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Hydrogen Engine

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A hydrogen engine is a type of internal combustion engine that utilizes hydrogen as its primary fuel source, converting chemical energy from hydrogen into mechanical energy through combustion. This technology aims to reduce greenhouse gas emissions and reliance on fossil fuels, contributing to cleaner energy solutions in transportation and power generation.
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A hydrogen engine is a type of internal combustion engine that utilizes hydrogen as its primary fuel source, converting chemical energy from hydrogen into mechanical energy through combustion. This technology aims to reduce greenhouse gas emissions and reliance on fossil fuels, contributing to cleaner energy solutions in transportation and power generation.
Development Project of a Hydrogen Fuelled Internal Combustion Engine for Heavy Duty Truck - Outline and Preliminary Test
by Kimitaka Yamane

2025, Journal of KONES

The hydrogen fuelled internal combustion engine represent a promising solution to the issues of global warming, fossil fuel break point and environmental protection. The development of a hydrogen fuelled internal combustion engine for... more
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Advantages and disadvantages of composite pistons for small power combustion engines
by Jakub Wieczorek

2025, DOAJ (DOAJ: Directory of Open Access Journals)

In vorliegender Ausarbeitung wurden Ergebnisse der technologischen und metallographischen Forschungen von kleine Maße aufweisenden Kolben für Verbrennungsmotoren geringer Leistung beschrieben. Halbzeuge der Kolben wurden mittels... more
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IJTERI ISSN (Online): 2456-4788 HYDROGEN USE IN IC ENGINES: A REVIEW
by Bhupesh Kurrey

2024

The researchers are working towards to find out the best replacement for the fossil fuel; if not at least to offset the total fuel demand. In regards to emission, the fuel in the form of gaseous state is much than liquid fuel. By... more
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Influence of initial charge conditions on engine performance and emission of a DISI hydrogen-fueled engine under various injection timings
by Arash Nemati

2024, Turkish Journal of Engineering and Environmental Sciences

Using hydrogen as an alternative fuel in spark-ignition engines has great potential for reducing exhaust emissions. Hydrogen engines have CO2 and HC free combustion and lean operation, offering lower nitrogen oxide (NOx) emission. In this... more
Figure 2a. Model validation under injection timing of 130° CA BTDC, spark timing of 11° CA BTDC. Figure 2a. Model validation under injection timing of
Figure 1. 3D moving mesh at 40° CA BTDC.
Figure2c. Model validation under injection timing of 80° CA BTDC, spark timing of 8° CA BTD¢ Figure 2b. Model validation under injection timing of 100° CA BTDC, spark timing of 11° CA BTDC.
Figure 3. Comparison between calculated and measured NO; emission under injection timing of 130° CA BTDC, spark timing of 11° CA BTDC.
Figure 4b. Effect of initial charge temperature on cylin- der temperature under injection timing of 80° CA BTDC. Figure 4a. Effect of initial charge temperature on cylin- der pressure under injection timing of 80° CA BTDC. Figure 4a. Effect of initial charge temperature on cylin- Figure 4b. Effect of initial charge temperature on cylin-
Figure 5a. Effect of initial charge temperature on cylin- der pressure under injection timing of 100° CA BTDC. Figure 5a. Effect of initial charge temperature on cylin- der pressure under injection timing of 100° CA BTDC.
Figure 6a. Effect of initial charge temperature on cylin- der pressure under injection timing of 130° CA BTDC. Figure 6a. Effect of initial charge temperature on cylin-
Figure 6b. Effect of initial charge temperature on cylinder temperature under injection timing of 130° CA BTDC. Figure 5b. Effect of initial charge temperature on cylinder temperature under injection timing of 100° CA BTDC.
Figure 8. IMEP comparison for various initial charge temperatures. 4.2. Effect of initial charge pressure
Figure 7. NO, comparison for various initial charge temperatures.
Figure 9. ISFC comparison for various initial charge temperatures.
Figure 11b. Effect of initial charge pressure on cylinder temperature under injection timing of 100° CA BTDC. Figure 1la. Effect of initial charge pressure on cylinde1 pressure under injection timing of 100° CA BTDC. Figure 11b. Effect of initial charge pressure on cylinder Figure 1la. Effect of initial charge pressure on cylinder pressure under injection timing of 100° CA BTDC.
Figure 10b. Effect of initial charge pressure on cylinder temperature under injection timing of 80° CA BTDC. Figure 10a. Effect of initial charge pressure on cylinde pressure under injection timing of 80° CA BTDC. Figure 10a. Effect of initial charge pressure on cylinder Figure 10b. Effect of initial charge pressure on cylinder Figures 10a, 10b, lla, 11b, 12a, and 12b indicate the in-cylinder pressure and temperature at various initial charge pressures. As can be seen, by increasing the initial pressure, the mean and maximum in-cylinde1 pressure is increased. Figures 10b, 11b, and 12b show that in-cylinder temperature is inversely proportional to the initial charge pressure. This is due to the fact that with the increasing of the initial pressure, the inlet ait mass increases; however, the injected hydrogen mass is fixed. This implies that the fuel-air mixture becomes leaner than that at the naturally aspirated condition. Since nonluminous burning will occur in the lean mixture condition, this can explain why the in-cylinder temperature decreases with the increasing of the initial pressure. temperature under injection timing of 80° CA BTDC.
Figure 13. NO, comparison for various initial charge pressures. et al., 2007).
Figure 12b. Effect of initial charge pressure on cylinder temperature under injection timing of 130° CA BTDC. Figure 12a. Effect of initial charge pressure on cylinde1 pressure under injection timing of 130° CA BTDC. Figure 12b. Effect of initial charge pressure on cylinder Figure 12a. Effect of initial charge pressure on cylinder
Figure 14. IMEP comparison for various initial charge pressures.
Figure 15. ISFC comparison for various initial charge pressures.
Table. Engine specifications. engine and contained a maximum of 73,246 cells at a 130° CA BTDC. Different grid densities were tested prior to the study, and the present resolution was found to give adequately grid-independent results with a reasonable CPU run time. The AVL FIRE code solves the compressible, turbulent, 3-dimensional transient conservation equations for reacting multicomponent gas mixtures. The turbulent flow within the combustior chamber is simulated using the RNG k — « turbulence model, modified for variable-density engine flows (Har and Reitz, 1995), considered to be one of the suitable turbulence models for engine simulation studies. Initia conditions for turbulence kinetic energy (TKE) and turbulence length scale (TLS) were calculated according tc the formulation provided in the AVL FIRE User Manual (2006) and described below:
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Using Solar-Powered 3-Phase Squirrel Cage Induction Motors as Prime Movers for Synchronous Electric Generators
by Abdul Rauf Sulemana

2024

Most industries and communities in Ghana and around the world depend on synchronous electric generators powered mostly by fuel-run internal combustion engines to augment their energy deficit. Industries, communities and individuals... more
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Doppelstrahlsprays zur Verbesserung von Hochdruck-Dieselinjektorsystemen
by Franz Durst

2023, Mtz - Motortechnische Zeitschrift

Einzelloch-Injektoren für Dieselmotoren werden hinsichtlich der durchschnittlichen Tropfengröße und des Anfangszeitpunkts der Sprayausbildung betrachtet. Um eine bessere Sprayqualität und eine höhere Dynamik für Mehrfacheinspritzungen zu... more
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Experimental study of combustion performances and emissions of a spark ignition cogeneration engine operating in lean conditions using different fuels
by Mohamed Salah Boulahlib

