The current work presents a recent development of the Extended Coherent Flamelet Model (ECFM) for 3D combustion modeling in spark-ignited gasoline engines. The reference-based ECFM model, originally published in 2003, computes the conditional unburned and burned gas species mass fractions from both real species and species tracers. This current work is motivated by two limitations of the reference-based model. First, the difference between convection of species tracers and convection of real species leads to small discrepancies between the two, due to high velocity gradients during gas exchange. This can lead to inaccurate estimation of the progress variable and consequently to negative conditional mass fractions in the burned gases after ignition. Second, the reference-based ECFM model assumes implicitly that the unburned and burned states correspond to the same mixture fraction. This assumption is valid for low stratification cases, but it can lead to substantial conditioning errors for highly stratified systems like gasoline direct injection (GDI) engines. To address these shortcomings, a new species-based ECFM (SB-ECFM) implementation is presented. In this species-based model, the unburned and burned gas states are entirely defined by the transported species in each zone. It is shown that SB-ECFM more reliably defines conditional quantities and the progress variable. This enhancement allows the use of a second-order central scheme in space when using full decoupling of auto-ignition and premixed flame progress variables as proposed in Robert et al., Proc. Comb. Inst, 2015, while the reference model is limited to the first-order upwind scheme in this case. Finally, simulations of a GDI engine are presented at different loads and rpm conditions. It is shown that, with the higher order scheme, SB-ECFM demonstrates very good agreement with measured pressure.

During the last several decades, investigations of the operation of internal combustion engines utilizing exhaust gas recirculation (EGR) has increased. This increased interest has been driven by the advantages of the use of EGR with respect to emissions and, in some cases, thermal efficiency. The current study uses a thermodynamic engine cycle simulation to explore the fundamental reasons for the changes of thermal efficiency as functions of EGR. EGR with various levels of cooling are studied. Both a conventional (throttled) operating condition and a high efficiency operating condition are examined. With no EGR, the net indicated thermal efficiencies were 32.1% and 44.6% for the conventional and high efficiency engines, respectively. For the conditions examined, the cylinder heat transfer is a function of the gas temperatures and convective heat transfer co-efficient. For increasing EGR, the gas temperatures generally decrease due to the lower combustion temperatures. For increasing EGR, however, the convective heat transfer coefficient generally increases due to increasing cylinder pressures and decreasing gas temperatures. Whether the cylinder heat transfer increases or decreases with increasing EGR is the net result of the gas temperature decreases and the heat transfer coefficient increases. For significantly cooled EGR, the efficiency increases partly due to decreases of the heat transfer. On the other hand, for less cooled EGR, the efficiency decreases due at least partly to the increasing heat transfer. Two other considerations to explain the efficiency changes include the changes of the pumping work, and the specific heats during combustion. For the constant loads considered, as more EGR is used, the throttle is opened which decreases the pumping work. For certain throttled cases, for increasing EGR, the increase of the cylinder heat transfer is somewhat offset by the decrease of the pumping work. For the high efficiency operating condition (with less throttling), the benefits of decreasing pumping work are much less. In all cases, the decreasing temperatures result in decreasing specific heats which provide some improvement in thermal energy conversion to work.

Generally, the turbulent-flame velocity of natural gas is significantly lower than diesel in the combustion process, which results in the thermal loads of natural gas engines being significantly higher than those of diesel engines under the same stoichiometric condition without EGR. In this study, a heavy-duty natural gas engine is taken as the research object, which is used to measure the temperatures to analyze the heat transfer characteristics in the cylinder head water jacket around the valve bridges, under different speeds and loads, as well as different coolant temperatures and pressures. Twelve thermocouples are inserted by drilling through the metal in the cylinder head with different heights to measure the metal temperatures at thermally critical areas such as the valve-bridge regions. Therefore, the local heat flux and the extrapolation to coolant wall temperatures are obtained by Fourier’s Law under different engine operating conditions. In addition, the thermal balance tests of the engine are also carried out, and the energy distributions are analyzed in different parts of the engine. The results of the research show that: a) the engine cooling condition has a direct impact on the engine cylinder head temperature. If the cooling temperature is low, the temperatures of the cylinder head’s measuring points have the same increases as the increasing coolant temperature. When the coolant temperature is high, the measuring temperatures have hardly any difference from the increases in cooling temperature. With increasing cooling pressure, the temperature increase at all measuring points, and the temperature of the measuring points varies substantially under high load conditions compared with the low load condition. The results indirectly indicated that local nucleate boiling appeared in the water jacket. b) The heat transfer characteristic curve of the water jacket was obtained from the processing of experimental data. Wall heat flux increases with increasing load, and the relationship between wall heat flux and wall temperature is no longer linear. The heat transfer characteristic curve indicates that the convective heat transfer and boiling heat transfer both appeared in the cooling water jacket. c) With the decrease of engine load, the percentage of crankshaft power in the combustion heat gradually decreases, then the percentage of the heat taken away by the cooling water increases gradually. At the same time, the percentage of the heat taken away by exhaust has changed little. d) The engine cooling temperature has a substantial influence on the engine thermal balance, and the cooling pressure has little effect on the engine thermal balance. With increasing cooling temperature, the heat taken by the cooling water decreased, which lead to an increase in the proportion of crankshaft power. It can be concluded that properly increasing the coolant temperature of the engine can improve the fuel economy of the machine.

