Development and Validation of a Reduced Chemical Kinetic

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Development and Validation of a Reduced Chemical Kinetic Mechanism for CFD Simulations of Natural Gas/Diesel Dual Fuel Engines Andrew Hockett, Greg Hampson, and Anthony J Marchese Energy Fuels, Just Accepted Manuscript • DOI: 10.1021/acs.energyfuels.5b02655 • Publication Date (Web): 26 Jan 2016 Downloaded from http://pubs.acs.org on February 3, 2016

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Development and Validation of a Reduced Chemical Kinetic Mechanism for CFD Simulations of Natural Gas/Diesel Dual Fuel Engines Andrew Hockett1, Greg Hampson2, and Anthony J. Marchese1,* 1

Department of Mechanical Engineering, Colorado State University, Fort Collins, Colorado, USA 2

Woodward Inc., Loveland, Colorado, USA

KEYWORDS: dual fuel, reduced mechanism, chemical kinetic mechanism, natural gas, CFD, diesel engine

ABSTRACT: A reduced chemical kinetic mechanism consisting of 141 species and 709 reactions has been constructed for simulating the combustion of both natural gas and diesel fuels in a dual fuel engine. Natural gas chemical kinetics are modeled as a mixture of methane, ethane, and propane, while the diesel fuel chemical kinetics are modeled as n-heptane. The new reduced mechanism combines reduced versions of a detailed nheptane mechanism and a detailed methane through n-pentane mechanism, each of which was reduced using a direct relation graph method. The reduced dual fuel mechanism is validated against ignition delay computations with full detailed mechanisms, adiabatic HCCI simulations with full detailed mechanisms, experimental premixed laminar flame speeds of CH4/O2/He mixtures at 40 and 60 atm, ignition delay and lift-off length from a diesel spray experiment in a constant volume chamber, and finally against dual fuel engine experiments using multi-dimensional CFD simulations. The engine simulations were performed for direct comparison against natural gas/diesel dual fuel engine experiments at varying injection timings, engine loads, substitution 1

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percentages, and natural gas compositions. An engine experiment with additional propane was used to induce engine knock for the purpose of validating the reduced mechanism’s ability to predict natural gas auto-ignition. The results show that the newly reduced mechanism accurately reproduces the chemical kinetic behavior of the detailed mechanism, including laminar flame speed at high pressure, ignition delay and lift off length in the diesel spray experiment, and pressure and heat release rate in the engine experiments. In the experiment where engine knock occurred, the model predicts the phasing and magnitude of a sudden acceleration in combustion rate and reproduces the observed high frequency pressure oscillations.

1. INTRODUCTION The incentive to power internal combustion engines with natural gas continues to increase due to both economic and emissions advantages of natural gas in comparison to traditional gasoline and diesel fuels. Natural gas is expected to remain less expensive than petroleum fuels on an energy equivalent basis through 20401. For many engine applications the fuel cost savings is sufficient to offset the initial capital investment of converting equipment to operate on natural gas. In addition, lean premixed combustion of natural gas avoids the rich regions within diesel sprays where soot is formed. Furthermore, the U.S. Environmental Protection Agency (EPA) has initiated regulations on greenhouse gas (GHG) emissions and fuel efficiency standards for light duty vehicles by 2018 and heavy duty vehicles by 20252,3. Because natural gas is composed primarily of methane, natural gas can offer an approximate 25% reduction in carbon dioxide emissions compared to diesel fuel due to the higher hydrogen to carbon ratio and higher heating value4. However, the reduction in total GHG emissions can be compromised if the natural gas does not burn to completion and is exhausted as unburned hydrocarbon (UHC) emissions, because methane has a global warming potential that is 25 times greater than carbon dioxide over a 100 year period5. Therefore, if UHC emissions can be minimized, then natural gas engines can offer a possible path towards meeting new GHG emissions regulations with significant fuel cost savings. Diesel engines can be readily converted to natural gas/diesel dual fuel engines by fumigating the intake manifold with natural gas and using a small diesel injection as an ignition source. Ignition of the diesel fuel 2


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subsequently initiates combustion of the premixed natural gas air mixture near the diesel flame. If the natural gas to air ratio is within the flammability limits, then a propagating premixed flame will consume the remaining natural gas in the cylinder. The amount of natural gas that is substituted for diesel is known as the substitution percentage, z, and is defined on an energy basis using the lower heating values of the fuels. Depending on the speed and load, the substitution percentage can be limited by a variety of constraints including excessive pressure rise rate, engine knock, pre-ignition, injector tip temperature, nitric oxide (NOX) emissions, and UHC emissions due to incomplete combustion of natural gas6. A greater understanding of the in-cylinder physics can lead to novel engine technologies that can expand the natural gas substitution limits and minimize incomplete combustion of natural gas. To this end, multi-dimensional computational modeling is often employed to gain valuable information about in-cylinder physics and aid in the development of new engine designs and combustion strategies. Prior to the present study, a sufficiently accurate and compact chemical kinetic mechanism was not available to enable accurate simulation of natural gas/diesel dual fuel combustion. In particular, no reduced dual fuel mechanism was available that accounted for the variation in chemical reactivity of natural gas due to variation in chemical composition. Niemann et al.7 used pure methane as a natural gas surrogate in a multi-dimensional model for natural gas/diesel reactivity controlled compression ignition (RCCI) under the assumption that a natural gas mixture with high methane content should behave similarly to pure methane. However, Bourque et al.8 found that ignition delay periods and flame speeds for typical pipeline natural gas compositions with small concentrations of heavier hydrocarbons differ substantially in comparison with pure methane. Rahimi et al.9 presented a 76 species reduced chemical kinetic mechanism for natural gas/diesel homogeneous charge compression ignition (HCCI) engines. This mechanism was developed by combining a reduced n-heptane mechanism with the GRI-Mech 3.0 natural gas mechanism10 and then using a genetic optimization algorithm to adjust rate constants for three different engine operating conditions consisting of different natural gas 3


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substitution percentage, total equivalence ratio, exhaust gas recirculation (EGR), and compression ratio. However, only one natural gas composition was validated in the study by Rahimi et al. and therefore a rate constant optimization and validation study would likely be necessary before modeling a different natural gas composition. Furthermore, the mechanism of Rahimi et al. was validated for HCCI simulations using a five zone model and not for sufficiently resolved multi-dimensional CFD simulations of premixed propagating natural gas flames or non-premixed diesel spray flames, which are present in conventional dual fuel engines. The objective of the present study was to develop a reduced chemical kinetic mechanism from highly detailed base mechanisms for use in multi-dimensional CFD simulations of natural gas/diesel dual fuel engines where premixed propagating natural gas flames and non-premixed diesel spray flames are present. The reduced dual fuel mechanism is validated against ignition delay computations with full detailed mechanisms, adiabatic zerodimensional HCCI simulations with full detailed mechanisms, experimental premixed laminar flame speeds of CH4/O2/He mixtures at 40 and 60 atm, ignition delay and lift-off length from a diesel spray experiment in a constant volume chamber, and finally against dual fuel engine experiments using multi-dimensional CFD simulations performed using the commercial CONVERGE™ code. The newly developed mechanism did not require any rate constant tuning and successfully reproduces in-cylinder pressure measurements from dual fuel engine experiments conducted under multiple operating conditions including varying substitution percentage, injection timing, engine load, two different natural gas compositions, and a case with additional propane to induce engine knock. Moreover, the mechanism is sufficiently compact such that three-dimensional, Reynolds averaged Navier-Stokes (RANS) based engine simulations can be conducted within a reasonable computational time (less than 24 hours).

