FUEL MIXTURES

Priority

[0001] This application claims priority under 35 U.S.C. § 119 to provisional application U.S. Serial Number 62/874,088 filed on July 15, 2019, the entire contents of which are incorporated herein by reference.

Field

[0002] The disclosure pertains to compression ignition and control of same in internal combustion engines.

Background

[0003] Current transportation technologies, including electric vehicles, consume a very large amount of fossil fuels and produce a significant amount of CO2 emissions which contribute to global climate change. For these reasons, there is a great need for a new technology in the transportation market that can significantly reduce energy consumption and pollutant emissions.

[0004] Various known advanced combustion modes, for example, Low Temperature

Combustion (LTC), are capable of high efficiency operation with ultra-low levels of pollutant emissions. However, these have poor control over the combustion process, which result in a relatively narrow operating range. Another mode, called Thermally Stratified Compression Ignition (TSCI), can provide cycle-to-cycle control over advanced combustion such that high efficiency and low emissions operation is achievable over a broad operating range. This known mode of TSCI uses direct water injection using a separate water injector, which separate injector in addition to complicating engine hardware, could not be directly employed in production hardware for modem gasoline or diesel engines without requiring modifications to the engine architecture.

[0005] There is thus a need for an advanced combustion method that provides cycle-to-cycle control over the rate of combustion and over the start of combustion, independently; and that achieves high efficiencies with near-zero levels of soot, NOx, and other harmful pollutant emissions; and that is employable on production hardware without requiring any changes to the architecture of modern production diesel or gasoline engines.

Summary

[0006] In one aspect, the disclosure is directed to a method for controlling compression ignition combustion phasing in an internal combustion engine, the method comprising providing a high heat of vaporization fuel charge, the fuel charge having a latent heat of vaporization of between about 600 to about 1800 kJ/kg; and directly injecting a spray of the high heat of vaporization fuel charge into a cylinder of an internal combustion engine during the intake stroke, the internal combustion engine having a gas exchange stage and a combustion stage, the injecting from a single injector and occurring at least immediately after Top Dead Center (TDC) during the gas exchange stage. In other aspects, a split injection strategy wherein a first portion of the fuel charge, e.g. a water-ethanol mixture, is direct injected during the intake stroke e.g. on a production engine, and a second portion of the fuel charge is direct injected during the compression stroke. In another aspect, injection timing and amount of fuel charge direct injected during the compression stroke is used to control the rate of combustion.

[0007] In another embodiment, the method disclosed enables the injection timing of the intake stroke injection to be used as independent control over the start of combustion. In one aspect, the fraction of fuel that evaporates off of the combustion chamber walls (e.g., the cylinder liner, piston, and head) versus the fraction of fuel that evaporates in the air in the cylinder is thereby controlled. As the directly injected liquid evaporates, it absorbs heat in the phase change to the gaseous state. For every injection event, a certain fraction of the injected liquid evaporates in the air in the cylinder and a certain fraction evaporates off of the cylinder walls and other component parts. When the evaporation occurs in the cylinder air, it absorbs heat from the incoming air, which decreases the air temperature in the cylinder; when the evaporation occurs on the combustion chamber walls, heat is absorbed from the walls, which have a substantially higher specific heat, which evaporation circumstance does not change the temperature significantly. By controlling the injection timing, the fraction that evaporates in the air versus the walls can be varied, and therefore enabling control of the gas temperature in the cylinder on a cycle-to-cycle basis. And by controlling the gas temperature in the cylinder, the start of combustion can be controlled on a cycle-to-cycle basis. This effect can be varied by changing the injector included spray angle for the single direct injection.

[0008] The disclosure further provides a method for achieving high engine efficiencies with near- zero levels of soot, NOx, and other harmful pollutant emissions, which advantages can be implemented on production hardware without requiring any changes to the architecture of modem production diesel or gasoline engines. Finally, when the fuel charge comprises wet ethanol (i.e., a mixture of water and ethanol) as the high heat of vaporization water- fuel mixture, a domestically mass-produced biofuel can be employed by leaving the water content in the fermented ethanol, thereby saving a significant amount of energy otherwise required to produce such fuel.

Brief Description of the Drawings

[0009] Fig. 1 is a schematic of the experimental test cell employed in the Example.

