressure plumes of direct injected fuel 146 that radiate outward from the (e.g. centrally located) direct injector 122 into the main chamber 114. Direct injection of the remaining balance of the cycle fuel into the main chamber 114 may exemplarily initiate within 1-5 CAD, within 1-3 CAD, or within 1-2 CAD of the spark of the prechamber at 204. Exemplary initiation of the direct injection of the remaining balance of the cycle fuel may occur within a crank angle range of between -15° to +5° ATDC, later than or equal to -10° ATDC, later than or equal to -5° ATDC, or later than or equal to -3° ATDC, for low engine load conditions. At higher engine loads, the direct injection timing typically needs to be advanced to deliver the larger quantity of cycle fuel, with direct injection of the remaining balance of the cycle fuel into the main chamber 114 may initiate within exemplary and non-limiting crank angle ranges of between -30° to 0° ATDC, -25° to -5° ATDC, -25° to -5° ATDC, or -20° to -10° ATDC.

[0078] Sparking of the prechamber at 204 causes the fuel in the prechamber to ignite causing a release of heat in the prechamber and, given the small volume of the prechamber, a rapid increase in the pressure within the prechamber. At 208, this heat and pressure in the prechamber produces at least one jet flame 144 out of at least one orifice 128 of the prechamber nozzle 125 that extends out into the main chamber 114. The at least one jet flame 144 extends in the direction of the high pressure plumes of direct injected fuel 146. As noted above, because the prechamber is fuel-rich, the excess fuel in the at least one jet flame extends the burn of the at least one jet flame promoting contact of the jet flame with the plumes of fuel.

[0079] Subsequently at 210, the at least one jet flame 144 begins combustion of the direct injected fuel in the main chamber 114. This combustion occurs before TDC in the stroke cycle andproduces a large initial spike in HRR as the comparatively high volatility fuel is ignited at the start of combustion. The combustion process may overall take between 30-60 CAD to complete.

[0080] Relative to the other timing in the process above, the start of combustion at 210 starts at the end of the compression stroke or at the beginning of the power stroke at 212. It will be recognized that the direct fuel injection and the combustion of said fuel will extend into the power stroke. Further, under low engine load conditions, the start of combustion at 210 may occur early in the power stroke, while under heavy load conditions, combustion will start during the compression stroke.

[0081] Figure 16 is a graph that presents the crank angle at the start of combustion (SOC) as represented by 5% total fuel mass fraction burn (CA5) for the PC-MCC system and operation as disclosed herein as compared to diesel CDC, and piloted and un-piloted MCCI combustion with E10 (gasoline / ethanol blend) and El 00 (pure ethanol), in comparison to changes in intake valve close (IVC) temperature. The prechamber spark timing was exemplarily -20° ATDC and a prechamber equivalence ratio of <[) = 1.3 at prechamber spark timing. The piloted and un-piloted MCCI combustion is comparative technology to PC-MCC, where the engines in-cylinder air intake valve closure temperature is elevated to achieve ignition with the low cetane E10 and E100. As shown in the graph, while piloted and un-piloted MCCI combustion of E10 and E100 fuels is possible, it is highly temperature-sensitive, which presents a challenge as it is hard to control. The PC-MCC system and operation as disclosed herein exhibits low temperature sensitivity, approximating that of diesel CDC, which is a robust, reliable combustion system that is does not require complex intake air temperature controls.

[0082] The equivalence ratio of the prechamber fuel mixture at spark timing dictates the fuel energy input for the jet flames, the flame speed, and the adiabatic flame temperature. A spark timing sweep was used to characterize the implications of prechamber fuel mixture equivalence ratio on PC-MCC ignition assistance performance. Spark timings from 14-29° BTDC and equivalence ratios of 0.7 (fuel-lean), 1.0 (stoichiometric), and 1.3 (fuel-rich) at spark were used for E10 and E100 fuels. While an equivalence ratio of <|) = 1.3 exhibited similar’ ignition delays between E10 and El 00 fuels and similar to that of CDC, as the equivalence ratio decreased, the ignition delay for both E10 and El 00 increased with El 00 showing greater sensitivity to decreased equivalence ratio.- l-