2. CHEMICAL MECHANISM REDUCTION METHOD The new reduced dual fuel chemical mechanism was constructed from previously published detailed mechanisms. Diesel fuel was modeled using n-heptane due to similar Cetane number and heating value with 4


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diesel fuel11. Natural gas was modeled using mixtures of methane, ethane, and propane as a surrogate. While pipeline quality natural gas is predominantly methane, its reactivity is greatly influenced by the presence of ethane and propane, which are typically present in low concentrations (< 10% by volume combined), but have much higher reactivity than methane. Larger hydrocarbons such as butane and pentane were not accounted for in the reduced mechanism, because the assumption was made that their concentration in the pipeline natural gas mixture being modeled would be low enough not to significantly impact combustion behavior12,13. A reduced methane/ethane/propane mechanism was formulated from the detailed methane through n-pentane mechanism (293 species, 1588 reactions) from Healy et al.14 at the National University of Ireland Galway, hereafter referred to in this paper as the NUIG mechanism. This detailed natural gas mechanism has been previously validated against experimental ignition delay data from rapid compression machine and shock tube experiments at temperatures between 630 and 1550 K, pressures up to 30 bar, and at lean, stoichiometric, and rich conditions. A reduced n-heptane mechanism was formulated from the detailed n-heptane mechanism from Curran et al.15,16 at Lawrence Livermore National Laboratories (LLNL) with 654 species and 2827 reactions, hereafter referred to as the LLNL mechanism. The detailed LLNL mechanism has previously been validated against experimental ignition delay data from rapid compression machine and shock tube experiments for temperatures between 550 and 1700 K, pressures between 1 to 42 atm, and equivalence ratios from 0.3 to 1.5. Both the NUIG and LLNL detailed mechanisms were developed by many of the same authors and therefore employ similar values for most of the reaction rate constants. Also the LLNL n-heptane mechanism includes many of the natural gas reaction pathways present in the NUIG mechanism because higher hydrocarbons break down into smaller hydrocarbon species that are important in natural gas chemistry. However, not all of the reactions and rate constants for smaller hydrocarbon species are identical between the NUIG natural gas mechanism and the LLNL n-heptane mechanism. Without a rigorous validation of the LLNL n-heptane mechanism against experimental data for natural gas mixtures, the choice was made to reduce the two detailed mechanisms separately and then combine the reduced mechanisms. 5


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The detailed base mechanisms were reduced using the direct relation graph with error propagation and sensitivity analysis (DRGEPSA) method17, which is an automatic mechanism reduction algorithm included within the CONVERGE™ software package. The NUIG mechanism was reduced based on ignition delay calculations of two different natural gas mixtures in air as described in Table 1. The target species included helium and argon to retain the third body efficiency information because these inert species are often used in laminar flame experiments at high pressure. The 85% methane composition was chosen to be similar to the composition measured in the engine experiments. This reduction resulted in a 49 species skeletal mechanism for methane, ethane, and propane mixtures. A preliminary study of this reduced mechanism found that the ignition delay for pure propane in air did not agree well with the detailed NUIG mechanism. Therefore, another reduction was performed for a composition with higher propane than ethane content, which resulted in a 46 species skeletal natural gas mechanism. Upon comparing the 46 and 49 species mechanisms, there were 4 species that were not in the initial 49 species mechanism: C3H6OOH1-3, C3H6OOH1-3O2, IC3H7O2, and C3KET13. These additional species are important in the low temperature branching pathways for propane. Consequently, the 46 species skeletal mechanism yielded much better agreement with the detailed mechanism for ignition delay of pure propane. Ignition delay plots of pure methane, ethane, and propane in air are given in the supplementary information. Next, the DRGEPSA method was used to reduce the LLNL n-heptane mechanism to a 131 species and 656 reactions skeletal mechanism. The three reduced mechanisms were then combined to form the final 141 species and 709 reactions reduced chemical kinetic mechanism for mixtures of methane, ethane, propane, and n-heptane, which is hereafter referred to as the CSU141 mechanism. An illustration of the reduction procedure is given in Figure 1. Any duplicate reactions between the three reduced mechanisms were given the rate constant values from the reduced natural gas mechanisms to ensure that engine knocking from natural gas auto-ignition is accurately represented. Care was taken to find and add reactions from the detailed base mechanisms that were not in any of the individual reduced mechanisms, but whose species were all present after combining the reduced mechanisms. For laminar flame speed calculations, 6


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the natural gas transport data from Bourque et al.8 was combined with the transport data from the LLNL nheptane mechanism, whereby any duplicate species were given the transport data from the natural gas data file. The CSU141 mechanism files (species, reactions, thermodynamic data, and transport data) are included in the supporting information. Table 1: Conditions used for reducing the base mechanisms using DRGEPSA. Detailed Base Mechanism Target species Fuel composition (mole frac.) Temperature (K) Pressure (bar) Equivalence ratio, φ Ignition delay error (%) Sensitivity analysis fraction

LLNL n-heptane15,16

NUIG Natural Gas14

NUIG Natural Gas14

n-C7H16, O2, OH, HO2, CH4, C2H6, C3H8, He, Ar, N2 n-C7H16 = 1.0

CH4, C2H6, C3H8, O2, OH, HO2, CO2, H2O, Ar, He, N2

CH4, C2H6, C3H8, O2, OH, HO2, CO2, H2O, Ar, He, N2

CH4 = 0.85, C2H6 = 0.1, C3H8 = 0.05 750, 800, 850, 900, 950, 1000, 1050, 1100, 1150, 1200 10, 30, 80 0.2, 0.4, 0.6, 0.8, 1.0, 1.2, 1.4 0.5 0.75

CH4 = 0.93, C2H6 = 0.02, C3H8 = 0.05 750, 800, 850, 900, 950, 1000, 1050, 1100, 1150, 1200 10, 30, 80 0.2, 0.4, 0.6, 0.8, 1.0, 1.2, 1.4 0.5 0.75

700, 750, 800, 850, 900, 950, 1000, 1050, 1100, 1150, 1200 10, 30, 80 0.2, 0.5, 1.0, 1.5, 2.0 0.1 0.75

Figure 1: Procedure used for reducing detailed mechanisms and combining reduced mechanisms. Ignition delay and zero dimensional HCCI engine calculations were performed using CHEMKIN®. The ignition delay period was defined by the maximum temperature derivative corresponding to the largest 7


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temperature increase. In this paper, the CSU141 mechanism is not compared against experimental ignition delay data, because a heat loss model would need to be added in the calculations and this would introduce additional error making it difficult to assess the accuracy of the reduction method. The reader should refer to the detailed mechanisms for comparisons with experimental ignition delay periods and then consider the percent error between the detailed mechanism and the reduced mechanism.

3. ENGINE EXPERIMENTAL PROCEDURE Natural gas/diesel dual fuel engine experiments were used to validate the CSU141 mechanism. Engine experiments were conducted using a four cylinder light-duty General Motors (GM) 1.9 L, common rail, turbocharged diesel engine. The specifications of the engine are listed in Table 2 and the diesel injector specifications are listed in Table 3. Engine experiments were conducted with two different natural gas composition as listed in Table 4, which are referred to as the low reactivity natural gas and the high reactivity natural gas. The high reactivity natural gas mixture had noticeably higher levels of ethane and propane. It should be noted that both of the natural gas compositions were municipal natural gas, which underscores the wide range of natural gas reactivity of “pipeline quality” natural gas. Table 2: Engine specifications. Cylinder arrangement Displacement volume Bore Stroke Connecting rod length Compression ratio Injection system Valves per cylinder Rated power Rated torque

4 Inline 1.9 L 82 mm 90.4 mm 161 mm 17.5 Bosch Common Rail 4 DOHC 110 kW (147.5 hp) @ 4000 RPM 315 Nm (232 ft-lbf) @ 2000 RPM

Table 3: Diesel injector specifications.