The use of exhaust gases as a diluent for IC engine mixtures has been shown to be advantageous for lower emissions for certain operating conditions. These exhaust gases may be introduced via an external circuit (perhaps including cooling), or they may be retained (internal) in the cylinder by the use of certain valve timings. Three (3) types of dilution processes were examined in this work: external EGR, negative valve overlap, and delayed exhaust valve closing. The first process is based on external dilution, and the last two processes are based on internal dilution. Thermodynamic evaluations for an automotive engine were completed for a constant load (bmep = 900 kPa) and constant speed (2000 rpm). The thermal efficiencies decreased for all three dilution processes in similar fashion as the burned gas fraction increased. This was shown to be a result of increasing heat losses for increasing burned gas fraction. The calculations showed that even though the gas temperatures decrease with the use of exhaust gas dilution, the cylinder heat transfer actually increases. This was shown for six different heat transfer correlations typically used for these calculations. This is a result of the dependences of the convective heat transfer on cylinder gas pressure and temperature. For the cases examined, cylinder pressures increased and cylinder temperatures decreased as the dilution level increases. In all cases examined, the net result was an increase of the convective heat transfer coefficient which dominated the decrease of temperatures.

A novel experimental procedure is presented which allows simultaneous identification of heat and work transfer parameters for turbocharger compressor models. The method introduces a thermally transient condition and uses temperature measurements to extract the adiabatic efficiency and internal convective heat transfer coefficient simultaneously, thus capturing the aerodynamic and thermal performance. The procedure has been implemented both in simulation and experimentally on a typical turbocharger gas stand facility. Under ideal conditions, the new identification predicted adiabatic efficiency to within 1%point and heat transfer coefficient to within 1%. A sensitivity study subsequently showed that the method is particularly sensitive to the assumptions of heat transfer distribution pre and post compression. If 20% of the internal area of the compressor housing is exposed to the low pressure intake gas, and this is not correctly assumed in the identification process, errors of 7–15%points were observed for compressor efficiency. This distribution in heat transfer also affected the accuracy of heat transfer coefficient which increased to 20%. Thermocouple sensors affect the transient temperature measurements and in order to maintain efficiency errors below 1%, probes with diameter of less than 1.5mm should be used. Experimentally, the method was shown to reduce the adiabatic efficiency error at 90krpm and 110krpm compared to industry standard approach from 6% to 3%. However at low speeds, where temperature differences during the identification are small, the method showed much larger errors.