2023, International Journal of Hydrogen Energy

This work presents an experimental study describing a six-cylinder spark ignition engine running with a lean equivalence ratio, high compression ratio, ignition delay and used in a cogeneration system (heat and electricity production).... more
‘ig. 1 — Experimental rig. 1: Engine; 2: Electrical motor; 3: Engine cooling unit; 4: Heat exchanger; 5: Panel control. 6: Thermal ircuit; 7: Fuel tanks 8: Exhaust gas analyser; 9: Engine chassis.
The study of the combustion products is carried out by an ECOM J2KN PRO analyzer incorporating a fumes suction pump, a 300 mm heating probe with triple function (com- bustion, and drawing temperature), electrochemical cells for Oz, CO, NO, NO». Chemical electrolyte cells are used for the Please cite this article in press as: Boulahlib MS, et al., Experimental study of combustion performances and emissions of a spar ignition cogeneration engine operating in lean conditions using different fuels, International Journal of Hydrogen Energy (2017), https:/ doi.org/10.1016/j.ijnydene.2017.10.041
Fig. 3 — NOx evolution versus electrical output.
Fig. 5 — CO evolution versus electrical output. The presence or not of carbon monoxide Is an indicator of the quality of combustion. Carbon monoxide acts as a pollutant gas and disrupts chemical reactions on the energy release aspect and oxygen use. Indeed, the formation of CO gives off less heat than the reaction forming CO2. Moreover, this reaction uses only one oxygen atom and thus promotes the formation of NOx, which are used for oxygen.The appearance of carbon monoxide implies a decrease in the amount of carbon dioxide (Fig. 5). The CO emissions are reduced by the use of pure methane and improve even better with the hydrogen addition in some regimes. The values of CO emissions of methane and methane-hydrogen mixture are very close. We note that natural gas gives a higher level of CO emissions regardless of the electrical energy delivered due to the presence of other polycarbonate gases. Combustion of NG (methane with addition of ethane and propane) occurs at higher temperatures than methane alone (Fig. 4). Carbon monoxide pollution is therefore much more important for natural gas. Combustion of methane is less polluting than NG, H»CH, (31%). The small difference between natural gas and methane can be explained by the difference in temperature obtained. The formation of nitrogen oxides is influenced by the maximum temperature during combustion and by the oxygen level. This suggests that the combustion temperature of CH, and NG is higher than H2CH, (Fig. 4), and that the excess oxygen caused by the formation of carbon monoxide allows nitrogen oxides to form. On the other hand, the emissions from the H,CH, are much less important. At constant pro- duced electric power, the proportion of NOx in emission is reduced by about 31% from the reference case of pure methane. The reaction mechanisms of hydrogen are different from those of carbon chains, so it is likely that the oxidation reaction of hydrogen is concurrent with that of nitrogen ox- ides. The present compression ratio is expected to produce more NO, as mentioned by Zhao et al. [53], but it appears that the effects of reducing ignition delay associated with a very lean condition are greater than those of the CR. The ignition delay, allows the fuel to burn late during the expansion, which results in lowering the combustion temperature which falls due to the dilution effects, which decreases the rate of NOx formation. The delay in ignition results in a delayed release of the heat of combustion which will be lost during the expan- sion phase and the increase in the volume of the cylinder. This will result in greater dilution and will result in a reduction in
HCs are complex compounds with a solid phase and a liquid phase. These particles are hazardous in the presence of polycyclic aromatic hydrocarbons (PAHs) and unburned CH,. The hydrocarbons originate from heterogeneous effects in the
hydrogen on HC emissions compared to NG and Chy are contradictory according to the operating parameters of the different engines. When the rate of excess air increases to- ward the lean burn limit associated with other conditions, HC
Fig. 9 - Maximum In pressure for used fuels.
Fig. 9 shows the evolution of the maximal internal cylinder pressure, for different fuels (NG, CH, and CH,/H,) at a stabi- lized speed of 1500 rpm. The maximal pressure is higher for the CH,/H2 mixture, indicating that the work is more impor- tant. The results of methane are also higher than that of Please cite this article in press as: Boulahlib MS, et al., Experimental study of combustion performances and emissions of a spark ignition cogeneration engine operating in lean conditions using different fuels, International Journal of Hydrogen Energy (2017), https:// doi.org/10.1016/j.ijhydene.2017.10.041
Table 1 — Engine specifications.
Table 2 — Accuracies of measurements. Please cite this article in press as: Boulahlib MS, et al., Experimental study of combustion performances and emissions of a spark ignition cogeneration engine operating in lean conditions using different fuels, International Journal of Hydrogen Energy (2017), https:// doi.org/10.1016/j.ijnydene.2017.10.041
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Using of an Electrochemical Compressor for Hydrogen Recirculation in Fuel Cell Vehicles
by Prof. Dr.-Ing. Thomas von Unwerth

2023, Fuel Cells

The automotive industry sees hydrogen‐powered fuel cell (FC) drives as a promising option with a high range and short refueling time. Current research aims to increase the profitability of the fuel cell system by reducing hydrogen... more
Fig. 2 Gas diffusion through the FC membrane.
Fig. | Hydrogen recirculation with jet pump and blower.
Fig.4 Schematic overview of an electrochemical hydrogen compressor
Fig. 3 Hydrogen recirculation with jet pump and EHC.
Fig.5 Current/voltage characteristic of the reference fuel cell stack.
Fig.6 Gas ratio of the anode of a reference fuel cell stack.
Fig. 7 Results of the simulation Study. Pumping hydrogen from the anode to the cathode with 1.13 Acm™ and 80°C.
g.8 Results of the simulation study. Reduction of nitrogen from 2% on te anode to ~40 ppm on the cathode by 1.13 Acm™ and 80°C.
Fig.9 Results of the experimental study. Mole fraction of 60% Hz gas mixture on the anode side and ~99.4% on the cathode side as a func- tion of the current at 80°C and 80% r.H.
Fig. 11 Results of the experimental study. Tests with relative humidity and temperature.
Fig. 10 Results of the simulated relative humidity of the EHC by 1.13 Acm™ and 80°C.
Fig. 13 Results of the experimental study. U/I characteristics of the EHC with variation of the gas composition at 80 °C and 80% r.H.
Fig. 12 Results of the experimental study. The relative humidity increases from anode to cathode with the increase of the partial pressure.
Fig. 14 Structure for the calculation model for a FC vehicle.
Fig. 16 Mechanical integration of EHC in FC system, Alternative B. Fig. 15 Mechanical integration of EHC in FC system, Alternative A.
Fig. 17 Relative humidity along the channel, Alternative B by 1.13 Acm? and 80°C.
Table 1 Vehicle model parameters.
Table 2 Comparison of the mean values in the new European driving cycle (NEDC) and WLTP.
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Numerically Modeling of Diesel Engine and Analysis the Effects of Double Injection Strategies on Performance and Pollutant Emissions
by Shahram Khalilarya

2023, International journal of automotive engineering

Modern diesel engines should have higher pollutant emissions standards with better performance and by using split injection strategies which could optimize the air – fuel mixture, this purpose could be achieved. After achieving the... more
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Modeling and Experiments on Mixture Formation in a Hydrogen Direct-Injection Research Engine
by Victor Hugo Mena Salazar

2023, SAE International Journal of Engines

Direct injection offers a large number of degrees of freedom, as it strongly influences the mixture stratification process. Experiments on a single cylinder research engine fuelled by H 2 , carried out at Argonne National Laboratory,... more
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Modeling and Experiments on Mixture Formation in a Hydrogen Direct-Injection Research Engine
by Victor julian Salazar

2023, SAE International Journal of Engines

Direct injection offers a large number of degrees of freedom, as it strongly influences the mixture stratification process. Experiments on a single cylinder research engine fuelled by H 2 , carried out at Argonne National Laboratory,... more
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Niedrig viskose Öle – eine Herausforderung für die Auslegung der Zylinderlaufbahn
by Andreas Götze

2023, Der Verbrennungsmotor - ein Antrieb mit Vergangenheit und Zukunft

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Retrofitting of auxiliary components to enhance the performance of a single-cylinder, air-cooled gasoline two-stroke engine
by fawzi ali