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Name Nozzles Nozzle orifice diameter Included spray angle Discharge coefficient

Bosch CRIP 2-MI 7 141 micron 148 0.86

Table 4: Measured natural gas composition at two different facilities. Species CH4 C2H6 C3H8 CO2 N2

Low Reactivity Natural Gas 0.9461 0.0285 0.0031 0.0125 0.0093

High Reactivity Natural Gas 0.8812 0.0823 0.0075 0.0167 0.0123

For the experiments described herein, the engine was modified to allow multi-port injection of natural gas. This modification was accomplished by building a natural gas accumulator with attached Woodward natural gas injectors and connecting tubing from each injector to the intake manifold runners, such that the gas was injected perpendicular to the air flow direction. In a similar fashion a propane injector was connected to cylinder 2 using a ‘T’ connector in the tube line for natural gas. The propane injection system was used to provide a means of varying the natural gas reactivity to examine the effect of natural gas reactivity on engine knock. The engine and injectors were controlled by a Woodward Engine Control Module (ECM), which enabled individual cylinder control of diesel and natural gas injection timing and duration. The diesel fuel consumption was measured using Micro Motion CMF025 Coriolis mass flow meters on the supply and return of the fuel tank, which have an accuracy of ±0.1% of the measured rate. Natural gas consumption was measured using a Micro Motion CMF010 Coriolis mass flow meter, which has an accuracy of ±0.35% of the measured rate. The air mass flow rate was measured with an automotive hot wire anemometer, which has an accuracy of ±3%. Incylinder pressure transducers (Kistler 6058A) were installed in the glow plug hole of each cylinder and pressure data were acquired on a National Instruments data acquisition system. Crank angle data were acquired with a resolution of 0.25 crank angle degrees. Apparent heat release rates (AHRR) were calculated based on 100 cycle 9


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averaged pressure data and assuming a constant ratio of specific heats of 1.3. A 5-point moving average was applied to the AHRR profiles. In order to accurately model the breakup and vaporization of diesel sprays, it is important to obtain a profile of the injected mass flow rate throughout the injection duration, which is specific to the injector being modeled. As described in Hockett et al.18, a rate of injection meter was used to measure the rate shape, the injection duration, and the delay period between applied electric current and start of injection. This rate of injection data was gathered for each of the four injectors used in the engine experiments. These rate of injection experiments also provided injected mass as functions of electronic pulse width and injection pressure, which was used to set the injected mass in the engine simulations. The experimental operating conditions are listed in Table 5. The nomenclature for referring to the different operating conditions in Table 5 uses a ‘D’ to designate diesel followed by the mass of diesel injected per cylinder per cycle in units of milligrams rounded to the nearest integer. In a similar manner, the letter ‘G’ signifies the mass of natural gas in milligrams and ‘P’ signifies the mass of propane in milligrams. Start of injection (SOI) timing sweeps were performed for two different fuel combinations at 2200 RPM and the highest boost pressure attainable at this engine speed. The D25G17 SOI sweep had a moderate substitution rate of 43% natural gas by energy and the D12G22 sweep had a higher substitution rate of 68%. Additionally, a load sweep at 2000 RPM and a lower boost pressure was performed by increasing the mass of natural gas while keeping the injected diesel mass constant. The load sweep used the higher reactivity natural gas composition while the SOI sweeps used the lower reactivity natural gas composition as described in Table 4. Additionally, in the D12G22P2 experiment, an additional 2 mg of propane was added to a case from the D12G22 SOI sweep to induce engine knocking and assess whether the model could predict natural gas auto-ignition. Therefore, these experiments encompass a wide range of substitution percentages, engine loads, natural gas equivalence ratios, natural gas reactivity, and TDC conditions for validation of the reduced mechanism.

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Table 5: Experimental operating conditions. Fuel case name Engine Speed (RPM) Injected diesel per cylinder (mg) Natural gas per cylinder (mg) Propane per cylinder (mg) Natural gas substitution, z (%) Natural Gas Composition Diesel injection pressure (bar) Injection Duration (CAD) Electronic SOI (°ATDC) Boost pressure (kPa abs) Manifold temperature (°C) Net IMEP (bar) Natural gas equivalence ratio Total equivalence ratio

D25G17 2200 25.4 16.5 0 43.3 Lower Reactivity 700 13.4 -8.8 -13 169.3 167.2 43.6 43.5 15.0 15.6 0.357 0.362 0.845 0.856

D12G22 2200 12.1 ± 3 21.8 0 67.9 Lower Reactivity 700 6.2 -8 -14 168.2 163.6 44.6 44.4 10.9 10.7 0.479 0.512 0.712 0.765

D11G8 2000 10.6 8.0 0 47.6 Higher Reactivity 700 5.9 -12 120 39.6 7.0 0.25 0.55

D11G18 2000 10.6 17.9 0 67.1 Higher Reactivity 700 5.9 -12 120 38.6 11.7 0.55 0.74

D12G22P2 2200 11.7 ± 3 22.1 2.3 71.9 Lower Reactivity 700 5.3 -10 163 48.2 11.7 0.600 0.824

4. COMPUTATIONAL METHOD FOR MULTI-DIMENSIONAL ENGINE SIMULATIONS Computational modeling of the in-cylinder physics was performed using the CONVERGE™ code19. All combustion simulations presented herein began at intake valve closure (IVC) and used a 1/7th sector of the cylinder geometry, wherein each sector included one of the injector’s seven nozzles as depicted in Figure 2(a). To initialize the velocity and turbulence fields at IVC, a preliminary simulation of the intake stroke was performed utilizing the geometry of the full cylinder with valves and intake and exhaust ports as depicted in Figure 2(b). The results of this intake stroke simulation was then used to map the velocity and turbulence fields in the sector simulations. A single cylinder GT-POWER model was used to find the trapped mass at IVC, the exhaust residual mass, the IVC pressure, and the IVC bulk temperature. The sector simulations assumed homogenous composition, temperature, and pressure of the gas at IVC. While keeping the trapped mass fixed, the IVC pressure and temperature used in the sector simulations were slightly adjusted (±5 kPa) from the values 11


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provided by GT-POWER in order to obtain closer agreement with the experimental pressure during compression. In some cases, small adjustments to piston and head wall temperatures (±20 K) were necessary to match TDC pressure.

Figure 2: (a) 1/7th sector geometry used for engine combustion simulations. (b) Cylinder and port geometry from the GM 1.9L engine used to obtain the velocity and turbulence fields at IVC for use in the sector simulations. The cylinder liner is hidden to show the piston geometry.