The project investigates the effects of peanut FAME on diesel engine combustion and thermal efficiency. The cold flow properties and viscosity were tested and were found that the cloud point (CP) and pour point (PP) of peanut FAME were 17°C and 8°C respectively, and was able to achieve CP of 0°C when blended 20:80 (wt%) with diesel No. 2 (P20). The dynamic viscosity of peanut FAME was 4.2cP (P100) and 2.85cP at 54°C (P20), both fuels are within the ASTM standard for biodiesel. The lower heating value (LHV) of peanut FAME was 37.10MJ/kg (P100) and 41.3MJ/kg (P20) compared to 41.7MJ/kg for diesel No.2 (D100), which supports the usage of peanut FAME in compression ignition engines. At residence time of 5ms from start of injection, P50 has shown positive combustion characteristics with ignition delay of 1.072ms at 2600rpm, 4.78 bmep (100% load). The P50 heat release displayed similar development compared with diesel No. 2, where premixed phase combined with diffusion combustion and reaching a maximum of 20.0J/CAD, which was higher than 17.5J/CAD for D100. Convection flux for both D100 and P50 had values of 1.4MW/m 2 . The total heat flux, calculated by Annand model, produced maximum values of 2.1MW/m 2 for D100 compared with 2.3MW/m 2 for the P50. The mechanical efficiency was only a 4% loss when observing the transition from D100 to P50. These findings support peanut FAME as a viable option when blended and used with diesel engines in order to meet the standards set forth by the RSF-2 and EISA allowing the U.S. to decrease foreign energy dependency and benefiting society through a cleaner burning fuel than is currently in use.

The US Army Single Fuel Forward policy mandates that deployed vehicles must be operable with aviation fuel JP-8. Therefore, an investigation into the influence of JP-8 on a diesel engine’s performance is currently in progress. The injection, combustion, and performance of JP-8, 20–50% by weight in diesel no.2 mixtures (J20-J50) produced at room temperature were investigated in a 77mm indirect injection, high compression ratio (23.5) diesel engine, in order to evaluate its effectiveness for application in Auxiliary Power Units (APUs) at 2000rpm continuous operation (100% load/BMEP 4.78 bar). Due to the viscosity requirements for proper injection the new fuel can contain as high as 100% JP-8 (J100). The blends had an ignition delay of 1.03ms regardless of the amount of JP-8 introduced. J50 and diesel no.2 exhibited similar characteristics of heat release, the premixed phase being combined with the diffusion combustion. The maximum combustion pressure remained relatively constant for all blends, 72.7bar for diesel and decreased slightly by 0.40bar for J50, with the peak pressure position being delayed by 0.5CAD for the J50. The instantaneous volume-averaged gas combustion temperature reached 2162K for diesel versus 2173K for J50; displaying a 1.2CAD delay in the position of the maximum temperature and retaining the higher temperature for a longer duration for J50. The heat flux in the engine cylinder exhibited comparable maximum values for all blends (diesel: 2.12MW/m 2 , J50: 2.14MW/m 2 ). The cylinder heat losses were at a minimum during combustion before TDC with increased convection losses at TDC for all fuels and the beginning of the power stroke. The heat losses associated with the system increased slightly with the addition of JP-8. The BSFC for diesel no.2 was 242(g/kW/hr) and increasing by only 0.7% for J50. The engine’s mechanical efficiency displayed similar values for all blends, 83% and decreasing by only 1% for J50. Taking into account each fuels’ corresponding density, the engine’s overall efficiency remained relatively constant at 29% with the addition of the JP-8. The engine investigation demonstrated that up to 50% JP-8 by weight in diesel can be injected and burnt in a small diesel engine with a combustion duration of approximately 5ms, while maintaining the engine overall efficiency. The study validates JP-8 as an excellent source for power generation in a diesel APU based on its combustion characteristics. The next stage of research shall be the full emissions investigation.

Glow plugs are a possible ignition source for direct injected natural gas engines. This ignition assistance application is much different than the cold start assist function for which most glow plugs have been designed. In the cold start application, the glow plug is simply heating the air in the cylinder. In the cycle-by-cycle ignition assist application, the glow plug needs to achieve high surface temperatures at specific times in the engine cycle to provide a localized source of ignition. Whereas a simple lumped heat capacitance model is a satisfactory representation of the glow plug for the air heating situation, a much more complex situation exists for hot surface ignition. Simple measurements and theoretical analysis show that the thickness of the heat penetration layer is small within the time scale of the ignition preparation period (1–2 ms). The experiments and analysis were used to develop a discretized representation of the glow plug domain. A simplified heat transfer model, incorporating both convection and radiation losses, was developed for the discretized representation to compute heat transfer to and from the surrounding gas. A scheme for coupling the glow plug model to the surrounding gas computational domain in KIVA-3V was also developed. The glow plug model successfully simulates the natural gas ignition process for a direct injection natural gas engine. As well, it can provide detailed information on the local glow plug surface temperature distribution, which can aid in the design of more reliable glow plugs.