2023, Jurnal Mekanikal

A unique stepped-piston engine was developed by a group of researchers at UTM. The development work engulfs design, prototyping and testing over a predetermined period of time. The aim is to demonstrate environmental-friendly capability... more
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Effect of Cylinder-by-Cylinder Variation on Performance and Gaseous Emissions of a PFI Spark Ignition Engine: Experimental and 1D Numerical Study
by Gerardo Valentino

2023, Applied Sciences

Combustion stability, engine efficiency and emissions in a multi-cylinder spark-ignition internal combustion engines can be improved through the advanced control and optimization of individual cylinder operation. In this work,... more
Figure 1. Engine layout with LP EGR circuit. The engine is designed following the so-called “downsizing” concept. At high loads the engine works under knock-limited conditions and this requires to delay the combustion phasing (50% of mass fraction burned—MFB 50%), especially at lower speeds where a higher knock tendency usually occurs. On the other hand, at high speeds the A/F mixture has to be particularly enriched to maintain the Temperature at Turbine Inlet below a certain maximum allowable level. These control strategies, although mandatory for the examined engine, greatly penalize the fuel consumption at high loads. For this reason, the base engine configuration is properly modified for research purpose through the installation of an external low-pressure (LP) exhaust gas recirculation (EGR) circuit as depicted in Figure 1
ee As discussed in the introduction section, experiments carried out on the examined engine have highlighted significant differences in cylinders, mainly ascribed to a non-uniform effective Cylinder-by-cylinder variations are apparent in Figure cylinder pressure trace and the rate of heat release (RO he combustion evolution between in-cylinder air/fuel (A/F) ratio [15]. 2a which shows the experimental in- HR) for two representative operating points (3000@9 and 3000@13, 0% EGR). As aforementioned, the pressure data refer to the ensemble average over 270 consecutive cycles. For both operating conditions, the pressure curves of two cylinders are overlapped in the compression stage, denoting the same air volumetric efficiency, while Cyl #2 provides higher than Cyl #1. The same behavior is found at each inves pressure peak and combustion rate igated operating point, suggesting a systematic difference in pressure cycles. Further experimental tests on the adopted fuel injection system have variation in port injec collected through injec highlighted a different fuel supply to cylinders, resulting from a ors fuel rates (Figure 2b). The results plotted in Figure 2b were tion tests realized at am-bent air pressure with fuel injection system removed from the engine. A drive signal to injectors was used to simulate the injection timing; a certain num mass per stroke delivered by injec pressure traces shown in Figure 2a, fuel mass flow rate. In of two injectors was switched, ind ber of injection timings were considered to measure the gasoline ors. In particular, consistently with the in-cylinder he injector corresponding to Cyl #2 provides a higher erestingly, no difference was found also when the relative position icating the fuel rail geometry as responsible of the cylinder-by-cylinder variation. Indeed, the gasoline rail mounted on the examined engine at test bench represen rail, mounted on the commercial S ts anon-optimized prototype geometry. The optimized version of engine, does not exhibit differences in the gasoline mass flow rates between port injectors. All the previous considerations underline that rail geometry is main responsible for the experimental cylinder-by-cylinder variation analyzed in this research activity. Figure 2. Comparison of experimental in-cylinder pressure traces and burn rates profiles at 3000 rpm, two IMEP levels (9 and 13 bar) and EGR = 0% (a). Differences of the injected gasoline mass per stroke between port injectors (b).
Figure 3. Relative air-fuel ratio (A) in Cyl #1, Cyl #2 and overall engine at different operating points as a function of EGR rate. Figure 3 shows the relative air—fuel ratio (A) in Cyl #1, Cyl #2 and the engine exhaust at different engine o perating points and EGR rates. A of Cyl #1 was estimated from the composition of the exhaust gases of that cylinder through the carbon balance of species. Similarly, the overal This allows to indirectly evaluate the A value of Cyl #2 under the hypothesis o volumetric efficiency in the compression s age. Even though the engine conditions are close to stoic lean (A between 1.03 and 1.08) and rich (A between 0.96 and 0.98) mixtures are o Cyl #1 and Cyl #2, respectively. Considering that the engine does not provide an individual set of injection and combustion phasing for each cylinder, the same spark advance and injection parameters are chosen for both cylinders, optimizing the engine load. As a consequence, a difference in the single cylinder load is found at each operating condition. engine A was verified from the overall exhaust gas composition. f an equal for both cylinders, as suggested by the comparison of pressure traces hiometric, btained in
Figure 4. Indicated load in Cyl #1 and Cyl #2 at different engine operating points as a function of EGR rate.
Figure 5. Engine Indicated Specific Fuel Consumption at different operating points as a function of EGR rate.
Figure 6. MFB 10% in Cyl #1 and Cyl #2 at different operating conditions as a function of EGR rate. leaving a shorter period available to complete the combustion process. Further increase in combustion duration against the EGR is found at lower engine load, because of the greater impact of internal EGR which contributes to increase the overall in-cylinder residual content at decreasing the load.
Figure 7. MFB 10-50% in Cyl #1 and Cyl #2 at different operating conditions as a function of EGR rate.
Figure 9. Numerical/experimental comparisons of combustion core duration MFB 10-50% for Cyl #1 (a) and Cyl #2 (b) at various operating conditions.
Figure 10. Numerical/experimental comparisons of ISFC (a) and TIT (b) at various operating conditions
Figure 11. Numerical/experimental comparisons of nitrogen oxide (NO) for engine (a) and Cyl #1 (b) at various operating conditions. As for pollutants, Figures 11-13 show the comparison between numerical and experi- mental levels of regulated gaseous emissions: NO, CO and HC, respectively. Consistently with the presentation of the experimental results, the variations of considered pollutant emissions are reported as a function of the EGR rate.
Figure 12. Numerical/experimental comparisons of percent carbon monoxide (CO%) for both engine and Cyl #1 at various operating conditions.
Figure 13. Numerical/experimental comparisons of engine unburned HC at various operating conditions.
Figure 14 shows the ISFC comparisons in the analyzed engine operating points. Interestingly, this figure highlights that not negligible ISFC improvements are realized and the lower ISFC levels resulting from the balanced A/F ratio (blue bars in Figure 14) mostly follow the EGR related trend. A maximum ISFC percent benefit equal to 5% (AISFC = 12.4 g/kWh at 3000@7, EGR% = 9%) is attained, while the average percent gain is around 2%. Figure 14. ISFC comparisons: Unbalanced A/F ratio vs. Balanced Stoichiometric A/F ratio at varying the engine operating condition.
Figure 15. Comparisons of NO (a), CO (b) and HC (c): Unbalanced A/F ratio vs. Balanced Stoichiometric A/F ratio at varying the engine operating condition.
Table 1. Engine characteristics.
Table 2. Overall set of experimental engine points. The measured net IMEP (Table 2) is properly derived from the acquired in-cylinder pressure traces over the entire engine cycle and represents the difference between the gross IMEP (in-cylinder pressure over compression and expansion strokes) and the pumping mean effective pressure (PMEP), evaluated over the intake and exhaust strokes.
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Verbesserung des Systemverhaltens von trockenlaufenden und geschmierten Friktionssystemen durch Ingenieurkeramische Werkstoffe
by Asli Arslan

2023

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Ultra low emission vehicle development: a study on the drivers for car firms and the implications for sustainable development policy
by Marc Dijk

2022, International Journal of Automotive Technology and Management

This paper studies the factors that lead automotive manufacturing firms to invest in Ultra Low Emission Vehicles (ULEVs). ULEVs are vehicles that emit extremely low levels of harmful gases compared to other vehicles. In contrast to... more
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CNG direct injection spark-ignition engine with high turbulence and high compression ratio : numerical and experimental investigations
by Gilles Coma