[0010] Fig. 2 is a graph of the cooling potential (units of DK) of ethanol, WE80 (80% ethanol 20% water by mass), methanol, butanol, and gasoline as a function of equivalence ratio assuming all of the heat needed to evaporate the fuel change comes from the air.

[0011] Fig. 3 is a graph of the intake temperature required to maintain a CA50 of 7.0 deg aTDC vs. injection timing for two coolant/oil temperatures (368K/358K and 353K/343K) with two different injector included spray angles (60° and 150°), with error bars showing the expanded uncertainty with a confidence interval of 95%.

[0012] Fig. 4 is a graph of combustion efficiency (pc) vs. injection timing for two coolant/oil temperatures (368K/358K and 353K/343K) with two different injector included spray angles (60° and 150°), with error bars showing the expanded uncertainty with a confidence interval of 95%.

[0013] Fig. 5 is a graph of emission index (El) of unbumed hydrocarbons (uHC) vs. injection timing for two coolant/oil temperatures (368K/358K and 353K/343K) with two different injector included spray angles (60° and 150°), with error bars showing the expanded uncertainty with a confidence interval of 95%.

[0014] Fig. 6 is a graph of emission index (El) of CO vs. injection timing for two coolant/oil temperatures (368K/358K and 353K/343K) with two different injector included spray angles (60° and 150°), with error bars showing the expanded uncertainty with a confidence interval of 95%.

[0015] Fig. 7 is a graph of CA25toCA75 burn duration vs. injection timing for two coolant/oil temperatures (368K/358K and 353K/343K) with two different injector included spray angles (60° and 150°), with error bars showing the expanded uncertainty with a confidence interval of 95%.

[0016] Detailed Description

[0017] The following detailed description of embodiments of the disclosure are made in reference to the accompanying figures. Explanation about related functions or constructions known in the art are omitted for the sake of clearness in understanding the concept of the invention to avoid obscuring the invention with unnecessary detail. Embodiments of the disclosure described herein provide.

[0018] Embodiments of the invention described herein provide a method for control of advanced combustion through a split direct injection of a high heat of vaporization fuel or water-fuel mixtures. In one embodiment, the method provides a new advanced combustion mode of autoignition through compression ignition enabled by water-fuel mixtures. By using water-fuel mixtures, with a specific split injection strategy, this new advanced combustion mode can be realized with cycle-to-cycle control over the start and rate of combustion. The combustion strategy offers significantly lower engine-out emissions and simultaneously higher efficiencies than conventional combustion modes like spark ignition or diesel combustion. The method provides: 1) at least one injection event during the intake stroke where the timing of the injection event is varied to control the start of combustion; 2) at least one injection event during the compression stroke where the timing and mass fraction of fuel are varied to control the rate of combustion; and 3) multiple injection events during either the intake stroke or the compression stroke to improve the combustion efficiency and reduce the unburned hydrocarbon emissions. In one practice, the method disclosed establish that delivering the high heat of vaporization fuel charge in multiple, shorter injections where each injection contains a small amount of fuel can increase the combustion efficiency and reduce the unbumed hydrocarbon emissions, especially for the compression stroke injection process.

[0019] In one embodiment, the disclosure provides a method for controlling compression ignition combustion phasing in an internal combustion engine. In one practice, the internal combustion engine comprises a four-stroke engine as known and commercially available in the art, which comprises, without limitation, a reciprocating piston connected to a crankshaft and configured to travel up and down in a cylinder, the piston having a piston head or crown, with the cylinder comprising internal walls defining the cylinder cavity, an intake valve, an outlet valve, and a liner. Operatively and as used herein, the four strokes include 1) intake (otherwise known as induction or suction), 2) compression, 3) combustion (otherwise known as power or ignition), and 4) exhaust (otherwise known as outlet). Top Dead Center is the position of the piston farthest from the crankshaft and occurs between the exhaust and intake stroke (known herein as the“gas exchange stage”) and at the combustion stroke (known herein as the “combustion stage”).