The authors investigated the formulation, combustion and emissions of polypropylene (PP)–diesel fuel mixtures in a direct injection diesel engine. The fuel has been obtained by an original technology they developed, in which the low or high density polypropylene (LDPP, HDPP), have been mixed in a nitrogen atmosphere at 200 °C, 10–40% by wt. in diesel fuel. The kinematic viscosity of the polypropylene-diesel fuels was investigated between 25–250 °C and the results showed that viscosity of the plastic mixtures is much higher than that of diesel alone, ranging from 10 cSt to 500 cSt, and depending on the plastic structure, content, and temperature. The TGA and DTA analysis has been conducted to investigate the oxidation and combustion properties of pure PP and polymerdiesel fuels. The results showed that at about 125 °C, the LDPP melts, but does not decompose up 240 °C, when the oxidation starts, and has a peak of heat release at 340–350 °C, and the process is completed at 400 °C. The engine’s injection system used, was a piston-barrel type pump, capable of an injection pressure of 200 bars. The injector had 4 × 0.200 mm nozzles with a conical tip needle. The 25% PP-diesel mixture had a successful ignition in a direct injection 110 mm bore, omega combustion chamber engine. The ignition delay for polypropylene-diesel mixtures was longer by about 0.5 ms (at 1200 rpm), compared with diesel. The heat release showed a different development compared with the reference diesel fuel, the premixed phase being inhibited while a slow diffusion combustion phase fully developed. The maximum combustion pressure has been 83 bars for diesel and decreased by 2 bars for the blended fuel, while the bulk gas maximum temperature (calculated) reached about 2500 K for diesel vs 2600 K for polypropylene mixture. The heat flux calculated by the Annand model has shown lower values for diesel fuel with a maximum of about 2.7 MW/m 2 compared with 3.0 MW/m 2 for PP blended fuel with similar values for convection flux for both fuels at about 1.57 MW/m 2 and a higher radiation flux of about 1.44 MW/m 2 for PP fuel versus 1.27 MW/m 2 for diesel. The heat lost during the cycle shows low values for the premixed combustion stage and increased values for the diffusion stage for both fuels. The exhaust temperatures have been practically identical for both fuels for all loads, with emissions of NOx, and CO reduced by 40% for the alternative fuel, while the CO 2 exhibited almost the same values for both fuels. The smoke emissions decreased by 60–90% for the polypropylene blended fuel depending on the load, The engines’ overall efficiency was slightly lower for PP fuel at low loads compared with diesel combustion but at 100% load both reached 36%. The study showed that the new formulation process proposed by the authors is able to produce a new class of fuels from diesel blended with low density polypropylene, and resulted in hybrid fuels with very promising combustion prospects. The engine investigation proved that 25% PP fuels can be injected and burnt in a diesel engine at a residence time of about 5 ms from the start of injection, and the engine’s nominal power could be reached, with lower emissions than reference diesel fuel.

One of the most efficient ways to reduce the pollution and fuel consumption of an automotive engine is to downsize the engine, whilst maintaining a high level of power and torque. This is achieved by using turbochargers. In urban, and often in suburban, traffic conditions the engine power demand is weak in relation to the maximum power available, so the turbocharger runs at low speed. To appreciate and improve engine performance, it is necessary to know the characteristics of the turbomachinery in this functioning area, characteristics which are not given by turbocharger manufacturer. The reason for this lack of information will be explained and the experiments we are currently conducting at low turbocharger speed are presented. Experimentally, it has been demonstrated that the measured performances of the compressor are dependent on heat exchange (convection and conduction) and are also linked to the pressure and temperature of the lubricating oil. At the CNAM laboratory, the turbocharger test rig has been equipped with a special torquemeter, allowing rotation speeds of up to 120000 rpm, set up between the turbine and the compressor. The turbine is thus separated from the compressor and could be considered as a drive which provides mechanical power to the turbocharger (torquemeter + compressor + bearing unit). Temperature and pressure of the lubricating oil can be adjusted to an experiment’s requirements. The test bench lay out is described. To achieve accurate measurements and evaluate the influence of heat exchanges, tests have been carried out with the whole compressor thermally isolated and with preheated inlet air. The compressor can be assumed to be adiabatic, and the power given to the air flow can be calculated using the first law of thermodynamics. Mechanical bearing losses can be deduced from this calculation and torquemeter power, but also from measurements of lubricating oil flow, and oil temperature at inlet and outlet. The results of experiments for different lubricating oil temperatures and pressures and turbocharger speeds are presented. Real compressor characteristics curves are set up and a comparison of experimental mechanical power losses with a journal bearing CFD model is presented.