2022

This study carried on within the H2020 GasOn project scope of work showed the interest to improve the turbulence of a CNG direct injection spark ignited engine with high compression by a dedicated cylinder head design. The in-cylinder air... more
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Berechnungsmethodik zur tribologischen Funktionsoptimierung im Kolben-Zylinder-System
by Gunter Knoll

2022, MTZ - Motortechnische Zeitschrift

Die Trends in der Motorenentwicklung wie Steigerung von Motoreneffizienz und spezifischer Leistung führen zu einer stärkeren Beanspruchung der Kurbeltriebbauteile. Insbesondere das Kolben-Zylinder-System ist davon betroffen und hat... more
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Wirkungsgradmessungen von Ölpumpen unter Einfluss des Luftgehalts
by Bernhard Kirsch

2022, MTZ - Motortechnische Zeitschrift

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Optimization of Pre-Chamber Geometry and Operating Parameters in a Turbulent Jet Ignition Engine
by Rudolf Tomic

2022, Energies

A turbulent jet ignition engine enables operation with lean mixtures, decreasing nitrogen oxide (NOX) emissions up to 92%, while the engine efficiency can be increased compared to conventional spark-ignition engines. The geometry of the... more
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Thermodynamic Optimization of Proton Exchange Membrane Fuel Cell System
by S. Obayopo

2022

The proton exchange membrane (PEM) fuel cell is a promising candidate as zeroemission alternative power source for transport and stationary applications due to its high efficiency, low-temperature operation, high power density,... more
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Control of the Air Supply Subsystem in a PEMFC with Balance of Plant Simulation
by Alan Josue Ribera Rojas

2022, Sustainability

This paper deals with the design of a control scheme for improving the air supply subsystem of a Proton Exchange Membrane Fuel Cell (PEMFC) with maximum power of 65 kW. The control scheme is evaluated in a plant simulator which... more
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Influence Of Engine Speed And Injection Timings On In-Cylinder Heat Transfer For Port Injection Hydrogen Fueled Engine
by Khalaf Hamada

2022

The overall heat transfer process within the in-cylinder for port injection hydrogen fueled internal combustion engine (H2ICE) has been investigated through simulation, while the experimental data has been utilized for validation purpose... more
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Emission Reduction Strategies for Small Single Cylinder Diesel Engine Using Valve Timing and Swirl Ratio
by S. Thipse

2022, International Journal of Engineering

Small diesel engines are widely used for commercial vehicle and passenger car applications due to their higher torque requirements, fuel economy, and better thermal efficiency. These engines are exposed to different operating and... more
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Instationäre Modellierung und Sensitivitäsanalyse regelbarer CO2-Axialkolbenverdichter
by Peterson Cavalcante

2022

Im Rahmen dieser Arbeit wird eine dynamische Modellierung eines regelbaren PKW-Axialkolbenverdichters mit Schwenkscheibenmechanismus für das Kältemittel CO2 durchgeführt, die den Verdichter in seiner Gesamtheit beschreibt. Das erstellte... more
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Unthrottled Engine Operation using Variable Valve Actuation: The Impact on the Flow Field, Mixing and Combustion
by jerome helie

2022, SAE Technical Paper Series

The effect on the intake flow field, air fuel mixing processes, thermodynamic performance and emissions output has been investigated for a range of valve operating profiles. A standard speed load point of 2000 rpm and 2.7 bar IMEP 720°... more
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Combustion and Emission Characteristics of an LNG Engine for Heat Pumps
by Kyoungyoul Lee

2022, Energies

The gas engine for a GHP is fueled by liquefied natural gas (LNG) and power from a natural gas engine-driven compressor, which operates the air conditioning and heating system via liquefaction and evaporation of refrigerant. Since gas... more
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Effect of Cylinder-by-Cylinder Variation on Performance and Gaseous Emissions of a PFI Spark Ignition Engine: Experimental and 1D Numerical Study
by Luigi Teodosio

2022, Applied Sciences

Combustion stability, engine efficiency and emissions in a multi-cylinder spark-ignition internal combustion engines can be improved through the advanced control and optimization of individual cylinder operation. In this work,... more
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Mixture preparation and combustion characteristics in a compression-ignition gasoline engine
by Steven Begg

2022

The recent interest in controlled auto-ignition (CAI) in gasoline engines has been focused on the potential for higher efficiency, lower fuel consumption and emissions of carbon dioxide (CO 2) and the oxides of nitrogen (NO x). Strategies... more
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Influence of initial charge conditions on engine performance and emission of a DISI hydrogen-fueled engine under various injection timings
by arash nemati

2021

Using hydrogen as an alternative fuel in spark-ignition engines has great potential for reducing exhaust emissions. Hydrogen engines have CO 2-and HC-free combustion and lean operation, offering lower nitrogen oxide (NO x) emission. In... more
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Influence of initial charge conditions on engine performance and emission of a DISI hydrogen-fueled engine under various injection timings
by Arash Nemati

2021, Turkish Journal of Engineering and Environmental Sciences

Using hydrogen as an alternative fuel in spark-ignition engines has great potential for reducing exhaust emissions. Hydrogen engines have CO2- and HC-free combustion and lean operation, offering lower nitrogen oxide (NOx) emission. In... more
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Untersuchung des Betriebsverhaltens eines Wasserstoff-Wankelmotors
by frank herfert

2021

Thesis by Heinz Stutzenberger 1983 180 pages Engine » Audi-NSU KKM 871, operating with hydrogen + gasoline-hydrogen mix > Content ... > Vorwort 3 > Inhaltsverzeichnis 4 > 1. Einleitung 7 > 2. Verbrennungsablauf + Motorverhalten des... more
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FL1-3: A Two-Zone Thermodynamic Model for Hydrogen-Fueled S.I. Engines(FL: Fuels and Lubricants,General Session Papers)
by Sebastian Verhelst

2021, The Proceedings of the International symposium on diagnostics and modeling of combustion in internal combustion engines

Because of significantly different physical and chemical properties ofhydrogen, compared to hydrocarbons, running an engine on hydrogen or modeling the eombustion of hydrogen in engines is not straightforward. This paper fbcuses on... more
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Recent progress in hydrogen-fueled internal combustion engines
by Sebastian Verhelst

2021, International Journal of Hydrogen Energy

Hydrogen-fueled internal combustion engines (H2ICEs) have been the topic of research for many decades, and contemporary reviews have surveyed the relevant literature. Because of a number of relatively large R&D projects that have been... more
Figure 3: Hybrid light duty truck equipped with a 4L 91kW hydrogen engine, demon- strated by Tokyo City University [3]. Permission to reproduce granted by copyright holder Tokyo City University.
Figure 4: Microbus equipped with a 4.7L 105kW hydrogen engine, demonstrated by Tokyo City University [3]. Permission to reproduce granted by copyright holder Tokyo City Uni- versity.
Figure 6: Final map of brake specific NOx emissions for the 2nd generation single cylinde1 engine in the DOE project [39]. Permission to reproduce requested.
Figure 7: CFD results showing the mixture stratification at TDC resulting from two different nozzle geometries. Local equivalence ratios shown on 2 vertical planes and 1 horizontal plane, for a 2000 rpm and 13.5 bar bmep engine condition [39]. Permission to reproduce requested.
Figure 8: CFD results showing the ‘NOx potency’ at TDC resulting from two differ- ent nozzle geometries. As a measure of ‘NOx’ potency, the volumes are shown having local equivalence ratios of 0.5jAj2. 2000 rpm and 13.5 bar bmep engine condition [39]. Permission to reproduce requested.
Figure 9: Schematic of final engine and aftertreatment system of the NTSEL project [24]. Permission to reproduce requested.
Figure 10: Control strategies for obtaining target power output while limiting NOx emis- sions [24]. Permission to reproduce requested.
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Ultra low emission vehicle development: a study on the drivers for car firms and the implications for sustainable development policy
by Marc Dijk