[0083] Figure 17 presents graphs summarizing the spark timing sweep for the tested equivalence ratios across various engine performance metrics. Ignition assistance performance was found to be strongest (using CA5 as SOC) at the most retarded spark timing of -14° ATDC independent of fuel or equivalence ratio. As spark timing was advanced, the phasing of the heating contributions provide by the prechamber jet flames also consequently advanced, resulting in weaker interactions between the jet flames and the direct injected fuel in the main chamber due jet cooling / dissipation. Despite the range of SOC timings and combustion phasing (with reference to CA50), the volatile nature of E10 and E100 fuels enabled near complete oxidation (combustion efficiency > 99%) in all cases depicted. Further SOC scales with the equivalence ratio at all park timings considered, however, advancement of the spark timing at a fuel lean equivalence ratio of (|) = 0.7 beyond -20° ATDC resulted in complete misfire for E100. E100 fuel was also sensitive to spark timing at a stoichiometric equivalence ratio of <|) = 1.0. This further supports the conclusion of the enhanced sensitive to equivalence ratio of E10 and E100 fuel in PC-MCC systems and operation.

[0084] Within the disclosed PC-MCC and operation, both E10 and E100 are lower, but within about 5% of the diesel baseline gross indicated thermal efficiency (GIE). This is generally expected (slightly lower GIE at identical CA50 timings for PC-MCC) due to the heat transfer and dissipative gas exchange losses associated with the prechamber. E100 achieves later CA50 timings due to elongated rate of energy delivery, but within fuel type CA50 timings are within 1CAD for each fuel type. The CA50 timing advances with spark timing advance, which may be due to the elongated ignition delays that produce large, premixed burns near TDC. The intensity of the premixed burn may shorten the time delay between the transition from premixed burn to ratelimited burn and thus advances CA50. GIE increases as the total fueling requirement decreases and the CA50 timing advances.

[0085] As noted above, NOx emissions decrease in the disclosed PC-MCC systems and operation with a rich prechamber (e.g. equivalence ratio <]) = 1.1-1.4). This is due to reduced prechamber NOx emissions associated with a rich premixed burn and reduced MC NOx achieved with reduced premixed bum spike severity by the strong ignition assistance qualities achieved with rich prechamber operation. However, peak load PC-MCC NOx levels presented in Fig. 20 do not demonstrate a sensitivity to equivalence ratio, rather to fuel selection with E100 producing notably lower NOx than E10 or baseline diesel, likely as under peak loads, the MCC NOx outputovershadows the PC NOx production / reduction due to the substantial difference between the volumes of the fuels burned during these combustion phases. Therefore, the rich PC operation has the benefit of mitigation NOx emissions at light load while demonstrating strong ignition assistance capabilities at both low and high loads. E100 is believed to produce lower NOx emissions due to considerably lower HRR during premixed and rate-limited burn representing a lower intensity combustion.

[0086] The investigation represented in the graphs of Fig. 17 demonstrates when under a heavy load, at late spark prechamber spark timings, e.g. less than -20° ATDC, produce strong ignition assistance, but with reference to Fig. 14, occur at pressures exceeding -50 bar, and result in elevated max prechamber pressures (Fig. 17). This is compared to the comparatively light load operation depicted in Fig. 6A, where baseline pressure in the prechamber remains below -50 or -60 bar across a wide range of crank angle spark timings. Excessive prechamber pressure e.g. >250 bar can be detrimental to prechamber structural integrity and component health. Furthermore, spark plug electrode health is greatly impacted by pressure at the time of spark discharge. When repeatedly fired in excess of -50 bar, spark plug life is known to deteriorate and fail prior to the recommended service interval. In the PC-MCC system design, this is particularly a concern for operation at late spark timings as the heightened compression ratios relative to standard SI combustion will inherently increase the nominal near-TDC spark discharge pressures, especially at boosted conditions. Therefore, it is observed that for spark timings later than -20°ATDC result in heightened prechamber pressures that would adversely effect the operational life of the spark plug. Thus an operationally viable system for MCC pressures requires a spark timing that ensures that the spark in the prechamber occurs at a prechamber pressure less than -60 or -50 bar, and in further examples, less than 50 bar.

