Prechamber design for cylinder periphery mounted igniter in fuel agnostic prechamber enabled mixing-controlled combustion

Abstract

A mixing controlled combustion (MCC) engine includes a cylinder having a cylinder head and cylinder walls. A piston is configured for reciprocal movement within the cylinder. The piston includes a crown defining a bowl and an annular wall between the bowl and an exterior perimeter of the piston. The crown and the cylinder define a main chamber. A direct injector extends through the cylinder head on an axis common to the cylinder and the piston. The direct injector configured to provide a fuel to the main chamber. A prechamber system includes a prechamber body defining a prechamber and a nozzle extending through the cylinder head. The nozzle includes at least one orifice oriented radially towards the direct injector. The prechamber system includes a spark plug. The prechamber system includes a fuel injector configured to provide the fuel to the prechamber.

Description

PRECHAMBER DESIGN FOR CYLINDER PERIPHERY MOUNTED IGNITER IN 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 fuel reactivity or high octane number, for example blends of gasoline and ethanol, from pure gasoline to pure ethanol. 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 higher reactivity or higher cetane number fuels, such as 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 dioxide (CO2), nitrogen oxides (NOx), and soot.

[0004] A natural solution to combat high emissions is to replace diesel fuel with cleaner burning low carbon intensity fuels in the MCC process, thus the engine can maintain its performance characteristics enabled by MCC. However, clean burning low carbon intensity fuels,such as ethanol, methanol, natural gas, propane, hydrogen, and ammonia - all have very low cetane number and arc not conducive to the MCC process. In order to bum low-cctanc 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 the fuel 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 - where the more reactive diesel fuel serves as the ignition source. In premixed systems, a diesel pilot is used to initiate combustion of the premixed low cetane fuel. However, these systems exhibited a lack of combustion stability over the operational range, increased HC emissions at middle to low loads, and still have challenges with knock and pre-ignition like an SI engine. 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, propane, ammonia, and gasoline / ethanol fuel blends, such as E10, E15, and E85.

[0008] An example of a mixing controlled combustion (MCC) engine includes a cylinder having a cylinder head and cylinder walls. A piston is configured for reciprocal movement within the cylinder. The piston includes a crown defining a bowl and an annular wall between the bowl and an exterior perimeter of the piston. The crown and the cylinder define a main chamber. Adirect injector extends through the cylinder head on an axis common to the cylinder and the piston. The direct injector configured to provide a fuel to the main chamber. A prechamber system includes a prechamber body defining a prechamber and a nozzle extending through the cylinder head. The nozzle includes at least one orifice oriented radially towards the direct injector. The prechamber system includes a spark plug. The prechamber system includes a fuel injector configured to provide the fuel to the prechamber.

[0009] In further examples of the MCC engine, a fuel source is common to the direct injector and to the fuel injector and the fuel is low cetane and high octane. The prechamber has a volume between 2-5% of a clearance volume of the main chamber. The fuel injector is configured to provide 2-5% of a total cycle fuel to the prechamber, and the direct injector is configured to provide the balance 95-98% of the total cycle fuel to the main chamber. The at least one orifice is angled radially towards the axis at an included angle (P) between 12.5°-22.5° below horizontal. The at least one orifice includes four or fewer orifices. The at least one orifice includes 1, 2, or 3 orifices. A swept angle (a) between the orifices between 0-40° and optionally between 10-30°. The swept angle (a) is centered at the direct injector. The at least one orifice is a single orifice, and has an extent angle of 30° or less, and optionally of 20° or less. The single orifice has a cross- sectional area between 9-12.5 mm2. A total cross-sectional area of the at least one orifice is between 4.5-12.5 mm2, optionally between 6.8-12.5 mm2, and optionally between 9-12.5 mm2. The nozzle the prechamber is in alignment within the cylinder with at least a portion of the wall of the piston. The wall of the piston further includes a cove configured to accommodate the nozzle when the piston is at top dead center (TDC) within the cylinder. The cove extends into the wall from the bowl and a wall face, the cove being radially interior of exterior perimeter of the piston.

[0010] In further examples, of the MCC engine, the prechamber system is a first prechamber system and the MCC engine includes a second prechamber system positioned across the cylinder from the first prechamber system. The second prechamber system includes a prechamber body defining a prechamber, a nozzle extending through the cylinder head, the nozzle includes at least one orifice oriented radially towards the direct injector, the second prechamber includes a fuel injector and a spark plug.