2021, International Journal of Automotive Technology and Management

This paper studies the factors that lead automotive manufacturing firms to invest in Ultra Low Emission Vehicles (ULEVs). ULEVs are vehicles that emit extremely low levels of harmful gases compared to other vehicles. In contrast to... more
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Numerical Study of Hydrogen Fuel Combustion in Compression Ignition Engine Under Argon-Oxygen Atmosphere
by Fadzli Ibrahim

2021, Jurnal Teknologi

Gas emissions from automobiles are one of the major causes of air pollution in our environment today. In fact, emissions of carbon dioxide (CO2), a product of complete combustion, has become a significant factor of the global warming... more
Figure 2 Pressure and temperature for simulation and theory results
Figure 3 Pressure and temperature for H2 direct injection in air atmosphere The simulation process continued with hydrogen fuel injected into a compression ignition engine at 702° CA to 712° CA. Figure 3 shows the pressure and temperature results from the simulation. Pressure increased and decreased proportionally with the decrease and increase of in-cylinder volumes. The same trend was also observed for temperature. The maximum pressure and temperature obtained was 847 K and 5.5 MPa, respectively. Figure 4 shows a very small value of heat release rate with a maximum value of 1.54 X 10°[J/CA]. It can be concluded that no combustion occurred throughout the process. The maximum temperature of 847 K (as shown in Figure 4) prevented the hydrogen fuel from igniting, because the auto-ignition temperature for hydrogen is 858K [14].
Figure 4 Temperature and heat release rate for He DI in air medium
Figure 5 Temperature for H2-Dl in Ar-O2compared to cold flow temperature in air and Ar-O2 H2 gas with higher flame speed and faster combustion velocity [18], induces the high premixed combustion phase and produces a high heat release rate.
Figure 7 shows the species mass of reactant and product during the simulation process from the beginning at 552°CA until the end at 858° CA. Major products for hydrogen combustion with argon-oxygen medium are oxygen (Oz) and water (H2O). Besides that, other species, such as O, OH, HO2 and H202, contribute only a very minimum mass value. It was observed that O2 remained high in volume and proved that the combustion occurred under lean combustion. Since argon gas reacts with no other gases during the simulation, it is proposed that this gas is recirculated back to the engine when designing and producing a concept engine in the future [7].
Figure 8 Comparison temperature image for DI H2 in Ar-O2 and air atmosphere simulation process, at 552°CA and 662°CA, only small differences in temperature values were obtained. Later, during the injection period at 702°CA, the values of temperature in argon-oxygen medium had already reached the hydrogen auto-ignition temperature; but a different situation occurred in the medium in normal air. At 792°CA, the temperature remained high in the argon-oxygen medium due the combustion process; however, the temperature in normal air had already decreased towards the ambient temperature. Detailed values for pressure and temperature between hydrogen direct injection in argon-oxygen and normal air medium are shown in Table 3.
Figure 7 Mass of combustion product
Table 2 Simulation parameters for region and initialization in case set-up
Table 1 Engine specifications for Yanmar TF70M
Table 3. Temperature and pressure at 552°CA, 632°CA 702°CA, and 792°CA References
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Numerical flow simulation of a natural gas engine equipped with and unscavanged auto-ignition prechamber
by Daniel Favrat

2021

A three-dimensional Reynolds-Averaged Navier-Stokes (RANS) simulation was carried out on an unscavanged auto-ignition prechamber-cylinder configuration in order to understand better the ignition conditions inside the engine. The... more
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Konzentrationsausgleich in Rührgefäßen bei laminarer Strömung
by Janine T. Bohlmann

2021, Forschung im Ingenieurwesen

Der in konventionellen Riihrgefiiflen ablaufende Mischprozefl ist hinsichtlich der physikalischen Detailvorgdnge noch weitgehend ungekldrt. Ursache dafiir ist vor allem das sehr komplizierte dreidimensionale Str6mungsfeld, das yon der... more
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Study of Performance and Environmental Emissions of Gasoline Spark Ignition Engine
by Mehrnoosh Dashti

2021

There are concerns both about the excess usage of limited fossil fuels reservoirs, and the environmental emissions production of fossil fuels such as CO, CO2, NO X , UHC and particles which may have long term effects on sustainable... more
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Enhancement of a Hydrogen Engine Cavitation Utilizing Mixed Fuel: a Review and Experimental Case Study
by Mohamed R . Gomaa

2020, International Review of Mechanical Engineering (I.RE.M.E.)

Previous studies examining gasoline and hydrogen engines have typically focused on factors pertaining to air to fuel ratios, their effects on an ICE performance, combustion by-products produced during the ignition cycle and its subsequent... more
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Advanced Rotary Engine Studies
by frank herfert

2020

by Charles Jones, 1980, Curtiss-Wright 26 pages > Introduction > Background Work > Testing of other Sizes 3 > Current NASA Advanced Engine Study - Approach + Status 5 » 1. Turbocharging 6 » 2. Increased IMEP + Speed /... more
RC2-75 AIRCRAFT ENGINE PROTOTYPE
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Recent progress in the use of hydrogen as a fuel for internal combustion engines
by Nipun Lodha

2018

Hydrogen-fueled internal combustion engines (H 2 ICEs) have been the topic of research for many decades, and contemporary reviews have surveyed the relevant literature. Because of a number of relatively large R&D projects that have been... more
Fig. 2 — Forklift truck equipped with a HJICE, part of the demonstration activities within the Hydrogen Region Flanders — South-Netherlands project [21].
Fig. 3 — Hybrid light duty truck equipped with a 4L 91 kW hydrogen engine, demonstrated by Tokyo City University [3]. Permission to reproduce granted by copyright holder Tokyo City University.
Fig. 5 — Final map of engine efficiency for the 2nd generation single cylinder engine in the DOE project [39]. Copyright © 2012 SAE International. Reprinted with permission.
walls is a prerequisite for increasing efficiencies of hydrogen- fueled engines. Measurements of instantaneous heat flux through the combustion chamber wall, for stoichiometric PFI operation on hydrogen, gasoline and methane have shown that the heat fluxes recorded for hydrogen were much higher than for gasoline or methane [69-71]. This explains why, although combustion is much faster, the resulting efficiency is only as high or even lower for hydrogen compared to gasoline or methane (see also [49]), due to the higher heat losses. The balance becomes a lot better if lean mixtures are used, to take advantage of the wide flammability limits of hydrogen. For moderately lean mixtures, flame speeds remain higher than for gasoline or methane, while heat losses decrease substan- tially, leading to increased efficiencies. If part load points are
Fig. 7 — CFD results showing the mixture stratification at TDC resulting from two different nozzle geometries. Local equivalence ratios shown on 2 vertical planes and 1 horizontal plane, for a 2000 rpm and 13.5 bar bmep engine condition [39]. Copyright © 2012 SAE International. Reprinted with permission. Testing of the lean NO, trap continued on the final engine chosen within the project, a four cylinder 4.7 L heavy duty engine converted to naturally aspirated DI hydrogen opera- tion [24]. The engine equipped with the aftertreatment system is shown schematically in Fig. 9 and was run on the JEO5 test cycle. The aftertreatment system resulted in large reductions of NO, emissions. However, the transient testing also showed Also, as mentioned in the previous section, heat transfer models need to be improved. While this would allow a better prediction of how heat transfer affects efficiency, it is of even
Fig. 8 — CFD results showing the ‘NO, potency’ at TDC resulting from two different nozzle geometries. As a measure of ‘NO,’ potency, the volumes are shown having local equivalence ratios of 0.5;4j2. 2000 rpm and 13.5 bar bmep engine condition [39]. Copyright © 2012 SAE International. Reprinted with permission.
Fig. 9 — Schematic of final engine and aftertreatment system of the NTSEL project [24]. Copyright © 2010 SAE International. Reprinted with permission. (simulated) turbocharged lean engine operating strategy was chosen. The equivalence ratio was fixed at ¢ = 0.3 (A = 3.3), with throttling at engine loads below 4 bar bmep and turbo- charging at higher loads. The engine data was recalculated to a 3 L engine powering a passenger car and used in a vehicle simulation on the Urban Dynamometer Driving Schedule (part of the FTP cycle). NO, emissions were calculated to be within the SULEV II range, i.e. well below the target set by DOE, without any aftertreatment. Further calculations showed that downsizing the engine could result in increased efficiency while still meeting the DOE target for NO, emissions, albeit with a smaller margin. This again illustrates the trade-off between efficiency and NO, emissions. The authors suggest advanced powertrain layouts (hybridization) could be looked at in future to take full advantage of the high efficiency and ultra-low emissions regimes. An alternative approach was taken within the DOE project [40]. Based on the obtained engine data, discussed above, a
Fig. 10 — Control strategies for obtaining target power output while limiting NO, emissions [24]. Copyright © 2010 SAE International. Reprinted with permission.
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Influence of initial charge conditions on engine performance and emission of a DISI hydrogen-fueled engine under various injection timings
by Shahram Khalilarya