[0020] A high heat of vaporization fuel charge having a latent heat of vaporization of between about 600 to about 1800 kJ/kg is provided. In various non-limiting practices, the latent heat of vaporization is preferably between about 900 to about 1500 kJ/kg; between about 1000 to about 1200 kJ/kg; and between about 1100 to about 1150 kJ/kg. The fuel charge can be comprised of a single component or a mixture of components, for example, a mixture of one or more alcohols and water. Without limitation serviceable alcohols include those miscible in water and which meet the aforementioned latent heat of vaporization parameters; such alcohols include, without limitation, C _\ -C ₍ , alkanols, including linear and all branched configurations, e.g. methanol, ethanol, propanol, isopropanol, n-butanol, sec -butanol, isobutanol, tert-butanol, and the like. In one practice, the alcohol comprises one or more C ₁ -C ₄ alkanols. Without limitation, when the fuel charge is a mixture of such alcohol and water, the mixture comprises about 5% to about 40% water; and about 95% to about 60% alcohol, by mass. In one embodiment, the fuel charge comprises a mixture of water and ethanol, e.g. about 80% ethanol and about 20% water, by mass, herein referred to as WE80.

[0021] The fuel charge is directly injected into the cylinder (for multi-cylinder engines, each cylinder is operated as herein described) as a spray using a single injector, thus avoiding the need for additional injectors and related hardware and software. Suitable injectors for this purpose including those known in the art that the spray as plumes, the angle between the spray plumes referred to herein as an injector included angle. In non-limiting practices, the injector has an injector included angle of between about 150° to about 30°, including e.g. an injector included angle of between about 118° to about 60°, such as about 90°.

[0022] The high heat of vaporization fuel charge is injected (in whole or in part as hereinafter discussed), into the cylinder during the intake stroke, the injection occurring at least

immediately after Top Dead Center (TDC) during the gas exchange stage. In one non-limiting practices, the direct injection occurs from between about at least immediately after TDC during the gas exchange stage to about -330 degrees after TDC during the combustion stage; in another such practice, direct injection occurs from between about -330 degrees after TDC during the combustion stage to about -240 degrees after TDC during the combustion stage. As understood to the artisan, the phrases using terminology such as -330 degrees after TDC during the combustion stage is also understood to equate with 330 degrees before TDS during the combustion stage. In one embodiment, the step of directly injecting the spray can comprise multiple spray injections.

[0023] In another practice, a split injection protocol is employed as a method for operating a compression ignition internal combustion. A first portion of a high heat of vaporization fuel charge as described herein is directly injected into a cylinder of an internal combustion engine during the intake stroke, the injecting coming from a single injector and occurring at least immediately after Top Dead Center (TDC) during the gas exchange stage; and a second portion of the same high heat of vaporization fuel charge is directly injected into the cylinder using the same injector during the compression stroke. In one embodiment, the first portion comprises between about 70% to about 90%, by mass, of the fuel charge; and the second portion comprises between about 10% to about 30%, by mass, of the fuel charge. In one particular practice, the fuel charge comprises a mixture of water and ethanol, e.g. the fuel charge comprises about 80% ethanol and about 20% water, by mass (WE80), and the first portion comprises about 70% of the fuel charge, by mass; and the second portion comprises about 30% of the fuel charge, by mass.

[0024] The step of directly injecting the first portion and/or the second portion can comprises multiple spray injections during the intake stroke.

[0025] In yet another aspect, the disclosure provides a method of compression ignition in an internal combustion engine which comprises comprising a step of directly injecting a spray of the high heat of vaporization fuel charge described herein from a single injector into a cylinder of an internal combustion engine during the intake stroke, wherein the injecting step comprises multiple injections and a fraction of each injection evaporates in the air of the cylinder; that is, a fraction of each injection evaporates not from the available surfaces of the cylinder and the component pieces and parts therein, but instead evaporates within the air in the volume of the cylinder defined by the internal cylinder walls, whereas the remainder of that fraction evaporates from the available surfaces, such as the piston crown and other combustion chamber surfaces. In one embodiment, the fraction that evaporates in the air of the cylinder is between about 20% to about 95%; in another embodiment, the fraction that evaporates in the air of the cylinder is between about 30% to about 90%; in yet another embodiment, the fraction that evaporates in the air of the cylinder is between about 40% to about 85%; in another embodiment, the fraction that evaporates in the air of the cylinder is between about 15% to about 95%; in still another embodiment, the fraction that evaporates in the air of the cylinder is between about 5% to about 99%. In one aspect, the injection timing during the intake stroke determines what fraction evaporates in air versus what fraction evaporates on the on cylinder walls. In one practice in this regard, the direct injection from the single injector occur between about -350 degrees after Top Dead Center (TDC) during combustion stage to about -180 degrees after TDC during the combustion stage. In yet another practice, the direct injections from the single injector occur about -330 degrees after Top Dead Center (TDC) during combustion stage to about -240 degrees after TDC during the combustion stage.