[0011] In still further examples of the MCC engine, the direct injector includes a plurality of orifices through which the direct injector provides high pressure plumes of the fuel. The orifices of the direct injector are arranged in an arc asymmetrically biased towards the prechamber. Thefuel injector is configured to inject 2-5% of a total cycle fuel to the prechamber. The spark plug is configured to spark at a timing between -30° and -10° after top dead center (ATDC). The direct injector is configured to begin injection of 95-98% of the total cycle fuel to the main chamber after the spark plug spark. The prechamber is configured to produce jet flames out of the at least one orifice into the main chamber from combustion of the fuel in the prechamber. The prechamber is fueled to an equivalence ratio of 1.1 < (|> < 1.4.BRIEF DESCRIPTION OF THE DRAWINGS

[0012] Figure 1 depicts an example of a mixing controlled combustion cylinder with a side mounted prechamber as presently disclosed.

[0013] Figure 2 is a perspective view of a prechamber body.

[0014] Figure 3 is a cross-sectional view of a prechamber system.

[0015] Figure 4 is an image of a modeled prechamber jet flames in a main chamber.

[0016] Figure 5 is graph that depicts a basic fueling and spark strategy for ignition assistedMCC.

[0017] 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.

[0018] Figure 6B is a graph of simulated results for various prechamber equivalency ratios.

[0019] Figure 7 A is a top perspective view of an exemplary prechamber and main chamber.

[0020] Figure 7B is a side- sectional view of the exemplary depicts prechamber and main chamber of Fig. 7A.

[0021] Figure 8 presents graphs summarizing ignition assistance across various orifice swept angles.

[0022] Figure 9 presents graphs summarizing ignition assistance performance for various total orifice cross-sectional areas.

[0023] Figure 10 presents graphs summarizing ignition assistance performance based for various prechamber volumes.

[0024] Figure 13 depicts a further example of a prechamber with a single orifice.

[0025] Figure 11 is an image of a modeled prechamber combustion.

[0026] Figure 12 depicts an example of a modified direct injector with asymmetric fuel plumes.

[0027] Figure 13 depicts an example a mixing controlled combustion cylinder with two prechambers.DETAILED DISCLOSURE

[0028] 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.

[0029] Disclosed herein is an active prechamber ignition assistance for a mixing-controlled combustion (PC-MCC) engine and process which enables the MCC of low cetane / 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.

[0030] 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.

[0031] 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 abovethe depth of the bowl 1 12. 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.

[0032] Table 1 provides exemplary dimensions and characteristics for configurations of the prechamber.

[0033] 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 two orifices 128. The orifice(s) 128 are centered along a line 9 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 9 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 9 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 (c.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 theprominence 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.

[0034] 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.

[0035] 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.

[0036] 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:

[0037] 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.

[0038] 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 a common 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 adirect injector 122 is a 538-7297: C9.3B HRC-A common rail fuel injector, available from Caterpillar, Inc.

[0039] 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.

[0040] 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.

[0041] 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.

[0042] 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 bum duration is observed as the prechamber traverses from lean to fuel-rich operation.

[0043] Close inspection of the main chamber HRR for a rich prechamber fueling 0 = 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 occurs for 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 = 0.8), stoichiometric (e.g. 0 = 1.0), and rich (e.g. 0 = 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.

[0044] 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.

[0045] 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 bum 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 on equivalence 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.

[0046] 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 arc 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.

[0047] 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 assistance standpoint, 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.

[0048] 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 jetsare 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.

[0049] 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.

[0050] 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 igniter configurations 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.

[0051] 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 higherconcentrations 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.

[0052] 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 direct injector. 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.

[0053] 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 CA50 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.

[0054] 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. c|) ~ 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.

[0055] 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.

[0056] The prechamber volume impacts prechamber pressure during compression and combustion. As the volume increases, the observed pressure difference between main chamber and prechamber 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 are 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.

[0057] 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 ofthe larger combustion volume. The additional prechamber fuel energy in the larger volume enables higher internal jet core velocities that arc 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.

[0058] 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 CASO 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.

[0059] 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 fuel injector 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 opposingsides towards the high pressure plumes of direct injected fuel. These examples may promote ignition assistance and improve combustion at advanced spark timings.

[0060] 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).

[0061] 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.