2017

Using hydrogen as an alternative fuel in spark-ignition engines has great potential for reducing exhaust emissions. Hydrogen engines have CO 2 -and HC-free combustion and lean operation, offering lower nitrogen oxide (NO x) emission. In... more
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Application and cost–benefit analysis of solar hybrid power installation on merchant marine vessels
by George Papaioannou

2017, Ocean Engineering

Photovoltaic systems are renewable energy sources with various applications and their implementations in energy production and saving are verified. Installing those systems onto merchant marine vessels could prove to be an efficient way... more
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Ultra low emission vehicle development: a study on the drivers for car firms and the implications for sustainable development policy
by Carlos Montalvo

2016, International Journal of Automotive Technology and Management

This paper studies the factors that lead automotive manufacturing firms to invest in Ultra Low Emission Vehicles (ULEVs). ULEVs are vehicles that emit extremely low levels of harmful gases compared to other vehicles. In contrast to... more
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Ultra low emission vehicle development: a study on the drivers for car firms and the implications for sustainable development policy
by Carlos Montalvo

2016, International Journal of Automotive Technology and Management

This paper studies the factors that lead automotive manufacturing firms to invest in Ultra Low Emission Vehicles (ULEVs). ULEVs are vehicles that emit extremely low levels of harmful gases compared to other vehicles. In contrast to... more
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Impact of Retarded Spark Timing on Engine Combustion, Hydrocarbon Emissions, and Fast Catalyst Light-Off MAY 0 5 2005
by Info Meca Tronical