[0026] The disclosure further relates to a method of controlling the rate of combustion in an advanced compression ignition combustion process of an internal combustion engine comprising providing a high heat of vaporization fuel charge as herein described; and directly injecting a spray of the high heat of vaporization fuel charge into a cylinder of an internal combustion engine during the compression stroke, the injecting from a single injector and occurring between about -140 degrees after Top Dead Center (TDC) during combustion stage to about -30 degrees after TDC during the combustion stage. In another practice, the injecting from a single injector and occurring between about -100 degrees after Top Dead Center (TDC) during combustion stage to about -50 degrees after TDC during the combustion stage.

[0027] In one aspect, the methods disclosed herein apply to after engine start.

[0028] The following example is illustrative of the disclosure and not limiting to same. [0029] Example:

[0030] Experimental Set-up

[0031] Experiments were conducted fully instrumented single cylinder research engine. The engine consisted of a 4-cylinder production, diesel 1.7L General Motors/Isuzu engine head, with one cylinder mounted to a Ricardo Hydra engine block and the other three cylinders deactivated. A custom, wide, shallow-bowl piston is used. This contrasts the standard, diesel- style re-entrant bowl used in production, which is not ideally suited for a homogeneous, low temperature combustion mode due to its’ large squish region and high surface area, which both contribute to higher heat transfer losses unnecessarily. The production camshafts are used unmodified and provide 12 degrees of positive valve overlap. The engine geometry and valve timings are shown in Table 1.

[0032] A diesel common rail was used and pressurized by a Bosch CP3 pump. Two production Bosch solenoid-style diesel injectors were used in this work: DLLA150P (150° included angle) and DLLA60P (60° included angle). Custom intake and exhaust manifolds were used, each connected to a custom-made plenum. Infineum R655 lubricity additive was added to the wet ethanol (<500ppm) to lubricate the injection system since the injection system was normally lubricated with the diesel fuel being injected. The wet ethanol was tested in HCCI combustion with and without the lubricity additive and it was found that the lubricity additive did not have an effect on combustion (i.e., there is no effect of the lubricity additive on the reactivity of the fuel mixture).

[0033] Table 1. Engine geometry and valve timings.

Compression | -| 16.0 Engine Speed TRPM1 1200

EVC [deg aTDC] 366

[0034] A custom-made LabView program was built to serve as both a data acquisition system and an engine control unit. As a data acquisition system, all high- and low- speed measurements are taken, and real-time combustion analysis is computed. As an engine control unit, the LabView code provides direct control over the injection pressure and timing, among other things. In the following experiments, the fuel injection pressure was maintained at 500 bar. The injection timing was varied while the injection duration was adjusted at each injection timing to maintain a constant fuel flowrate.

[0035] Crank angle measurements, in increments of 0.1°, were provided by a Kistler encoder coupled to a pulse multiplier. Four high- speed Kistler pressure transducers (in-cylinder, intake, exhaust, and fuel rail) were read at each crank angle increment. Fuel flow was measured with a Coriolis flow meter. A Horiba MEXA-7100DEGR motor exhaust gas analyzer provided C02, 02, CO, NOX, and unbumed hydrocarbon (uHC) measurements from the exhaust gas. All engine tests were conducted on an active GE dynamometer at 1200 RPM. For each case studied, 300 consecutive cycles were saved and ensemble-averaged. A custom MATLAB code then processed the ensemble- average data, and outputted the results from combustion analysis. The heat release analysis and uncertainty analysis were performed by the custom MATLAB code follow the guide outlined in Gainey, B., Longtin, J. P., & Lawler, B. (2019). A Guide to Uncertainty Quantification for Experimental Engine Research and Heat Release Analysis. SAE International Journal of Engines, 12{ 5), 509+. The experimental test cell is shown in Fig. 1. All instrument ranges and uncertainties are shown in Table 2.

[0036] Table 2. Model, range, and instrument uncertainty of instruments used.