CONVERGE™ uses a fixed and uniform Cartesian grid and a cut cell method at the boundaries. The simulations used adaptive mesh refinement (AMR) based on temperature, velocity, and n-heptane mass fraction with the minimum cell size being 0.25 mm as recommended by Senecal et al.11. Fixed embedding of 0.25 mm cells was used in the near nozzle region. The base cell size had a cell width of 4 mm. The transport equations were solved using a cell centered finite volume method with 2nd order central differencing for the surface fluxes. The gas phase transport properties were based on the viscosity and thermal conductivity for air. The combustion chemistry was modeled using the CSU141 mechanism presented herein. To reduce computational cost, a multizone chemistry model was employed where the chemistry for cells similar in temperature, progress equivalence ratio, and mass fraction of n-heptane are solved simultaneously. The n-heptane criterion is necessary in dual fuel simulations to prevent cells with similar equivalence ratio, but different fuel types from being grouped

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together. When simulating a case where engine knock occurred, the hydroperoxyl radical (HO2) was added as a dimension in the multi-zone chemistry strategy. The assumption of a well stirred reactor in each cell was used. No sub-grid model for turbulence-chemistry interactions (TCI) was used to correct for commutation error between average cell temperature and average cell reaction rate. However, Pomraning et al.20 found that the well-stirred reactor assumption with a Reynolds averaged Navier-Stokes (RANS) turbulence model and grid refinement using AMR can successfully achieve grid convergence for premixed and non-premixed combustion at a cell width of 0.25 mm. They found that the converged solution agreed well with experimental measurements of premixed and non-premixed flames. Pomraning et al. concluded that in a sufficiently resolved RANS simulation of typical flow conditions in an engine (i.e. Reynolds number) the magnitude of the TCI commutation error from the well-stirred reactor assumption must not be significant in comparison to other sources of uncertainty if the goal is to capture global pressure and heat release rates. The RANS turbulence model used was the renormalization group (RNG) κ-ε model21. The turbulent heat transfer model from Amsden22 was used for the wall boundaries. The spray model utilized in this study was that of the Lagrangian drop-Eulerian fluid type. The liquid phase was modeled as a single component with the thermodynamic properties of diesel fuel obtained from the CONVERGE™ liquids property library (referred to as ‘DIESEL2’)19, which are the recommended properties from Convergent Science, Inc. The breakup of the spray parcels was modeled using the modified KH-RT droplet breakup mechanism, which does not use a defined breakup length as described in the CONVERGE™ theory manual19. A more comprehensive description of the spray model is provided in the supporting information for this work. The spray model was validated against liquid and vapor penetration measurements from experiments done by the Engine Combustion Network (ECN)23, which consisted of a vaporizing, but non-reacting spray injected into a constant volume chamber at typical TDC temperature and pressure conditions. Specific details about the ECN

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datasets used and comparisons of simulated and experimental liquid and vapor penetration can be found in the supporting information for this work.

5. RESULTS AND DISCUSSION 5.1. n-Heptane Ignition Delay Calculations Figure 3 shows comparisons of adiabatic zero dimensional ignition delay periods between the detailed LLNL n-heptane mechanism and the CSU141 reduced dual fuel mechanism. Comparisons are made for n-heptane/air mixtures at pressures of 10, 30, and 80 bar, temperatures between 700 to 1300 K, and equivalence ratios between very lean, φ = 0.2, and rich, φ = 2.0. These temperatures and pressures are relevant to diesel sprays at engine TDC conditions. The reduced CSU141 mechanism has excellent agreement with the detailed base mechanism across all conditions presented in Figure 3 and in particular captures the negative temperature coefficient (NTC) region resulting from the two stage ignition process for n-heptane. Figure 3(d) shows excellent agreement at rich mixtures between 700 and 1000 K, which are conditions relevant to diesel spray ignition. The maximum deviation for the φ = 2.0, 80 bar results occurs at 1100 K where the reduced mechanism predicts a 0.02 ms longer ignition delay period, which translates to a 0.25 crank angle degree at 2000 RPM. It should be noted that the ignition delay of diesel jets in dual fuel engines is not controlled solely by chemical kinetics and that turbulent mixing rates are also important. In addition to conventional dual fuel combustion, early injection schemes such as RCCI require accurate modeling of the chemical kinetics of diesel fuel over a wider range of equivalence ratios. Fuel air mixtures in RCCI engines can be anywhere between slightly rich to very lean and begin at temperatures lower than peak compression temperatures. Figures 3(a), 3(b), and 3(c) show excellent agreement for CSU141 at lower temperatures for very lean to stoichiometric conditions.

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Figure 3. Ignition delay plots for n-heptane/air mixtures comparing the CSU141 reduced mechanism with the detailed LLNL15,16 mechanism for (a) very lean φ = 0.2, (b) lean φ = 0.5, (c) stoichiometric φ = 1.0, and (d) rich φ = 2.0 mixtures.

5.2. Natural Gas Ignition Delay Calculations Comparisons of ignition delay between the detailed NUIG mechanism and the reduced CSU141 mechanism for two different natural gas/air mixtures consisting of methane, ethane, propane, and air are shown in Figure 4. The 90/6.7/3.3 mixture consisted of 90% methane, 6.7% ethane, and 3.3% propane by mole. The 70/20/10 mixture consisted of 70% methane, 20% ethane, and 10% propane by mole. The ignition delay plots show that the reduced CSU141 mechanism captures with excellent accuracy the behavior of the detailed mechanism for both mixtures across the range of pressures, temperatures, and equivalence ratios. As discussed earlier, experimental data are not included in these plots because of the need to model the heat transfer in ignition delay 15


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experiments from a rapid compression machine (RCM). The error in ignition delay between the reduced and the detailed natural gas mechanisms is less than 0.1% at all conditions presented and therefore it should be expected that the reduced mechanism can perform as well as the detailed mechanism when compared with experimental data. While the NUIG mechanism was only validated with experimental data up to 30 atm24, the 80 bar calculations were of interest because typical turbocharged light-duty diesel engines can have TDC compression pressures near 80 bar. Therefore, assuming the NUIG mechanism is accurate at diesel TDC pressures, the CSU141 mechanism should be able to predict auto-ignition in natural gas/diesel dual fuel engine simulations.

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Figure 4: Comparisons of natural gas/air ignition delay calculations between the reduced CSU141 mechanism and the detailed NUIG14 mechanism. The 90/6.7/3.3 mixture is shown in the left column for (a) very lean φ = 0.2, (b) lean φ = 0.5, and (c) stoichiometric φ = 1.0 mixtures. The right column shows the 70/20/10 mixture at (d) very lean φ = 0.2, (e) lean φ = 0.5, and (f) stoichiometric φ = 1.0 mixtures.

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The ignition delay calculations presented thus far were conducted at constant volume and fixed initial temperature/pressure. The performance of the mechanism was further validated for the effects of varying temperature and pressure from the compression and expansion strokes in an engine by performing zerodimensional adiabatic HCCI engine simulations using CHEMKIN®. The engine geometry for these calculations was identical to that of the experimental engine described previously. The engine speed was chosen to be 2000 RPM, which is close to the speed used in the engine experiments. Figure 5 shows computational comparisons between the detailed LLNL n-heptane mechanism and the CSU141 mechanism for n-heptane/air mixtures at φ = 0.5 and φ = 2.0. For these computations, the initial pressure was fixed at 1.7 bar and the initial temperature was varied. The CSU141 and the LLNL mechanisms exhibited excellent agreement in terms of start of combustion timing for the low temperature heat release, high temperature ignition, and pressure rise. The agreement tends to diverge slightly as the start of combustion retards past TDC. However, the results of Figure 5 strongly suggest that the CSU141 mechanism can predict start of combustion behavior of diesel sprays and auto-ignition in HCCI or RCCI engines to the same level of accuracy as the detailed LLNL n-heptane mechanism.