2016

An experimental study was performed to determine the effects of substantial spark retard on engine combustion, hydrocarbon (HC) emissions, feed gas enthalpy, and catalyst light-off. Engine experiments were conducted at cold engine... more
Figure 1.1. Cumulative engine-out and tailpipe-out HC emission over an FTP-75 for a six- cylinder SI engine.
igure 2.1 Intake charge motion control plate (CMCP) with 67% asymmetrical area reduction for the single cylinder engine. location of the spark retard ignition timing limit was extended.
ire 2.2 Fast-response HC analyzer probe sampling locations for single-cylinder experiments. Exhaust port and runner probes 7 cm and 37 cm from exhaust valves, respectively.
Figure 2.4 Overview of gas delivery system equipment used for quench gas and secondary air experiments.
Figure 2.5 Active three-way catalyst (serial no: XX901JP) specifications. Table 2.3. Multi-cylinder engine specifications.
Figure 2.6 Schematic of fast-response NDIR sampling head from Cambustion [8].
control plate (CMCP) installed in the intake port. a function of various spark timings and relative air/fuel ratios with and without the charge motiot
gure 3.3 Combustion durations as a function of spark timing for different relative air/fuel ratios and fluid temperatures without charge motion control plate (CMCP) at 3.0 bar Net-IMEP and 1500 RPM.
sure 3.4 Combustion duration as a function of spark timing for different relative air/fuel ratios and fluid temperatures with charge motion control plate (CMCP) at 3.0 bar Net-IMEP and 1500 RPM.
ire 3.5 Location of 50% MFB as a function of spark timing without charge motion control plate (CMCP) for various relative air/fuel ratios and fluid temperatures at 3.0 bar Net-IMEP and 1500 RPM.
igure 3.6 Location of 50% MFB as a function of spark timing with charge motion control plate (CMCP) for various relative air/fuel ratios and fluid temperatures at 3.0 bar Net-IMEP and 1500 RPM.
igure 3.7 Thermocouple measured port exit exhaust gas temperature as a function of 50% MEB location for various relative air/fuel ratios with and without the charge motion control plate (CMCP) at 3.0 bar Net-IMEP, 1500 RPM, and 20° C fluids.
gure 3.8 Converter-in sensible enthalpy rate as a function of COV of Net-IMEP for different relative air/fuel ratios with and without the charge motion control plate (CMCP).
‘igure 3.10 Steady-state HC flow rate and emission index HC emissions as a function of 50% MEB location for various relative air/fuel ratios and fluid temperatures at 3.0 bar Net-IMEP, 1500 RPM, with charge motion (CMCP). HC emissions.
‘igure 3.11 Typical time-resolved HC concentration profile measured in the exhaust port, 7 cm from the exhaust valves. Major features shown and noted during gas exchange, from exhaust valve opening (EVO) to exhaust valve closing (EVC).
Figure 3.12 In-cylinder pressure and exhaust port and runner time-resolved HC concentration measurements for three different spark timings (Sp = 15°, 0°, -16° BTDC) at 3.0 bar Net-IMEP, A = 1.0, CMCP, and 20° C fluids.
gure 3.13 In-cylinder pressure and time-resolve HC concentration measured in the port and runner for 90°, 40°, 20° C fluid temperatures, spark timing = -1° BTDC, A = 1.0, and without charge motion (CMCP). combined with the plug flow model, yielding a HC mass flow rate. sonditions that were required by the plug flow model. Data from the fast-response FID was the
Figure 3.15Engine cycle simulation thermodynamic model of open system, including crevices, work transfer to piston, and heat transfer to cylinder boundaries. temperatures.
manuals. Figure 3.16 GT-Power model of single-cylinder engine with intake and exhaust system.
‘igure 3.17 Port exit exhaust mass flow rate and exhaust gas temperature predicted by the engine model at the exit of the exhaust valves versus crank angle for various spark timings at 3.0 bar Net-IMEP, 1500 RPM, A = 1.0, and 20° C fluids.
Figure 3.18 Spatial and cyclic variations of the time-resolved HC measurements in the exhaust port, 7 cm from exhaust valves at several port-exit locations.
ire 3.19 Model results from analysis of HC measurements. GT-Power simulation results of in-cylinder pressure and exhaust mass flow rate at the port exit and runner. Measured HC concentration and resulting HC mass flow rate computed by a plug flow model.
Figure 3.20 Steady-state HC emissions as a function of spark timings measured in three locations under A = 1.0, 20° C fluids.
Figure 3.22 Steady-state HC emissions as a function of spark timing for three different locations under A = 1.1 and 20° C fluids.
Figure 3.23 Emission index HC emissions as a function of the location of 50% MFB for three different locations under A = 1.1 and 20° C fluid.
Figure 3.24 Phasing of quench gas injection as a function of crank angle. Data shown with in- cylinder pressure and solenoid trigger signal for 3.0 bar Net-IMEP, 15° BTDC spark timing, and A = 1.0.
Figure 3.25 Change in HC and NO, emissions as a function of the ratio of mass of quench gas to mass of charge (mCO2Quench/mCharge,j-+fye;) for three spark timings with A = 1.0 and 20° C fluids.
Figure 3.26 Steady-state HC emissions as a function of spark timing at four locations (cylinder exit, port exit, runner, and mixing tank) under A = 1.0 and 20° C fluids.
Figure 3.27 Emission index HC emissions as a function of 50% MFB location at four locations under A = 1.0 and 20° C fluids.
Figure 3.28 Steady-state HC emissions as a function of spark timings at four locations unde A= 1.1 and 20° C fluids.
Figure 3.29 Emission index HC emissions as a function of 50% MFB location at four locations for A = 1.1 and 20° C fluids.
re 3.30 Exhaust secondary air injection experiments conducted at an engine relative air/ fuel ratio of 0.85 (engine) aS a function of with exhaust relative air/fuel ratio. Data shown for 3.0 bar’ Net-IMEP, 1500 RPM, 0° BTDC spark timing, and 20° C fluids.
ure 3.31 Hydrocarbon flow rate and runner exhaust gas temperatures as a function of various secondary air injection timings for a fixed engine (Agngine = 0.85) and exhaust stoichiometry (Acxhayst = 1.2). Data shown for 3.0 bar Net-IMEP, 1500 RPM, and 20° C fluids.
sure 4.1 Baseline calibration MAP, RPM, spark timing, and relative air/fuel ratio as a function of time after crank. Data from an ambient startup under idle load and speed conditions.
re 4.2 Spark timing as a function of time after startup for various spark modifications (A® spark): aggressive retarded spark timing modification, A@,,, = -15°
Figure 4.3 Measured UEGO relative air/fuel ratio (A) as a function of cycle no. after startup. Data shown for accelerated closed loop control compared to baseline calibration.
yure 4.4 COV of Net-IMEP as a function modified spark timing, 4 to 20 seconds after crank (130 cycles). In-cylinder pressure data acquired from all four cylinders at idle speed and load conditions following an ambient start.
ire 4.5 Location of 50% MFB as a function of cycle number after 20° C startup. Data shown for various spark timing modifications under idle speed and load conditions.
ire 4.6 Combustion duration (10-90%) as a function of cycle number after 20° C startup. Data shown for various spark timing modifications under idle speed and load conditions.
gure 4.7 Cylinder no. 4 port exit and converter-in hydrocarbon concentrations (ppmC 1) for the first 20 seconds following a 20° C start. Baseline timing and fueling calibration under idle load and speed conditions.
gure 4.8 Cylinder no. 4 pressure, CO, CO, and HC time-resolved concentrations measured at the exhaust port exit of cylinder no. 4 as a function of crank angle (CA). Data shown for three different spark modifications (absolute spark timings: 3°, -1°, -6° BTDC) from cycle 200 to 210, stoichiometric closed loop control and idle speed and load conditions (2.3 bar Net-IMEP, 1000 RPM, A= 1.0).Grayed areas show period of exhaust flow. flow.
igure 4.10 Cumulative, 1 to 20 seconds, ambient startup CO and HC mass emissions measured at the exhaust port exit of cylinder no. 4 and converter-inlet as a function of modified spark timing under idle speed and load conditions (Port exit emissions are scaled by a factor of 4 for comparison to converter-in levels).
concentrations for cylinders no. 3 and no. 4 for a spark timing of 3° BTDC under idle speed and
gure 4.12 Cylinder no. 3 and no.4 in-cylinder pressure and port exit time-resolved HC concentrations as a function of crank angle (CA). Data shown from cycle 200 to 210 after startup, absolute spark timing -4° BTDC, stoichiometric closed loop control, and idle speed and load conditions (2.3 bar Net-IMEP, 1000 RPM, A= 1.0).Grayed areas show period of exhaust flow.
igure 4.13 Port exit measured exhaust gas temperatures and relative air/fuel ratios (A) calculated from time-resolved CO and CO, measurements for cylinders no. 3 and no. 4 as a function of modified spark timing. Data shown for idle load and speed conditions under closed loop stoichiometric control.
Figure 4.14 Spark-ignition engine exhaust gas CO, and CO concentration data as a function of relative air/fuel ratio.
re 4.15 Converter-in CO + O, emissions and emission index HC levels as a function of intake MAP. Data shown for stoichiometric closed loop control and 20° C idle speed and load conditions (2.3 bar Net-IMEP, 1000 RPM, A = 1.0). CO+O, concentration shown with and without oxygen concentration corrections due to HC emissions.
‘igure 4.16 Measured exhaust gas and component temperatures as a function of time after crank, Data shown for baseline timing and fueling calibration under idle load and speed conditions. metal skin and component temperatures were observed to vary less than 20° C.
Figure 4.17 Cumulative converter-in feed gas HC emissions as a function of cumulative converter-in sensible enthalpy. Data shown for 1 to 20 seconds following an ambient startup for four spark timing modifications (A@,,,) under idle speed and load conditions.
igure 4.18 Cumulative HC mass emissions prior to catalyst light-off as a function of light- off time following an ambient start. Data shown for various spark modifications (A6,,,) under idle speed and load conditions (Light-off defined as 50% reduction in HC emissions).
gure 5.1 Plug flow exhaust oxidation sub-model flowchart with linked engine predicted cylinder-exit conditions, heat transfer, and chemical kinetic mechanism. each zone had a specific exhaust geometry and heat transfer correlation, Fig. 5.2.
Figure 5.2 Diagram of exhaust port and runner zones.
3 Simulation of single-cylinder exhaust mass flow versus crank angle after exhaust valve opening. Exhaust mass modeled using two elements; mass from compressible blown process and mass from incompressible displacement.
5.2.2 INITIAL CONDITIONS
figure 5.6 Mole fraction of exhaust gas composition for highly diluted reactive mixture of hydrocarbons.
7 Thermocouple measured exhaust gas and component temperatures for 15°,0°, and -15° BTDC spark timings. Steady-state warmed-up exhaust system with 20° C fluids, 3.0 bar Net-IMEP, 1500 RPM, and A = 1.0.
ple 5.2 Nusselt number correlations for exhaust runner with argument factors pulses and pipe bends [26, 27]. 5.3.3 AVERAGE EXHAUST HEAT TRANSFER COEFFICIENT
5.8 Average port (left) and runner (right) heat transfer coefficient as a function of spark timings. Data shown for mass elements from exhaust blowdown and displacement process.
5.9 Schematic of exhaust port and runner temperature validation of heat transfer sub- model. Exhaust port and runner HC oxidation modeled as a heat addition per unit length. Thermocouple modeled exhaust gas temperature at port exit and runner.
‘igure 5.10 Instantaneous exhaust gas, runner wall and thermocouple temperature as a function of time. Thermocouple response data shown for hot exhaust component temperatures.
‘igure 5.11 Thermocouple exhaust gas temperatures, measured versus modeled, for various spark timings. Comparison includes heat released due to exhaust system HC burn-up.
Figure 5.12 Cold-start SI engine-out speciated HC emissions for initial 63 seconds of engine operation using Japanese domestic fuel. Source: Yamamoto et al. [16].
Figure 5.13 Variation in engine-out hydrocarbon composition as a function of time durin; warm-up. Source: Kubo et al. [17].
Figure 5.14 Hydrocarbon oxidation pathways.
ure 6.1 Relative port, runner, and total exhaust system HC oxidation as a function of spark timing and relative air/fuel ratio. Data shown for 3.0 bar Net-IMEP, 1500 RPM, 20° C fluids, with intake charge motion (CMCP).
ire 6.2 Measured port and runner exhaust gas temperature as a function of spark timing and relative air/fuel ratio. Data shown for 3.0 bar Net-IMEP, 1500 RPM, 20° C fluids, with intake charge motion (CMCP).
ure 6.3 Measured and predicted fraction of port reacted hydrocarbons (left) and exhaust port time averaged and enthalpic gas temperatures (right) as a function of spark timing. Data shown for A = 1.0, 3.0 bar Net-IMEP, and 1500 RPM.
sure 6.4 Measured and predicted fraction of port reacted hydrocarbons (left) and exhaust port time averaged and enthalpic gas temperatures (right) as a function of spark timing. Data shown for 4 = 1.1, 3.0 bar Net-IMEP, and 1500 RPM.
ure 6.5 Measured and predicted fraction of runner reacted hydrocarbons (left) and exhaust runner time averaged and enthalpic gas temperatures (right) as a function of spark timing. Data shown for 4 = 1.0, 3.0 bar Net-IMEP, and 1500 RPM.
Figure 6.6 Measured and predicted fraction of runner reacted hydrocarbons (left) and exhaust runner time averaged and enthalpic gas temperatures (right) as a function of spark timing. Data shown for A = 1.1, 3.0 bar Net-IMEP, and 1500 RPM.
K showed modest burn up sensitivity (less than 10%) to temperature variations
Figure 6.10 Hydrocarbon mole fraction composition for three cases. Data shown for case 1 with the baseline HC mixture, case 2 is 100% iso-octane, and case 3 contains 55% ethylene, 35% pentane, and 10% methane. whereas, n-alkanes fragment into ethylene which reacts very rapidly to form highly reactive
igure 6.11 Calculated half-lives (t59) of various hydrocarbon mixtures as a function of temperature for exhaust residence times. Data shown for A = 1.0, 1% HC gas mixture.
CO, by 15 - 35%.
Figure 6.14 CO oxidation time constant as a function of exhaust relative air/fuel ratio (A) Data shown for three different spark timings (exhaust gas temperatures).
Figure 6.15 Measured time averaged and predicted adiabatic enthalpic runner gas temperatures as a function of exhaust relative air/fuel ratio (A).
Figure 6.16 Cylinder no. 4 port exit and collector-in exhaust mass flow rate as a function of crank angle. mixed, and subjected to a step change in air/fuel ratio at the runner collector before reaching the
Figure 6.17 Cylinder no. 4 port exit and collector-in exhaust gas temperature as a function of crank angle.
Figure 6.18 Diagram of multi-cylinder exhaust system with runner collector mixing.
Figure 6.19 Cylinder relative air/fuel ratio as a function of distance from exhaust valve seats.
‘igure 6.20 Measured and predicted fraction of runner reacted hydrocarbons (left) and exhaust runner time averaged and enthalpic gas temperatures (right) as a function of spark timing. Data shown for idle speed and load conditions, 1 to 20 seconds after 20° C startup. Model initial conditions based upon air/fuel ratio data from cylinder nos. 3 and 4.
between the two fired cylinders. cycle position is determined from spark discharge times differences
ure C.4.1 In-cylinder pressure and exhaust port and runner time-resolved HC concentration measurements for three different spark timings (Sp = 15°, -1°, -16° BTDC) at 3.0 bar Net-IMEP, A = 1.0, CMCP, and 90° C fluids.
sure C.5.1 Model results from analysis of time-resolved HC measurements. GT-Power simulation results of in-cylinder pressure (upper left) and exhaust mass flow rate at the port exit and runner.(upper right). Measured HC concentrations (lower left) and predicted HC mass flow rate (lower right). Data shown for 3.0 bar Net-IMEP, 1500 RPM., 15° BTDC spark timing, A = 1.0, and 20° C fluids.
igure C.7.2 Port exit and runner EGTs and tailpipe-out HC flow rate as a function of exhaust relative air/fuel ratio. Data shown for continuos secondary air injection at a fixed engine relative air/fuel ratio (Aengine = 9-85) and 0° BTDC spark timing.
zure D.1.1 In-cylinder pressure, RPM, MAP, spark timing, and lambda (A) as a function of time after ambient startup. Data shown for baseline spark and fuel calibration (A®.park = 0°). Calculated Net-IMEP and COV of Net-IMEP, averaged RPM, MAP, spark timing, and A, for 1 to 20 seconds after crank. ).1.1 In-cylinder Pressure, RPM, MAP, Spark Timing, and Relative Air/Fuel Ratic
ure D.3.1 In-cylinder pressure, RPM, MAP, spark timing, and lambda (A) as a function of time after ambient startup. Data shown for -10° spark timing modification (A®.park = -10°). Calculated Net-IMEP and COV of Net-IMEP, averaged RPM, MAP, spark timing, and A, for 1 to 20 seconds after crank. D.3.1 In-cylinder Pressure, RPM, MAP, Spark Timing, and Relative Air/Fuel Ratio
re D.3.2 Time-resolved port exit (FFID #1) and converter-in (FFID #2) HC concentrations, port exhaust gas temperature (EGT), and exhaust component temperatures as a function of time after ambient startup. Data shown -10° spark modification (A®spar, = -10°). Calculated sensible enthalpy and averaged HC concentrations for 1 to 20 seconds after crank.
ure D.4.1 In-cylinder pressure, RPM, MAP, spark timing, and lambda (A) as a function of time after ambient startup. Data shown for -15° spark timing modification (A® spark = -15°). Calculated Net-IMEP and COV of Net-IMEP, averaged RPM, MAP, spark timing, and A, for 1 to 20 seconds after crank.
Figure F.1.1| Exhaust gas hydrogen (H3) concentration as a function of CO concentration [14].
Port, runner, and calculated burn-up exhaust gas temperature as a function of combustion phasing (location of 50% MFB). Burn-up temperatures are based upon complete oxidation of HC without the effects of heat transfer in the runner. Data shown for 3.0 bar Net-IMEP, 1500 RPM, 20° fluids, 4 = 1.0 with charge motion (CMCP).
F.4.1 Spark Timing = 15° BTDC and A = 1.0 re F.4.1_ Time histories of cylinder-exit temperature, velocity, mass flow, and hydrocarbon mass emissions as a function of crank angle. Data shown for A = 1.0 and spark timing = 15° BTDC.
Figure F.4.2_ Mass element temperature and exhaust gas mole fraction histories as a function of time for a highly diluted reactive mixture. Data shown for A = 1.0 and spark timing = 15° BTDC.
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Application of genetic algorithm to both sides pressure optimization of PEMFC
by Hamid H Nemati and 
1 more
Mohammad T Shakeri