Model Used For Range Uncertainty

[0037] Results:

[0038] Cooling Potential

[0039] The latent heat of vaporization of neat ethanol and water is 0.85 J/mg and 2.26 J/mg, respectively. With a blend of 80% ethanol and 20% water by mass (WE80), the effective latent heat of vaporization of wet ethanol is 1.13 J/mg, slightly higher than methanol (1.1 J/mg), twice as large as butanol (0.58 J/mg), and greater than 350% higher than gasoline (0.31 J/mg for gasoline). Fig. 2 shows the cooling potential of ethanol, WE80, methanol, butanol, and gasoline as a function of equivalence ratio. Here, cooling potential is defined as the change in air temperature resulting from complete evaporation of the fuel in the air. The amount of fuel injected into the cylinder will affect how much cooling potential is available. Therefore, in Fig. 2, the cooling potential is both a function of the latent heat of vaporization of the fuel and the stoichiometric air fuel ratio (AFR). Note, the stoichiometric AFR of neat ethanol is 9.0026 and the stoichiometric AFR of WE80 is 7.202. Table 3 tabulates the fuel properties for the fuels used in Fig. 2. Fig. 2 shows that WE80 and methanol have the highest cooling potentials. The cooling potential is calculated by the following equation:

F * AFR stoich * h fg, fuel Cooling Potential =

C p, air

[0040] Where hfg, fuel is the latent heat of vaporization of fuel, F is the equivalence ratio, AFRstoich is the stoichiometric air-fuel ratio, and Cp, air is the specific heat of air.

WE80 is available as a domestically produced biofuel and leaving the water content in the ethanol allows some energy and cost savings during production. WE80 also has a slightly higher lower heating value than methanol (21.4 MJ/kg vs. 19.9 MJ/kg). In the following experiments, WE80 was used with an equivalence ratio of 0.4. This means there is 63K of cooling potential available with WE80, 750% greater than gasoline’s cooling potential of 8.4K. However, not all of the available cooling potential can be used to cool the air. Wall wetting will result in heat from the liner and/or piston crown being used to evaporate the fuel. Controlling the fraction of available cooling potential that is used can therefore control the intake temperature on a cycle-to- cycle basis. With 63K of cooling potential, there is the potential for cycle-to-cycle control of a large range of intake temperature.

[0041] Table 3: Fuel properties, including molecular formula, stoichiometric air-fuel ratio (AFRstoich), and latent heat of vaporization (hfg) under STP conditions

[0042] Injection timing sweep

[0043] To control the fraction of fuel that evaporates in the air vs. on the wall, the amount of fuel that wets the walls must be controlled. With such a high heat of vaporization, the spray penetration length during the intake stroke (low in-cylinder temperature/pressure) will be very large. The location of the piston during injection can therefore control the length a spray can travel before hitting the walls. With this in mind, an injection timing sweep was performed with two different injectors and two different wall temperatures. The effect that injector included angle and wall temperature have on the range of intake temperature control was examined as discussed below. The two injector included spray angles used in this study were 60° and 150°. Although the wall temperature was not measured directly, changing the coolant and oil temperature were used to change the wall temperature. The two coolant/oil temperatures used in this study are 368K/358K (95°C/85°C) and 353K/343K (80°C/70°C).

[0044] To determine the effective range of intake temperature control, the injection timing sweeps were performed with a constant fueling rate of WE80 (18.75 ± 0.50 mg/cycle) and a constant CA50 (7.0 ± 0.5 deg aTDC). The upstream intake temperature was adjusted at each injection timing to control combustion phasing at the desired CA50. Notably, although the intake temperature varied largely, the amount of air flowing into the cylinder each cycle did not vary appreciably. This was because the charge is cooled by fuel evaporation (i.e., the fuel’s cooling potential), resulted in the same temperature at IVC and therefore the same combustion phasing. The air flowrate did not decrease due to a decrease in air density as intake temperature increased because IVC temperature actually remained constant, and therefore, the equivalence ratio also remained constant (0.4 ± 0.02).

[0045] The two injector included spray angles used in this study, 150° and 60°, are vastly different, targeting different parts of the combustion chamber. The 150° injector is typically used in CDC, where fuel is injected near TDC into a deep-bowled piston, creating a diffusion flame. When this injector is used to deliver fuel during the intake stroke and into a cylinder whose piston geometry is a wide, shallow bowl, the spray plumes tend to impinge on the liner and the available spray penetration length is always approximately half of the cylinder bore. Only near TDC do the edges of the piston bowl intercept the spray. This means that once the edge of the bowl moves down past the spray plumes, the ability to extend the spray length by injecting later is lost. This contrasts the 60° injector, whose available spray penetration length increases as the piston recedes to BDC until about -240 deg aTDC, where the spray will begin to impinge on the liner. The maximum available spray penetration length for the 60° injector occurs at -240 deg aTDC and is 2 times the cylinder bore. As a result, the injection timing sweep for the 150° injector was only carried out to -300 deg aTDC. Another reason for this w a s that the combustion efficiency became excessively low (<80%) and

combustion became unstable after -300 deg aTDC due to excessive wall impingement and poor evaporation.