Figure 5: Adiabatic zero-dimensional HCCI engine simulation comparisons between the detailed LLNL15,16 and reduced CSU141 mechanism for varying initial temperatures of n-heptane/air mixtures at (a) φ = 0.5 and (b) φ = 2.0. 18


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Figure 6(a) shows calculated HCCI pressure curves for the 90/6.7/3.3 natural gas mixture in air at φ = 0.5 for both the detailed NUIG mechanism and the reduced CSU141 mechanism for various initial temperatures at a fixed initial pressure of 1.7 bar. Figure 6(b) shows a similar comparison for the 70/20/10 natural gas mixture. As shown, the reduced mechanism captures the ignition behavior of the detailed mechanism for these natural gas mixtures. At lower initial temperature when the start of combustion begins to retard past TDC, the reduced mechanism begins to predict earlier start of combustion than the detailed mechanism. However, the results of Figure 6 strongly suggest that the CSU141 will predict the end gas auto-ignition of natural gas to the same level of accuracy as the detailed NUIG mechanism.

Figure 6: Zero-dimensional adiabatic HCCI engine simulation comparisons between the detailed NUIG14 and reduced CSU141 mechanism for varying initial temperatures at φ = 0.5 for (a) the 90/6.7/3.3 natural gas mixture in air and (b) the 70/20/10 natural gas mixture in air.

5.4. Laminar Flame Speed Calculations of Methane at High Pressures High pressure methane laminar flame speed measurements using spherically propagating flames in a constant volume chamber were conducted by Rozenchan et al.25 at 40 and 60 atm. These high pressure experiments were 19


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conducted with helium as the inert and with reduced oxygen concentration to prevent flow instabilities and maintain laminar behavior. Specifically, laminar flame speed measurements in CH4/O2/He mixtures at various equivalence ratios were conducted at 40 atm using an oxidizing gas consisting of 17 % O2 and 83% He by volume and at 60 atm using an oxidizing gas with 15% O2 and 85% He by volume. For these experiments, the unburned gas temperature was held fixed at 298 K. Figure 7 compares the experimental and calculated results using the reduced mechanism at both 40 and 60 atm. In the 40 atm case, the CSU141 mechanism agrees with experiments to within 6% at lean to stoichiometric conditions and slightly over predicts the flame speed at rich conditions. For the 60 atm case, the reduced CSU141 mechanism over predicts the flame speed across the range of equivalence ratios investigated, but the deviation diminishes as the mixture becomes leaner. The relative error in flame speed at the leanest condition, φ = 0.85, is only 8% and at φ = 0.9 is 16%. Martinez et al.26 compared a variety of natural gas mechanisms to experimental laminar flame speeds for methane/air mixtures at 20 atm and different equivalence ratios. At φ = 0.9 they found a wide spread in predictions with most mechanisms over predicting the flame speed by 20% to 100%. Because the agreement between the simulations and experiments for methane laminar flame speed at high pressure had less than 20% error, CSU141 is as effective as any of the natural gas mechanisms evaluated by Martinez et al. In addition, the highly turbulent flame speeds experienced in a diesel dual fuel engine can often be more dependent on the turbulence intensity than the laminar flame speed, which allows for greater error in laminar flame speed predictions to be tolerated. Because the natural gas/air mixtures in dual fuel engines are usually lean and contain mostly methane, the results in Figure 7 are encouraging and suggest that the reduced CSU141 mechanism can be used to predict flame speed for lean natural gas mixtures at high pressures in dual fuel engines.

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Figure 7: Comparisons of experimental laminar flame speeds with CSU141 computations for CH4/O2/He mixtures at (a) 40 atm and (b) 60 atm. The experimental data is from Rozenchan et al.25. The indicated O2 % refers to the molar composition in the O2/He oxidizer gas.

5.5. Combusting Diesel Spray The ECN spray database23 includes a set of data for a diesel spray at an injection pressure of 1400 bar into a chamber filled with 21% oxygen at a nominal temperature of 1000 K and density of 30 kg/m3, which are typical TDC conditions for high compression ratio diesel engines. This data set was used to validate the combined capabilities of the spray, turbulence, and reduced chemical kinetic mechanism to simulate reacting diesel sprays. Table 6 lists some of the important experimental conditions for this spray, while more information can be found on the ECN website23. Experimental ignition delay and flame lift-off length are used as benchmarks for validating the simulations. The experimental ignition delay is based on the time between start of injection and the time of rapid pressure rise. In the simulations, the ignition delay is the elapsed time when the temperature derivative reaches a local maximum that corresponds with the largest temperature rise. The experimental lift-off length is the axial distance from the nozzle to the flame as measured by OH chemiluminescence. The simulated lift-off length is indicated by the location where the OH mass fraction becomes larger than 2% of the maximum in the domain. The simulated lift-off length was measured at 2 ms, which is well after steady state behavior is established. For these simulations, the liquid properties of diesel 21


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were represented using the properties of n-tetradecane, because it was found that the use of n-tetradecane had a 1 mm improvement in lift-off length compared to using “DIESEL2” from the CONVERGE liquids property library. Table 6: Experimental conditions for the ECN combusting diesel spray used for model validation23. Fuel Injection pressure (bar) Injected mass (mg) Injection duration (ms) Nominal ambient temperature (K) Nominal ambient density (kg/m3) Oxygen % volume Nozzle orifice diameter (mm) Discharge coefficient Fuel temperature (K)

Diesel #2 1400 14.3 5.1 1000 30 21 0.100 0.8 436

Figure 8 shows the simulation results using a plane through the center of the spray colored by the simulated mass fraction of OH and also indicates the lift-off length. Table 7 presents a comparison of the simulated and experimental ignition delay and lift-off length. The simulated ignition delay is within 14% error of the experimental. When the ignition delay periods are converted to crank angle degrees at an engine speed of 2000 rpm, the simulated ignition delay is longer than the experimental by only 0.4 crank angle degrees, which is acceptable for engine simulations. The simulated lift-off length is within 20% error of the experimental with a difference of 2 mm. The error in ignition delay and lift-off length obtained in this study is comparable to the level of error achieved by other authors. For instance, Bajaj et al.27 compared simulations and optical spray experiments of reacting n-heptane sprays across a wide range of ambient conditions and generally achieved liftoff length agreement within 25% and ignition delay agreement within 30%. The error in ignition delay and liftoff length are the result of an accumulation of multiple error sources including the accuracy of the detailed mechanism, turbulence-chemistry interactions omitted by assuming a well-stirred reactor in each cell, solving only for the ensemble averaged flow via the RANS turbulence model, and the method used to define the experimental values27. The agreement between model and experiment demonstrates that the coupling of the 22


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reduced chemistry, turbulence, and spray models can provide reasonable predictions of non-premixed diesel jet flames that are useful for engine modeling.

Figure 8: Simulation result for OH mass fraction at 2 ms. The lift-off length is defined at the first location of 2% of the maximum OH mass fraction.

Table 7: Comparison of experimental and simulated ignition delay and lift-off length. Ignition delay (ms) 0.24 0.273

Experiment Simulation

Lift-off length (mm) 9.5 11.4

5.6. Multi-dimensional Engine Simulations The experimental and computational pressure and heat release rates for the D25G17 and D12G22 SOI timing sweeps are compared in Figure 9(a) and 9(b) respectively. Each simulation took approximately 12 hours to compute using 48 cores. As shown in Figure 9, excellent agreement between model and experiment was obtained for the start of combustion phasing, pressure rise rate, peak pressure, combustion duration, and shape of heat release rate as injection timing changed. A quantitative summary of the error from the engine simulations is given at the end of this section. There is some over prediction of the AHRR during the premixed ignition of the spray, as indicated by the peak in AHRR at start of combustion. For dual fuel combustion, the first peak in the AHRR curve includes heat release from the premixed burn of the diesel vapors and entrained natural gas in the head of the spray. The slight discrepancy in AHRR between model and experiment during 23


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spray ignition could be the result of using a single component fuel surrogate for the diesel chemistry, a single component liquid surrogate for the vaporization model, error in measurement method for the injector rate shape, experimental uncertainty in the mass of diesel fuel injected, and using a sector simulation, which assumes perfect symmetrical behavior for each spray. Overall, the model has effectively captured the general trends in pressure and AHRR behavior for the two injection timings. The model predictions are particularly noteworthy given the fact that no adjustments (i.e. tuning) of model constants was done between different injection timing cases. The only differences made between the input files other than injection timing included accounting for small variations in manifold pressure and the ingested air mass as these parameters varied slightly in the experiments as injection timing was advanced.