[0087] Therefore, to avoid excessive wear on the spark plug, the spark timing in the prechamber must be advanced to occur prechamber pressures less than -60 or -50 bar, and in further examples, less than 50 bar. However, as previously described, advancement of the prechamber spark timing results in less fuel energy available for ignition assistance with weaker prechamber jet flame - fuel plume interaction. This further delays ignition. Figures 18-20 present a further operational variation to that as shown and described with respect to Figs. 5, 14, and 15. With the spark timing advanced beyond -22 / 23° ATDC to limit prechamber pressure at sparking to nominally below 50 bar, a pilot fuel injection from the direct injector 122 into the main chamber114 is added in order to prolong the ignition assistance to main fuel injection occurring closer to TDC. In effect, the prechamber jet flames provide an ignition assistance to the pilot injection, which in turn ignites the main injection. Depending upon the reactivity of the fuel being used the prechamber jet flames may completely ignite the pilot or serve as a catalyst to promote accelerated spray breakdown an kinetically controlled fuel decomposition, this promoting ignition of the main injection.

[0088] One advantage observed of this arrangement is that while in application the volume of the prechamber is structurally fixed and the resulting variation in fuel mass provided to the prechamber is thus limited to little variation relative to the balance of the fuel in the main chamber, the direct injected pilot fuel plumes are not so limited, thus providing a wider range of fuel energy input to be delivered with the pilot. This provides a mechanism for direct control of the amount of ignition assistance provided.

[0089] Figure 18 is a crank angle timing diagram 250 of a PC-MCC operating strategy with a direct injection pilot. The timing diagram 250 presents fuel injections for E10 and E100 as well as diesel for comparison. Under the heavier load, the prechamber fuel injection 252 is advanced to exemplarily between -80° to -40° ATDC. The prechamber fuel injection may be between 1-7% of the total fuel mass. The spark window 254 was exemplarily between -38° and - 25° ATDC. The In one experiment, the pilot injection was maintained at 9.1% of total fuel mass for both E10 and E100 investigations. In examples, the pilot injection may be between 5-15%, 10- 15%, 5-10%, or 10-20%. The pilot start of injection (SOI) 256 timing was swept from 21.5 to 29.5° BTDC. For each pilot timing, the spark timing was maintained at 4.5 CAD advanced of the pilot SOI. The main injection 258 was maintained at 11.5° BTDC and included the balance 84% - 90% of the total fuel mass. As previously described, at the same injection pressure, E100 requires a longer injection to deliver a comparable fuel mass compared to E10 or diesel.

[0090] Figure 19 presents the pressure progression and heat release rate (HRR) in both the prechamber and the main chamber for E10 and E100 fuels, a pilot arrangement with a spark at 30° BTDC with a non-pilot arrangement with a spark at 20° BTDC. As depicted in the graph, the pilot injection MCC process further improved the correspondence of the combustion towards the baseline diesel combustion, bringing SOC to within < 1 CAD while simultaneously reducing the peak prechamber and spark discharge pressures.

[0091] Figure 20 is a flow chart that depicts an exemplary method 300 of prechamber ignition assistance of mixing controlled combustion with a pilot. The method 300 may be performed similarly to the method 200 and details previously described and may be carried out using any of the systems 100 as previously described wherein or any other suitable systems as may be recognized in view of the present disclosure.

[0092] At 302 the prechamber is actively fueled with the fuel injector 136. As previously noted, one of the purposes of the method 300 is to advance the spark timing while maintaining ignition assistance for MCC. The advanced spark timing thus may require an advanced prechamber fueling inasmuch as the prechamber must be fueled with the prechamber fuel fraction prior to the advanced spark. In examples, the prechamber fueling may exemplarily occur between -80° ATDC and -40° ATDC or at least prior to spark. The prechamber is fueled with exemplarily 1-7% of the total combustion cycle fuel.