[0046] Intake Temperature

[0047] Fig. 3 shows the intake temperature required to achieve a CA50 of 7.0 ± 0.5 deg aTDC vs. injection timing for both coolant/oil temperatures with both spray angles. Due to the limitation on spray penetration length imposed by the 150° injector, the range of control over the intake temperature was limited, as seen in Fig. 3. On the other hand, the 60° injector’s range of control over the intake temperature was very large. The range of intake temperature control of the 150° injector was 9.7K and 8.3K for the high and low coolant/oil temperatures, respectively.

The range of intake temperature control of the 60° injector was 57.5K and 47.3K for the high and low coolant/oil temperatures, respectively. The results show that with the 60° injector, cycle- to-cycle control of combustion phasing for advanced combustion strategies using wet ethanol as a fuel can be achieved.

[0048] It was expected that the required intake temperature would increase with retarding injection timings. This trend was seen for the entire limited injection timing sweep with the 150° and up until an injection timing of -240 deg aTDC with the 60° injector. The required intake temperature of -210 deg aTDC was actually lower than that of -240 deg aTDC. The required intake temperature of -180 deg aTDC w as then higher than both -210 deg aTDC and -240 deg aTDC. [0049] Combustion Efficiency and Emissions

[0050] Fig.4 shows the combustion efficiency (r|c) vs. injection timing for both coolant/oil temperatures with both injector included spray angles. The combustion efficiency of both injectors increased when the coolant/oil temperature was increased. However, the increase in combustion efficiency due to the increase in coolant/oil temperature was not the same for both injectors. At a coolant/oil temperature of 353K/343K, the 150° injector had a low combustion efficiency at each injection timing. As previously mentioned, the combustion efficiency was extremely low at -270 deg aTDC, combustion became unstable, and the injection timing sweep was truncated. Contrarily, the 60° injector tended to wet the piston crown, which lacked direct cooling and was therefore less sensitive than the liner to changes in coolant/oil temperature. As a result, the combustion efficiency of the 60° injector was higher than that of the 150° injector when the coolant/oil temperature was 353K/343K for similar injection timings. The combustion efficiency of the 60° injector increased less than the 150° injector when the coolant/oil temperature was raised to 368K/358K, resulting in similar combustion efficiencies among the two injectors for similar injection timings.

[0051] The emissions index (El) of unbumed hydrocarbons (uHC), and El CO vs. injection timing for both coolant/oil temperatures with both spray angles are shown in Fig. 5 and Fig. 6, respectively. The engine-out uHC emissions in these experiments followed the reverse of the trend outlined by the combustion efficiency. On the other hand, the CO emissions, specifically for both injectors with the coolant/oil temperatures of 368K/358K, did not directly follow the reverse of the trend outlined by the combustion efficiency. For both injectors, the CO emissions were lower with the higher coolant/oil temperatures. For the 150° injector, the CO emissions remained nearly constant even though the combustion efficiency is decreasing. For the 60° injector, the CO emissions decreased with retarding injection timing. While not being bound to any theory, it is believed that as the CO emissions are related to the fraction of fuel that evaporates off the walls, since that will result in colder walls with rich regions nearby, that might inhibit complete combustion. [0052] Burn Duration

[0053] Fig. 7 shows the CA25 to CA75 bum duration vs. injection timing for both coolant/oil temperatures with both spray angles. The difference in the amount of thermal stratification resulting from a change in wall temperature of ~15K is negligible. Therefore, no significant change in bum duration caused by a change in thermal stratification was expected due to a change in coolant/oil temperature. For a given injection timing with the 60° injector, the bum duration was not highly sensitive to the coolant/oil temperature, with the difference in burn duration for the two coolant/oil temperatures falling within the bounds of uncertainty. However, with the 150° injector, the burn duration was longer with the lower coolant/oil temperature for a given injection timing. The resulting low combustion efficiencies are coupled with the longer bum duration.