Figure 9: Comparisons of experimental and computational pressure and AHRR for (a) the D25G17 SOI sweep and (b) the D12G22 SOI sweep.

In-cylinder temperature plots in the plane of the spray axis from the D25G17 simulation with -13° SOI are shown in Figure 10. From Figure 10(a) it can be seen that the ignition of the spray in dual fuel combustion is similar to traditional diesel spray ignition, where a premixed auto-ignition reaction zone is observed near the head of the spray plume. The premixed auto-ignition zone grows around the spray and transitions into a non24


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premixed flame by approximately -1° ATDC as shown in Figure 10(b). The non-premixed spray flame then initiates a premixed propagating natural gas flame as seen by the growth of the green area, which indicates a flame temperature of approximately 2000 K as shown in Figures 10(c) through 10(f). Notice that the nonpremixed diesel flame continues to burn during natural gas flame propagation and therefore a dual fuel model must be able to simultaneously simulate the important physics in both non-premixed and premixed flames. The images in Figure 10 demonstrates that the essential physical phenomena of the diesel spray ignition, the nonpremixed diesel jet flame, and the premixed propagating natural gas flame necessary for reproducing global AHRR has been effectively captured via the coupling of the reduced chemical kinetic mechanism, the RANS turbulence model, and the spray models described herein.

Figure 10: Time development of temperature in the plane of the spray axis for the D25G17, -13° ATDC SOI case. Time proceeds sequentially from (a) to (f).

Figure 11 presents comparisons between the experiments and simulations for the natural gas load sweep using a fixed boost pressure and fixed diesel injection mass and timing. This load sweep therefore provides validation with variation in natural gas equivalence ratio. As previously explained, these experiments were conducted with a higher reactivity natural gas mixture due to greater ethane and propane content. Additionally, the boost 25


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pressure used for the load sweep was appreciably lower than that used in the SOI sweeps. The simulations were able to capture the start of combustion timing, the pressure rise rate, the shape and magnitude of the AHRR profile, and the combustion duration. Again, some over prediction was observed for the heat release rates during spray ignition, which can be attributed to the same sources of error discussed above for the SOI timing sweep. The peak pressure was somewhat under predicted and as a result the expansion pressure was also lower. Examination of the experimental AHRR for the D11G18 case near the timing of peak pressure shows a small increase in the AHRR that is not captured by the simulations. This discrepancy could be due to sources of error that are not specifically due to the accuracy of the chemistry in the reduced chemical mechanism, such as uncertainty in air and fuel mass flow measurements, the assumptions of a RANS based turbulence model, assuming constant wall temperatures, assuming homogenous composition and temperature at IVC, and the symmetry assumption from using a sector simulation. In general, the agreement in Figure 11 demonstrates the ability of the reduced mechanism to predict the combustion behavior responsible for the observed AHRR curves at two widely different natural gas equivalence ratios, natural gas substitution percentages, and total fuel mass. In particular, the CSU141 mechanism is able to capture the observed AHRR curves from a higher reactivity natural gas mixture than the previous mixture used in the SOI sweeps.

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Figure 11: Comparisons of experimental and computational pressure and AHRR for the dual fuel load sweep using a constant diesel injection of 11 mg.

Comparisons between 100 cycle averaged experimental engine data and volume averaged simulation results for the case with additional propane, D12G22P2, are shown in Figure 12(a). The experimental AHRR curve shows a sudden acceleration in combustion rate occurring just before peak pressure, which indicates end gas auto-ignition of natural gas. The simulation also captures the timing of this spike in AHRR, but the magnitude is somewhat higher leading to higher peak pressure and consequently higher pressure during expansion. Figure 12(b) compares experimental pressure from a single cycle and simulated pressure from a single cell at the same location as the pressure transducer tip. The single cycle experimental pressure shows fluctuations that can be classified as engine knock. The single cell simulation pressure shows pressure fluctuations with similar frequency and amplitude beginning when the fast combustion event occurs.

Figure 12: Comparisons between experiment and simulation for the additional propane case where engine knocking was observed. (a) 100 cycle averaged experimental pressure and volume averaged simulation

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pressure. (b) Single cycle experimental pressure and simulation pressure at a single cell with the same location as the pressure transducer tip.

The agreement in the AHRR profile for the propane addition experiment shows that the CSU141 mechanism can be useful in explaining the chemical kinetic processes that cause natural gas auto-ignition and knocking in dual fuel engines. Figure 13(a) through (d) are three dimensional images of the engine sector showing the development of a 2000 K iso-surface, which is indicative of the boundary between unburned and burned regions. At 1º ATDC in Figure 13(a) the combustion zone has enveloped the spray axis and the non-premixed diesel spray flame has been established. A portion of the temperature iso-surface has crossed the periodic boundary and therefore represents the combustion zone from the adjacent diesel spray moving into the sector. The swirl motion causes the premixed flame to pass through only one of the periodic boundaries. By 3° ATDC in Figure 13(b), the burned region has grown in a direction normal to the spray axis, which is representative of a propagating natural gas flame. Also, the burned region has moved downward along the bowl wall due to the impact of the diesel jet. The growing iso-surface in the squish region is due to natural gas flame propagation. Figure 13(b) shows that the premixed flame from the adjacent spray has merged with the flame from this sector, creating a cusp like crease where the two flames meet. In the 6.5° ATDC image of Figure 13(c), a planar cross section plot of local heat release rate has been added showing the location of the reaction zones in the bottom of the bowl in relation to the temperature iso-surface. For clarity of the local heat release planar cross section, the 2000 K iso-surface has been removed in Figure 13(d), which depicts three flames. The flame along the bowl wall is the non-premixed diesel flame spreading down into the bottom of the bowl and the other two are due to the merging premixed natural gas flames from adjacent spray axes. All three are expanding in size as they travel into the bottom of the bowl.

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Figure 13: Development of a 2000 K temperature iso-surface with a local heat release rate clip plane showing three flames in the bottom of the bowl region for the propane addition case.

The growth of these three flames during the uncontrolled fast combustion event are explained using the planar cross section of local heat release rate in Figure 14. As the flames move toward each other in Figure 14(a) and (b), volumetric heat release begins to occur throughout the volume of gas in front of and between the natural gas flames, which are in the process of merging together. Volumetric heat release does not occur near the walls of the bowl due to heat loss to the walls. A quarter of a degree later in Figure 14(c) the rate of heat release throughout the volume in front of the merging natural gas flames has increased considerably, indicating volumetric ignition. This volumetric heat release occurs at 8.50° ATDC, which is the same timing as the peak in AHRR from the simulation. At 8.75° ATDC in Figure 6.14(d), this volume is no longer producing heat and the flames have traveled to the walls where they are slowed down by heat transfer. Therefore, the simulation has

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shown that the uncontrolled fast combustion rate observed in the propane addition experiment is likely due to volumetric ignition of natural gas in front of merging natural gas flames originating from adjacent diesel sprays.

Figure 14: Plots of local heat release rate in the bowl region during the uncontrolled fast combustion event for the D12G22P2 propane addition case.