[0093] At 304, the spark plug 140 in the prechamber is activated, sparking the fuel in the prechamber. As noted, this occurs prior to the pressure in the prechamber reaching the threshold of 50 bar or 60 bar which is known to reduce spark plug longevity. In further examples, the spark may occur at a prechamber pressure between 30-50 bar. At exemplary heavy load conditions tested such prechamber pressure ranges may occur in advance of -26° ATDC, but variation in crank angle will occur based upon overall system design and operating load. In examples, spark timings between -26° and -34° ATDC were tested as described herein, and spark timings may exemplarily occur between -20° and -40° ATDC.

[0094] With the advanced spark timing, the pilot combustion is needed to extend the ignition assistance as described above. The direct injector 122 is operated at 306 to produce a pilot injection of fuel into the main chamber 114. The pilot injection is between 5-20% of the total combustion cycle fuel. In examples, the pilot injection may be between 5-15%, 10-15%, 5-10%, or 10-20% of the total combustion cycle fuel. In the example described, the pilot injection is 9.1% of the total combustion cycle fuel. In examples, the pilot fuel injection starts 1° or 2° CAD after the spark. Therefore, the pilot may exemplarily occur between about -39° and - 18° ATDC. In other examples, the pilot fuel injection occurs between exemplary ranges of between -34° and -24° ATDC, between -30° and -20° ATDC, or between -25° and -20° ATDC.

[0095] The combustion of the prechamber fuel progresses from the spark at 305 causing a release of heat in the prechamber and, given the small volume of the prechamber, a rapid increasein the pressure within the prechamber. At 308, this heat and pressure in the prechamber produces at least one jet flame 144 out of at least one orifice 128 of the prechamber nozzle 125 that extends out into the main chamber 114. The at least one jet flame 144 extends in the direction of the plumes of fuel of the pilot injection.

[0096] At 310 the pilot fuel is ignited by the jet flames from the prechamber. As noted above, because the prechamber is fuel-rich, the excess fuel in the at least one jet flame extends the bum of the at least one jet flame promoting contact and of the jet flame with the pilot fuel and ignition of the pilot fuel. The pilot fuel ignition extends the bum of the ignition assistance beyond the length of time of burn of the jet flames from the prechamber.

[0097] At 312, the direct injector 122 starts the direct injection of the remaining balance of the cycle fuel into the main chamber 114. This produces high pressure plumes of direct injected fuel 146 that radiate outward from the (e.g. centrally located) direct injector 122 into the main chamber 114. Direct injection of the remaining balance of the cycle fuel into the main chamber 114 may initiate within a crank angle range of between -15° to +5° ATDC, between -15° to 0° ATDC, between -15° to -5°, or between -15° to -10°, in the example described herein the direct injection of the balance of the cycle fuel occurred at -110ATDC. In examples, the balance of the cycle fuel is between 94% and 73% of the total combustion cycle fuel. Because the pilot fuel is already ignited within the main chamber, as the balance of the cycle fuel is injected, the balance of the cycle fuel is readily ignited at 314 with the flames of burning pilot fuel. This ignition is exemplarily 2° CAD or less from the start of the direct injection at 312. As previously noted, ignition of the balance of the cycle fuel may occur at the end of the compression stroke or the beginning of the power stroke, while the injection of the balance of the combustion cycle fuel and combustion thereof extends into the power stroke. The complete injection of the balance of the fuel may take about 30° to 50° CAD, depending upon the amount of fuel to be injected and the energy density of the fuel. As noted, the ignition of the fuel at 314 occurs at the end of the compression stroke or the beginning of the power stroke, and at 316 the piston completes the compression stroke and / or continues the power stroke.

[0098] While the flow charts of Figs. 15 and 20 describe respective examples of piloted and unpiloted versions of the PC-MCC operation. It will be recognized that the specific timings of fuel injection and spark are first dependent upon the specific fuel used and the physical configuration of the engine. Furthermore, during operation of any given engine and selected fuel- l-type, fuel injection and spark timings are adjusted during engine operation to achieve intended engine outputs and operating loads. Because these combustion strategics arc fuel agnostic to low cetane, high octane fuels and may be implemented with the same cylinder arrangement, engine operation may use unpiloted PC-MCC operation under lower load demands and piloted PC-MCC operation in response to higher load demands. Figures 21 and 22 present more broad timing diagrams giving a larger operational timing ranges as may respectively be used for the methods as described with respect to Figs. 15 and 20. It is recognized that in the diagrams of Figs. 21 and 22, that prechamber spark and SOI of the prechamber fuel typically occur within about 1-5 CAD, prechamber and main chamber fuel injection occur over relatively larger CAD durations, similarly main chamber MCC combustion occurs over larger CAD durations. The timing windows of Figs. 21 and 22 exemplarily represent a window in which the event starts and completes over expected operational conditions.