[0054] The length of the error bars in Fig. 7 provide guidance into the cyclic variability (cycle- to-cycle combustion variation) of the operating condition. The variability in the bum duration of the 150° injector was larger with the lower coolant/oil temperatures due to worsening evaporation conditions and poor combustion efficiency. On the other hand, the variability in the bum duration of the 60° injector w a s not significantly affected by the coolant/oil temperature. However, the cyclic variability in the bum duration did increase as the injection timing is retarded. Furthermore, it can be seen from Fig . 7 that the bum duration begins to trend upward as injection timing is retarded. This was seen at injection timings as early as -140 deg aTDC [21] and would not be surprising if it occurred as early as -240 deg aTDC.

[0055] Conclusions

[0056] Fuels with a high latent heat of vaporization have a large charge cooling potential. By adjusting the intake stroke injection timing of a direct injected fuel with a high latent heat of vaporization, the fraction of fuel that evaporates in the air vs. the fraction of fuel that wets and evaporates off of the walls are controlable, providing a means to control the temperature at intake valve closing, and thus the combustion phasing, on a cycle-to-cycle basis, which is something advanced combustion concepts have a great need for.

[0057] In this study, a mixture of 80% ethanol and 20% water by mass was used as the high latent heat of vaporization fuel. The experiments were conducted on a single cylinder, Ricardo Hydra engine block that mimics the geometry of a 1.7 L GM/Isuzu diesel engine. Intake stroke injection timing sweeps of direct-injected HCCI combustion were performed, adjusting the intake air temperature to match combustion phasing at 7.0 deg aTDC. These injection timing sweeps were performed with two different coolant/oil temperatures

(368K/358K and 353K/343K) with two different injector spray angles (60° and 150°).

[0058] The results show that through a range of injection timings from -330 deg aTDC to -180 deg aTDC, the fraction of fuel evaporating in the air ranges from 0.22 to 0.82. Advanced combustion concepts, like HCCI, TSCI and RCCI, have a significant need for high-speed control of combustion phasing. The large range of intake temperature control found in this study provides a means to achieve this high controllability of the combustion phasing of an advanced, low temperature combustion concept, on a cycle-to-cycle basis.

[0059] Advantages of the method disclosed herein include: similar efficiency levels to slightly more efficient than diesel engines (used for heavy duty application, on-highway trucking, and construction); significantly cleaner emissions characteristics than diesel engines; significantly cheaper than diesel engines; significantly better control and a larger operating range than other advanced combustion strategies; can use current production engine hardware without requiring any change to the engine architecture; can use a domestically mass-produced biofuel and saves energy during the production of the biofuel

[0060] Abbreviations

deg aTDC degress after Top Dead Center CAA- Crank angle location when x% of the fuel has burned (e.g. CA50)

CAD Crank angle degrees

CDC Conventional diesel combustion

CFD Computational fluid dynamics

CO Carbon monoxide

C0 ₂ Carbon dioxide

CoV Coefficient of variance

DI Direct injection

DWI Direct water injection

EGR Exhaust gas recirculation

El Emissions index

EVC Exhaust valve closing

EVO Exhaust valve opening

GHRR Gross heat release rate

HCCI Homogeneous charge compression ignition

IMEP _g Gross indicated mean effective pressure

IVC Intake valve closing

IVO Intake valve opening LTC Low temperature combustion

MPRR Maximum pressure rise rate

NO _x Oxides of Nitrogen (NO or NO ₂ )

NVO Negative valve overlap

Gross thermal efficiency llcomh Combustion efficiency

0 ₂ Diatomic oxygen

PCCI (diesel) Premixed

charge compression ignition

PFI Port fuel injection

PFS Partial fuel stratification

PM Particulate matter

PLIF Planar laser-induced fluorescence

P-V Pressure-volume diagram diagram

RCCI Reactivity controlled compression ignition

SACI spark assisted compression

SF Split fraction

SI Spark ignition

SOC Start of combustion

SOI Start of Injection TSA Thermal stratification

analysis

TSCI Thermally stratified compression ignition

uHC Unburned hydrocarbons

WE Wet ethanol

WI W ater inj ection

VCR Variable compression

ratio

Ratio of specific heats

[0061] While the invention has been shown and described with reference to certain embodiments of the present invention thereof, it will be understood by those skilled in the art that various changes in from and details may be made therein without departing from the spirit and scope of the present invention and equivalents thereof.