2016, 2009 IEEE 13th International Multitopic Conference

Genetic algorithm is very powerful tool at different optimization domain. It's known very well that pressure is one of the most important and effective parameter in fuel cell Performance. In this paper, twovariable GA was applied to... more
a@ is net water transport coefficient across the
PARAMETERS VALUE THAT PRODUCE MAXIMUM POWER Fig. 3. Maximum fitness during the generations. TABLE 4
Fig. 4. Polarization curves for different both sides equal pressure.
‘VALUES OF GENETIC ALGORITHM PARAMETERS TABLE 1
maximum power density, in other word, goal is to find the best value of anode and cathode pressure that have highest output power. We modeled the PEMFC using analytical method and used as fitness function. find the best value of anode and cathode pressure that ’ is molar flux between the channel and MEA. Also,
GEOMETRICAL AND FLOW RATE PAREMETERS that are shown in Table 4. Fig. 3 illustrates the convergence of the genetic code towards the best solution.
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Effects of the Stroke/Bore Ratio on the Performance Parameters of a Dual-Spark-Ignition (DSI) Engine †
by I. Sezer

2016, Energy & Fuels

In this study, the effects of the stroke/bore (S/B) ratio on the performance parameters of an engine having dual-spark ignition (DSI) were investigated theoretically by using a quasi-dimensional thermodynamic cycle model. A range of S/B... more
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The stratified combustion model of direct-injection spark-ignition engines
by Hoimyung Choi

2016, Proceedings of the Combustion Institute

This paper presents a new model for stratified combustion in direct-injection spark-ignition (DISI) engines, which considers the secondary diffusion flame in the burned region and the effect of mixturefraction fluctuation on the laminar... more
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