[0099] Figure 21 presents a timing diagram for unpiloted PC-MCC operation, where the prechamber jet flames are used to directly ignite the single direct fuel injection into the main chamber, which occurs around the top dead center (e.g. about -30 deg ATDC to +30 deg. ATDC). The unpiloted strategy would exemplarily be used at low- (1 to 5 bar IMEP) to mid- (5 to 10 bar IMEP) engine loads where prechamber spark ignition is possible -30 deg ATDC to +20 deg. ATDC. Figure 22 presents a timing diagram for pilot PC-MCC operation, where the prechamber jet flames are used to directly ignite pilot injection into the main chamber, the pilot injection exemplarily contains 5% to 20% of the total direct injected fuel, and occurs before the main direct injection of the balance of the combustion cycle fuel in the main chamber. The piloted PC-MCC strategy may exemplarily be used at high engine loads (>10 bar IMEP) where prechamber spark ignition is challenging around top dead center due to high prechamber gas pressures. Thus, the prechamber spark timing window is advanced to -50 deg. ATDC to -20 deg. ATDC, to ignite the pilot direct injection, and avoid high prechamber pressures.

[0100] Figure 23 presents graphs representing various combustion performance parameters for E10 and E100 between -34° and -14° ATDC for various fuel equivalence ratios in piloted and unpiloted PC-MCC combustion. For both E10 and E100, the addition of the pilot enabled the advancement of the prechamber spark timing range, to a point of lower pressure in the combustion cycle, as shown in Fig. 21, while maintaining strong ignition assistance of the main fuel injection while also reducing the peak prechamber and spark discharge pressures. For bothfuels, the peak prechamber pressure is greatly reduced and thus peak pressure values within the prechamber-main chamber system arc governed by the peak main chamber pressure. For the spark timing range considered, the spark discharge pressure was held within 33-45 bar, and less than the 50bar threshold recognized for spark plug performance. The CA50 timing advances in the presence of the pilot injection. Comparing the HRR between the piloted and unpiloted PC-MCC strategies shows that the addition of the pilot shortens the transition from premixed to rate-limited bum, resulting in the advanced CA50. As with the unpiloted strategy the SOC trend of the piloted strategy is also dependent on the prechamber equivalence ratio and spark timing. The influence of PC equivalence ratio on ignition assistance indicates that the pilot’s ignition and subsequent burn are dependent on the phi-dependent combustion characteristics and fuel energy input from the prechamber. For E100 fuel the piloted and unpiloted injection strategies show a strong sensitivity to lean prechamber operation. For the piloted injection strategy, the main injection misfired for spark timings advanced of -30° ATDC. However, the piloted injection strategy enabled lean E100 PC-MCC operation at advanced spark timings not available in the unpiloted injection strategy.

[0101] In the above description, certain terms have been used for brevity, clarity, and understanding. No unnecessary limitations are to be inferred therefrom beyond the requirement of the prior art because such terms are used for descriptive purposes and are intended to be broadly construed. The different systems and method steps described herein may be used alone or in combination with other systems and methods. It is to be expected that various equivalents, alternatives, and modifications are possible within the scope of the appended claims.

[0102] The functional block diagrams, operational sequences, and flow diagrams provided in the Figures are representative of exemplary architectures, environments, and methodologies for performing novel aspects of the disclosure. While, for purposes of simplicity of explanation, the methodologies included herein may be in the form of a functional diagram, operational sequence, or flow diagram, and may be described as a series of acts, it is to be understood and appreciated that the methodologies are not limited by the order of acts, as some acts may, in accordance therewith, occur in a different order and / or concurrently with other acts from that shown and described herein. For example, those skilled in the art will understand and appreciate that a methodology can alternatively be represented as a series of interrelated states or events, such as in a state diagram. Moreover, not all acts illustrated in a methodology may be required for a novel implementation.