A summary of agreement in combustion metrics between the engine experiments and simulations is given in Table 8. The ignition timing as indicated by the crank angle where 10% of the total integrated heat release has been reached (CA10) shows that the simulation was within a half crank angle at all operating conditions investigated. The simulations were able to predict the 50% heat release timing (CA50) within a half crank angle. The CA10-90 combustion duration, which is the crank angle difference between the 10% and 90% heat release timings, shows agreement within 5 crank angle degrees. Cases with larger error in combustion duration were the result of small disagreement in the heat release rate profile at the tail end of the combustion duration. The peak pressure was matched within 5 bar except for the D12G22P2 propane addition experiment where a 30


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slight over prediction in the magnitude of the fast combustion event lead to a greater increase in pressure. This fast combustion also lead to greater pressure rise rate error, which was otherwise within 1.1 bar/deg. The level of agreement between engine experiments and simulations shown in Table 8 demonstrates that the reduced mechanism is a useful model for dual fuel engine simulations.

Table 8: Comparison of experimental and simulated ignition delay and lift-off length.

Case Name D12G22, SOI = -10 deg. D12G22, SOI = -14 deg. D25G17, SOI = -8.8 deg D25G17, SOI = -13 deg. D11G8 D11G18 D12G22P2

CA10 Error (deg.) 0.41 0.15 0.50 0.13 0.25 0.13 0.14

CA50 Error (deg.) 0.03 0.11 0.37 0.14 0.39 0.23 0.12

CA10-90 Error (deg.) 1.10 1.60 1.12 4.75 4.00 1.47 2.51

Peak Pressure Error (bar) 0.06 2.63 0.12 0.22 1.70 4.43 6.62

Pressure Rise Rate Error (bar/deg) 0.77 0.56 0.05 1.09 0.97 0.69 5.52

6. CONCLUSIONS A reduced chemical kinetic mechanism consisting of 141 species and 709 reactions has been formulated for natural gas/diesel dual fuel combustion whereby n-heptane is used to represent diesel fuel and a mixture of methane, ethane, and propane represents natural gas. The final reduced mechanism is a combination of reduced versions of the detailed n-heptane mechanism from Curran et al.15,16 and the detailed methane through n-pentane mechanism from Healy et al.14. Ignition delay plots for n-heptane showed excellent agreement between the reduced CSU141 mechanism and the detailed LLNL mechanism across pressures from 10 to 80 bar, temperatures from 700 to 1400 K, and equivalence ratios from very lean (φ = 0.2) to rich (φ = 2.0). The natural gas chemistry was validated using ignition delay plots for two different natural gas compositions with varying ethane and propane contents and excellent agreement with the detailed NUIG mechanism was achieved at lean to stoichiometric conditions and pressures between 10 and 80 bar. However, the detailed mechanisms have only been validated with experimental data up to 30 bar, and therefore the agreement at 80 bar assumes the rate 31


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constants in the detailed mechanisms are accurate at higher pressure. Additionally, the reduced mechanism achieved excellent agreement with detailed mechanisms for initial temperature sweeps of adiabatic HCCI simulations for both n-heptane and two natural gas mixtures. Furthermore, the CSU141 mechanism was able to replicate experimental laminar flame speeds for lean CH4/He/O2 mixtures at 40 and 60 atm to within 16% relative error. Finally, the n-heptane chemistry in the CSU141 mechanism was able to predict the experimental ignition delay and lift-off length from a reacting diesel spray in a constant volume chamber at engine TDC conditions to within 14% and 20% error respectively. Comparisons between natural gas/diesel dual fuel engine experiments and multi-dimensional CFD simulations showed excellent pressure and AHRR agreement for two start of injection timing sweeps with different substitution percentages, a natural gas load sweep at fixed diesel injection, and an engine knocking case using additional propane to increase mixture reactivity. The simulations sufficiently captured the start of combustion timing, the pressure rise rate, the shape and magnitude of the heat release rate profile, and the combustion duration. In the engine knocking case, the simulation was able to predict the timing and magnitude of a sudden acceleration in combustion rate as well as the high frequency pressure oscillations. In-cylinder images from the simulation indicate that this fast combustion was due to volumetric heat release in the bottom of the bowl where propagating natural gas flames originating from adjacent spray axes merge together. Some of the simulations showed a small discrepancy in peak pressure and expansion pressure, but this could be due to error sources other than the chemical kinetics. Such sources of error include uncertainty in air and fuel mass flow rates, solving for the ensemble averaged flow using a RANS turbulence model, assuming homogenous composition and temperature at IVC, and the sector geometry assumption that combustion behavior is perfectly symmetric with respect to the spray axis. Despite these sources of error, the agreement achieved from coupling the reduced mechanism with the spray models and the RANS turbulence model has demonstrated that the reduced mechanism is a useful tool for capturing the important chemical kinetics in diesel ignition, natural gas

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flame propagation, and natural gas auto-ignition occurring in dual fuel engines. It particular, the reduced mechanism was able to accurately model varying natural gas reactivity without relying on rate constant tuning. Future validation work with natural gas/diesel RCCI engine experiments is necessary to confirm if the CSU141 mechanism is capable of modeling premixed auto-ignition of large reactivity gradients where the diesel/natural gas mixtures can vary from slightly rich to very lean. In addition, it should be reiterated that the CSU141 mechanism makes the assumption that heavier than propane hydrocarbons in the natural gas mixture are low enough in concentration that their effect on ignition delay and flame speed is negligible. Future work is necessary to develop a reduced dual fuel mechanism that includes the chemical kinetics of heavier hydrocarbons, such as butane, for dual fuel applications that require operation with a low methane number natural gas.

ASSOCIATED CONTENT Supporting Information. Included are additional details about validation of the spray model with experimental data from the ECN as well as the CHEMKIN® formatted files for the reduced CSU141 mechanism (species, reactions, thermodynamic data, and transport data). This material is available free of charge via the Internet at http://pubs.acs.org.

AUTHOR INFORMATION Corresponding Author *Andrew G. Hockett, [email protected]

ACKNOWLEDGEMENTS

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Thank you to Woodward Inc. for funding of this project and providing access to their engine test facility. Particular appreciation is extended to Jason Barta for helping conduct the engine experiments. The authors are grateful to Convergent Science, Inc. for their continuous technical support with the engine simulations.

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and Flame Speed Kinetics of Two Natural Gas Blends with High Levels of Heavier Hydrocarbons, Proc. ASME Turbo Expo, 2008, 3, 1051. (9)

Rahimi, A.; Fatehifar, E.; Saray, R.K. Development of an optimized chemical kinetic mechanism for

homogeneous charge compression ignition combustion of a fuel blend of n-heptane and natural gas using a genetic algorithm, Proc. Inst. Mech. Engineers, Part D: J. Automobile Engineering, 2010, 224, 1141-1159. (10) Smith, G.P.; Golden, D.M.; Frenklach, M.; Moriarty, N.W.; Eiteneer, B.; Goldenberg, M.; Bowman, C.T.; Hanson, R.K.; Song, S.; Gardiner, W.C.; Lissianski, V.V.; Qin, Z.; GRI-Mech homepage, http://www.me.berkeley.edu/gri_mech/. (11) Senecal, P.; Pomraning, E.; Richards, K.; Som, S. Grid-convergent spray models for internal combustion engine CFD simulations, Proc. ASME ICEF, 2012, ICEF2012-92043. (12) Dagaut, P. On the kinetics of hydrocarbons oxidation from natural gas to kerosene and diesel fuel, Phys. Chem. Chem. Phys., 2002, 4, 2079–2094. (13) Turbiez, A.; El Bakali, A.; Pauwels, J.F.; Rida, A.; Meunier, P. Experimental study of a low pressure stoichiometric premixed methane, methane/ethane, methane/ethane/propane and synthetic natural gas flames, Fuel, 2004, 83, 933–941. (14) Healy, D.; Kalitan, D.; Aul, C.; Petersen, E.; Bourque, G.; Curran, H.J. Oxidation of C1-C5 Alkane Quinternary Natural Gas Mixtures at High Pressures, Energy and Fuels, 2010, 24, 1521-1528.