[0062] 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, diesel-like performance is achieved with the use of low sooting fuels and can be operated globally stoichiomctrically, enabling the use of the low cost 3-way catalysis after treatments commonly available to SI engines while exhibiting diesel-like performance.

[0063] 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.

[0064] 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.

[0065] 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 mixing controlled combustion (MCC) engine comprising: a cylinder having a cylinder head and cylinder walls; a piston configured for reciprocal movement within the cylinder, the piston comprising a crown defining a bowl and an annular wall between the bowl and an exterior perimeter of the piston, wherein the crown and the cylinder define a main chamber; a direct injector extending through the cylinder head on an axis common to the cylinder and the piston, the direct injector configured to provide a fuel to the main chamber; and a prechamber system comprising a prechamber body defining a prechamber, a nozzle extending through the cylinder head, the nozzle comprising at least one orifice oriented radially towards the direct injector, the prechamber system comprising a spark plug and comprising a fuel injector configured to provide the fuel to the prechamber.

2. The MCC engine of claim 1, further comprising a fuel source common to the direct injector and to the fuel injector and the fuel is low cetane and high octane.

3. The MCC engine of claim 1, wherein the prechamber has a volume between 2-5% of a clearance volume of the main chamber.

4. The MCC engine of claim 1, wherein the fuel injector is configured to provide 2- 5% of a total cycle fuel to the prechamber, and the direct injector is configured to provide a balance 95-98% of the total cycle fuel to the main chamber.

5. The MCC engine of claim 1, wherein the at least one orifice is angled radially towards the axis at an included angle ((3) between 12.5°-22.5° below horizontal.

6. The MCC engine of claim 1, wherein the at least one orifice comprises 1, 2, or 3 orifices.

7. The MCC engine of claim 1 , comprising a swept angle (a) between at least one center of the at least one orifice between 0-40° and optionally between 0-30° or between 10-30° or between 20-30°.

8. The MCC engine of claim 7, wherein the swept angle (a) is centered at the direct injector.

9. The MCC engine of claim 1, wherein the at least one orifice is a single orifice, and has an extent angle of 30° or less, and optionally of 20° or less.

10. The MCC engine of claim 9, wherein the single orifice has a cross-sectional area between 9-12.5 mm2.

11. The MCC engine of claim 1, wherein a total cross-sectional area of the at least one orifice is between 4.5-12.5 mm2, optionally between 6.8-12.5 mm2, and optionally between 9-12.5 mm2.

12. The MCC engine of claim 1, wherein the nozzle of the prechamber is in alignment within the cylinder with at least a portion of the annular wall of the piston.

13. The MCC engine of claim 12, wherein the annular wall of the piston further comprises a cove configured to accommodate the nozzle when the piston is at top dead center (TDC) within the cylinder, the cove extending into the wall from the bowl and a wall face, the cove being radially interior of exterior perimeter of the piston.

14. The MCC engine of claim 1, wherein the at least one orifice comprises three or fewer orifices.

15. The MCC engine of claim 1, wherein the prechamber system is a first prechamber system, and further comprising:a second prechamber system positioned across the cylinder from the first prechamber system, and comprising a prechamber body defining a prechamber, a nozzle extending through the cylinder head, the nozzle comprising at least one orifice oriented radially towards the direct injector, the second prechamber system comprising a fuel injector and a spark plug.

16. The MCC engine of claim 1, wherein the direct injector comprises a plurality of orifices through which the direct injector provides high pressure plumes of the fuel.

17. The MCC engine of claim 16, wherein the orifices of the direct injector are arranged in an arc asymmetrically biased towards the prechamber.

18. The MCC engine of claim 1, wherein the fuel injector is configured to inject 1-7% of a total cycle fuel to the prechamber, and the spark plug is configured to spark at a timing between -35° and +5° after top dead center (ATDC).

19. The MCC engine of claim 16, wherein the direct injector is configured to begin injection of 93-97% of the total cycle fuel to the main chamber after the spark plug spark.

20. The MCC engine of claim 1, wherein the prechamber is configured to produce jet flames out of the at least one orifice into the main chamber from combustion of the fuel in the prechamber.

21. The MCC engine of claim 1, wherein the prechamber is fueled to an equivalence ratio of 1.1 < <|) < 1.4.

22. The MCC engine of claim 1, wherein the fuel injector is configured to inject the fuel at a pressure less than 300 bar- and the direct injector is configured to inject the fuel at a pressure greater than 500 bar.- l-

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