[0103] This written description uses examples to disclose the invention, including the best mode, and also to enable any person skilled in the art to make and use the invention. The patentable scope of the invention is defined by the claims and may include other examples that occur to those skilled in the art. Such other examples are intended to be within the scope of the claims if they have structural elements that do not differ from the literal language of the claims, or if they include equivalent structural elements with insubstantial differences from the literal languages of the claims.

Claims

CLAIMS1. A method of mixing controlled combustion in a cylinder with a prechamber, the method comprising: actively fueling the prechamber with a first fuel injector; sparking the prechamber with a spark plug operatively connected to the prechamber; injecting fuel into a main chamber from a second fuel injector; producing jet flames that exit at least one orifice of a nozzle of the prechamber and extend into the main chamber; and igniting the fuel injected by the second fuel injector into the main chamber with the jet flames.

2. The method of claim 1, further comprising providing the second fuel injector on a central axis of the main chamber and providing the prechamber at a side of the main chamber.

3. The method of claim 2, wherein the prechamber is at least partially radially external of an internal wall of a crown ring of a piston.

4. The method of claim 1, wherein the first fuel injector and the second fuel injector are connected to a common fuel source.

5. The method of claim 1, wherein a same fuel is injected by the first fuel injector and the second fuel injector, and the fuel has low reactivity and high octane.

6. The method of claim 5, wherein the fuel is selected from E10, E15, E85 and E100 fuel blends.

7. The method of claim 5, wherein the fuel is selected from ethanol, methanol, hydrogen, natural gas, ammonia, gasoline, and E10 and E15.

8. The method of claim 1 , wherein the first fuel injector fuels the prechamber with between 1% and 7% of a total combustion cycle fuel, and the second fuel injector injects a balance 93% - 99% of the total combustion cycle fuel.

9. The method of claim 1, wherein the first fuel injector provides a fuel at an equivalence ratio (c|)) greater than 1.0.

10. The method of claim 9, wherein the fuel is provided by the first fuel injector at an equivalence ratio between 1.1-1.4.

11. The method of claim 1, wherein the first fuel injector injects fuel at a pressure less than 300bar and the second fuel injector injects fuel at a pressure greater than 500bar.

12. The method of claim 1, wherein the prechamber is actively fueled by the first fuel injector by -40° ATDC.

13. The method of claim 1, wherein the spark plug sparks the prechamber between - 25° ATDC and -5° ATDC, and optionally sparks the prechamber between -20° ATDC and -10° ATDC.

14. The method of claim 1, wherein the second fuel injector begins injection of fuel into the main chamber between -15° and -1° ATDC, optionally less than or equal to -10° ATDC, less than or equal to -5° ATDC, or less than or equal to -3° ATDC.

15. The method of claim 1, wherein the spark plug sparks the prechamber at a prechamber pressure below 50 bar.

16. The method of claim 15, wherein injecting fuel into the main chamber from the second fuel injector is a first injection of fuel from the second fuel injector, the first injection of fuel comprising 5-20% of a total combustion cycle fuel, and further comprising:subsequently injecting a second injection of fuel from the second fuel injector into the main chamber, the second injection comprising a balance of the total combustion cycle fuel.

17. The method of claim 16, wherein the first fuel injector fuels the prechamber with between 1 % and 7% of the total combustion cycle fuel and the second injection of fuel comprises 73%-94% of the total combustion cycle fuel.

18. The method of claim 16, wherein the spark plug sparks the prechamber between -40° and -20° ATDC and the first injection of fuel from the second fuel injector occurs within 1 to 2 CAD from the spark plug spark.

19. The method of claim 16, wherein the first injection of fuel from the second fuel injector occurs after the spark plug spark and between -39° and -18° ATDC, and optionally between -35° and -20° ATDC, between -30° and -20° ATDC, between -25° and -20° ATDC, and / or between -25° and -20° ATDC.

20. The method of claim 16, wherein the first injection of fuel from the second fuel injector is ignited by the jet flames, and combustion of the first injection of fuel from the second fuel injector ignites the second injection of fuel from the second fuel injector.

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