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(15) Curran, H.J.; Gaffuri, P.; Pitz, W.J.; Westbrook, C.K. A Comprehensive Modeling Study of n-Heptane Oxidation, Combust. Flame, 1998, 114, 149-177. (16) Curran, H.J.; Gaffuri, P.; Pitz, W.J.; Westbrook, C.K. A Comprehensive Modeling Study of iso-Octane Oxidation, Combust. Flame, 2002, 129, 253-280. (17) Niemeyer, K.E.; Sung, C.-J.; Raju, M.P.; Skeletal mechanism generation for surrogate fuels using directed relation graph with error propagation and sensitivity analysis, Combustion and Flame, 2010, 157, 1760-1770. (18) Hockett, A. A Computational and Experimental Study on Combustion Processes in Natural Gas/Diesel Dual Fuel Engines, Ph.D. dissertation, Colorado State University, Fort Collins, CO, 2015. (19) CONVERGE™ 2.2.0 Theory Manual, Convergent Science Inc., 2014. (20) Pomraning, E.; Richards, K.; Senecal, P. Modeling Turbulent Combustion Using a RANS Model, Detailed Chemistry, and Adaptive Mesh Refinement, SAE Tech. Pap. Ser., 2014, 2014-01-1116. (21) Han, Z.; Reitz, R.D. Turbulence Modeling of Internal Combustion Engines Using RNG k-ε models, Combust. Sci. Technol., 1995, 106, 267-295. (22) Amsden, A. KIVA 3-V: A Block Structured KIVA Program for Engines with Vertical or Canted Valves, Los Alamos National Laboratory, 1997, Report No. LA-13313-MS. (23) Engine Combustion Network, Diesel Spray Combustion homepage, Sandia National Laboratories, http://www.sandia.gov/ecn/dieselSprayCombustion.php, (accessed 2014 through 2015). (24) Healy, D.; Curran, H.J.; Simmie, J.M.; Kalitan, D.M.; Zinner, J.M., Barrett, A.B.; Petersen, E.L.; Bourque, G. Methane/ethane/propane mixture oxidation at high pressures and at high, intermediate, and low temperatures, Combust. Flame, 2008, 155, 441-448. 36


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(25) Rozenchan, G.; Zhu, D.L.; Law, C.K.; Tse, S.D. Outward Propagation, Burning Velocities, and Chemical Effects of Methane Flames up to 60 atm, Proc. Comb. Inst., 2002, 29, 1461-1469. (26) Martinez, D.; Tozzi, L.; Marchese, A.J. A Reduced Chemical Kinetic Mechanism for CFD Simulations of High BMEP, Lean-Burn Natural Gas Engines, Proc. ASME ICES, 2012, ICES2012-81109. (27) Bajaj, C.; Ameen, M.; Abraham, J. Evaluation of an Unsteady Flamelet Progress Variable Model for Autoignition and Flame Lift-Off in Diesel Jets, Combust. Sci. Technol., 2013, 185, 454-472.

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Figure 1: Procedure used for reducing detailed mechanisms and combining reduced mechanisms. 190x54mm (300 x 300 DPI)


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Energy & Fuels

Figure 2: (a) 1/7th sector geometry used for engine combustion simulations. (b) Cylinder and port geometry from the GM 1.9L engine used to obtain the velocity and turbulence fields at IVC for use in the sector simulations. The cylinder liner is hidden to show the piston geometry. 101x54mm (300 x 300 DPI)


Energy & Fuels

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Figure 3. Ignition delay plots for n-heptane/air mixtures comparing the CSU141 reduced mechanism with the detailed LLNL mechanism15,16 for (a) very lean φ = 0.2, (b) lean φ = 0.5, (c) stoichiometric φ = 1.0, and (d) rich φ = 2.0 mixtures. 158x120mm (300 x 300 DPI)


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Energy & Fuels

Figure 4: Comparisons of natural gas/air ignition delay calculations between the reduced CSU141 mechanism and the detailed NUIG mechanism14. The 90/6.7/3.3 mixture is shown in the left column for (a) very lean φ = 0.2, (b) lean φ = 0.5, and (c) stoichiometric φ = 1.0 mixtures. The right column shows the 70/20/10 mixture at (d) very lean φ = 0.2, (e) lean φ = 0.5, and (f) stoichiometric φ = 1.0 mixtures. 171x181mm (300 x 300 DPI)


Energy & Fuels

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Figure 5: Adiabatic zero-dimensional HCCI engine simulation comparisons between the detailed LLNL and reduced CSU141 mechanism for varying initial temperatures of n-heptane/air mixtures at (a) φ = 0.5 and (b) φ = 2.0. 165x69mm (300 x 300 DPI)


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Energy & Fuels

Figure 6: Zero-dimensional adiabatic HCCI engine simulation comparisons between the detailed NUIG14 and reduced CSU141 mechanism for varying initial temperatures at φ = 0.5 for (a) the 90/6.7/3.3 natural gas mixture in air and (b) the 70/20/10 natural gas mixture in air. 165x69mm (300 x 300 DPI)


Energy & Fuels

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Figure 7: Comparisons of experimental laminar flame speeds with CSU141 computations for CH4/O2/He mixtures at (a) 40 atm and (b) 60 atm. The experimental data is from Rozenchan et al. The indicated O2 % refers to the molar composition in the O2/He oxidizer gas. 89x63mm (300 x 300 DPI)


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Energy & Fuels

Figure 8: Simulation result for OH mass fraction at 2 ms. The lift-off length is defined at the first location of 2% of the maximum OH mass fraction. 127x42mm (300 x 300 DPI)


Energy & Fuels

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Figure 9: Comparisons of experimental and computational pressure and AHRR for (a) the D25G17 SOI sweep and (b) the D12G22 SOI sweep. 184x69mm (300 x 300 DPI)


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Energy & Fuels

Figure 10: Time development of temperature in the plane of the spray axis for the D25G17, -13° ATDC SOI case. Time proceeds sequentially from (a) to (f). 162x66mm (300 x 300 DPI)


Energy & Fuels

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Figure 11: Comparisons of experimental and computational pressure and AHRR for the dual fuel load sweep using a constant diesel injection of 11 mg. 89x70mm (300 x 300 DPI)


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Energy & Fuels

Figure 12: Comparisons between experiment and simulation for the additional propane case where engine knocking was observed. (a) 100 cycle averaged experimental pressure and volume averaged simulation pressure. (b) Single cycle experimental pressure and simulation pressure at a single cell with the same location as the pressure transducer tip. 171x76mm (300 x 300 DPI)


Energy & Fuels

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Figure 13: Development of a 2000 K temperature iso-surface with a local heat release rate clip plane showing three flames in the bottom of the bowl region for the propane addition case. 107x107mm (300 x 300 DPI)


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Energy & Fuels

Figure 14: Plots of local heat release rate in the bowl region during the uncontrolled fast combustion event for the D12G22P2 propane addition case. 107x95mm (300 x 300 DPI)


Development and Validation of a Reduced Chemical Kinetic

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