The quest for increasing the efficiency of an internal combustion engine has been going on ever since the invention of this reliable work equipment. In recent times much attention has been focused on achieving higher thermal efficiency by reducing the heat loss to the surroundings by the application of “low heat rejection” concept using ceramic coating for engine components or ceramic components itself. In this direction, the present analysis focused on the formulation and development of a two zone combustion model for the prediction of combustion, heat release and performance of the conventionally Water cooled and low heat rejection engines. The low heat rejection concept is introduced in this model by developing and formulating a wall heat transfer model to consider the effect of instantaneous inside surface temperature on combustion and heat release of an LHR engine. The gas wall heat transfer by means of convection and radiation is calculated using various gas wall heat transfer models. The predictions by the computer model and experimental findings are demonstrated. The capability of this model and the comparison of the predicted results with experimental findings are highly satisfactory.

The catalytic converter is an important device for the emission control from spark-ignition engines. Several concurrent physical/chemical processes such as convective heat transfer, gas phase chemical reactions, surface reactions, flow oscillations, water vapor condensation and diffusion mechanisms add complexity to modeling of flows inside catalytic converters. Under cold-start conditions, the fact that catalytic converters do not become operational during the initial operation allows a significant fraction of the overall pollutants to be emitted. In the present study, these complex transient phenomena have been examined using a previously validated numerical model. 1 The numerical results suggest new material-dependent designs to improve both the transient conversion characteristics and the steady state conversion efficiency of catalytic converters. Moreover, from our model calculations, we have observed that for a given amount of the noble metal catalysts the light-off time and the monolith temperature are greatly affected by the noble-metal distribution along the honeycomb walls of a monolith. The results of the numerical simulations indicate that the light-off time is shortened by approximately 35% for CO, H 2 and C 3 H 6 when replacing a traditional homogeneous noble metal distribution by a simple, step-function distribution. 2 The emissions of CO, H 2 and C 3 H 6 from the exhaust gas are, therefore, reduced without increasing the cost of noble metal catalysts used in converters. In order to avoid further deterioration of catalysts due to the thermal effects, an optimum noble metal distribution needs to be investigated with the understanding that the optimum noble metal distribution proposed has to be practical for the manufacturing. Since the main source of the exhaust emissions is generated during the cold-start period of the converter operation, the reduction of emissions shown in our model calculations is quite substantial.

A laser-based system should be advantageous to a spark-plug based ignition system. Free choice of the ignition spot and precise timing constitute two major advantages. Multi point laser ignition could lead to higher efficiencies, and laser ignition as such is capable of igniting leaner mixtures than a spark plug, thereby decreasing thermal NO x and soot emissions. This paper is devoted to advances in optical diagnostics of laser ignition for future internal combustion engines. The focus of this paper is on diagnostics at high pressures, that is engine-like conditions. Laser ignition tests were performed with the fuels methane, hydrogen and biogas in static combustion cells with dimensions comparable to stationary engines. A Nd:YAG laser (5 ns pulse duration, wavelength 1064 nm, 1–20 mJ pulse energy) was used to ignite gaseous fuel/air mixtures at initial pressures of 1–3 MPa. Schlieren photography and laser-induced fluorescence (LIF) were used for optical diagnostics (flame kernel development, shock wave propagation). The lean burn characteristics were investigated. Schlieren photography was used to determine the velocity of the shock wave and to study the influence of the shock wave on temperature rise and energy loss. Using planar laser-induced fluorescence (PLIF), the spatial distribution of the combustion intermediates OH and formaldehyde were recorded. The temporally resolved imaging shows that the initial stages of the flame front evolution closely follows the turbulence and density fluctuations caused by the shock and pressure wave induced by the laser spark. In this paper, results from LIF spectroscopy and Schlieren photography are compared. Depending on the laser pulse energy and focus size, at later stages after the ignition the flame front propagation approaches the laminar burning regime and flame front speed decrease. Flame front break up at lean conditions indicates the limit of the ignitable mixture fraction when the speed due to spark-induced convection exceeds the flame propagation rate.