Crank Angle

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(THD) lubrication and

(THD) lubrication and

By the time piston reaches at the end of compression stroke that is, at 360 degree crank angle there is clear shift of positive hydrodynamic pressures to the skirt bottom surface such that these rise from mid point to the skirt bottom surface, the slope is steep but higher peak pressure values are close to the centre point of skirt bottom surface (Refer to Fig:6d). In the power stroke peak hydrodynamic pressures keep generating at the piston skirt bottom surface throughout the entire length of stroke showing less gentle slopes of instantaneous pressure fields. This shift is understandable as there is drastic directional shift of piston secondary motion coupled with exponential rise and subsequent fall of combustion gas pressures along with simultaneous increase in swept volume.
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Accelerating the Simulation of Chemically Reacting Turbulent Flows via Machine Learning Techniques.

Accelerating the Simulation of Chemically Reacting Turbulent Flows via Machine Learning Techniques.

175 Figure 4.38 Contour plots of temperature (left column) and OH mass fraction (right.. column) at varying crank angle degrees along the plane of the spray..[r]

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Bayesian models for the determination of resonant frequencies in a DI diesel engine

Bayesian models for the determination of resonant frequencies in a DI diesel engine

– from here out referred to as the resonant frequency. This model does not have a large effect on the deviance but, from a physics perspective, it is an important parameter – particularly from the aspect of what is desired from the model. It also stops the model from under-predicting the resonant fre- quency under the assumption that it is a stationary frequency. The minimal gain in the DIC, 9360, can be attributed to the very small decay in the res- onant frequency, therefore providing evidence that there is minimal change in the resonant frequency as the crank-angle increases. Updated posterior expectations of our parameters are now 6107 Hz for ω 0 and 2.1962 × 10 −10
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Modeling and Analysis of Cranckshaft

Modeling and Analysis of Cranckshaft

The pressure versus crank angle of this specific engine was not available, so the pressure versus volume (thermodynamic engine cycle) diagram of an ideal cycle similar engine was considered. This diagram was scaled between the minimum and maximum of pressure and volume of the engine. The four link mechanism was then solved by Excel programming to obtain the volume of the cylinder as a function of the crank angle. Figure 2 shows the scaled graph of pressure versus volume for this specific engine,

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Experiment Investigation of Combustion Characteristics Analysis of Blend of Diesel with Low Concentration Linseed Methyl Ester at Different Compression Ratio on VCR Diesel Engine

Experiment Investigation of Combustion Characteristics Analysis of Blend of Diesel with Low Concentration Linseed Methyl Ester at Different Compression Ratio on VCR Diesel Engine

Fig. 4 Curve showing varition of Heat Release v/s CA The curve showing variation of heat release and crank angle Due to increase in viscosity of fuels which are derived from edible or non edible vegetable oils it results in poor atomization and slower mixing which ultimately results in the longer ignition delay and the effect on ignition delay affect the heat release rate of the fuel. It can be observed from graph that the maximum heat release obtained for pure diesel is 27.16 C at C 18 and for corresponding maximum heat release rate obtained for LME10 at C 18 are 25.19 C .
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PERFORMANCE OF FOUR STROKE DIESEL ENGINE, FOCUSING COMBUSTION MODELING AND CYCLE ANALYSIS

PERFORMANCE OF FOUR STROKE DIESEL ENGINE, FOCUSING COMBUSTION MODELING AND CYCLE ANALYSIS

These Figures are made by mat lab programming with the use of modeling equations and theory. From these graphs can find pressure and volume from the cycle analysis at each crank angle and can predict the performance of engine combustion from this graph without using experimental work and time consuming, laborious work during experiment.

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Experimental Investigation of Performance and Emission Characteristics of Diesel Engine Working on Diesel with Neem Biodiesel and Ethanol Blend with EGR

Experimental Investigation of Performance and Emission Characteristics of Diesel Engine Working on Diesel with Neem Biodiesel and Ethanol Blend with EGR

The variation of cylinder pressure with respect to crank angle for neem biodiesel blends of B20% and B40% at constant pressure 180 bar is shown in figures 26 and 27. From the figures it is observed that the peak pressures variation is less because the properties such as calorific value, viscosity, and density are brought closer to diesel. Cylinder pressure of blends B20% and B40% with ethanol 5% without EGR at full load. Peak pressure was found to be 70.94 bar and 69.35 bar respectively. Cylinder pressure of blends B20% and B40% with ethanol 5% with EGR 5% the values found to be 70.31 bar and 69.43 bar respectively. Cylinder pressure of blends B20% and B40% with ethanol 5% with EGR 10% at full load the values found to be 68.94 bar and 69.88 bar respectively. The diesel cylinder pressure at full load found to be 67.05 bar. The cylinder pressure of blend B20% with ethanol 5% without EGR value found to be high, this is due to good mixture formation for bio-diesel at higher loads where temperatures are high, EGR serves as a heat absorbing agent, which reduces the cylinder charge temperature in the combustion chamber during the combustion process.
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Application of back propagation neural network in predicting nox  emission from i. c. engines

Application of back propagation neural network in predicting nox emission from i. c. engines

The entire experiment was carried out at the I.C Engine laboratory in a computerized single cylinder, four stroke, multi-fuel, variable compression ratio (VCR) engine as shown in Fig.1. The fuel used for the experiment was diesel. The setup consists of single cylinder, four stroke, multi-fuel, research engine (specified in Table 1) connected to eddy current type dynamometer for loading. The operation mode of the engine can be changed from Diesel to Petrol or from Petrol to Diesel with some necessary changes. In both the modes, the compression ratio can be varied without stopping the engine and without altering the combustion chamber geometry by specially designed tilting cylinder block arrangement. The injection point and spark point can be changed for research tests. Setup is provided with necessary instruments for measuring combustion pressure, diesel line pressure and crank-angle. These signals are interfaced with
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Simulated work loops predict maximal human cycling power

Simulated work loops predict maximal human cycling power

Joint power derived from human cycling experimental data Previously reported kinematics ( joint angle and angular velocity) and kinetics (net joint moment and power) during maximal cycling (Martin and Brown, 2009) were used for this investigation. To briefly summarize the experimental study, 13 highly trained cyclists (1 female, 12 males, mean±s.d. mass 74.8±6.5 kg) performed maximal isokinetic cycling trials at 120 rev min −1 for one 30 s trial. For this investigation, we used only data from the first complete cycle for each subject, which represents a non-fatigued state at a constant cycling velocity. During each trial, pedal reaction force, pedal and crank angle, and limb segment position were recorded at 240 Hz. Specifically, pedal reaction force was recorded from the right pedal using two 3-component piezoelectric force transducers (Kistler 9251, Kistler USA, Amherst, NY, USA). Pedal and crank angle were recorded using digital encoders (S5S-1024-IB, US Digital, Vancouver, WA, USA) attached to the right pedal and crank. Limb segment position, defined as the position of the hip, knee, ankle and fifth metatarsal head, was derived from measurements from an instrumented spatial linkage system (Martin et al., 2007). The limb segment positions were used to calculate ankle, knee and hip joint angle and joint angular velocity (termed ‘ experimental joint angle ’ and ‘ experimental joint angular velocity ’ in Fig. 1). Parasagittal plane net joint moment values (termed ‘ experimental joint moment ’ in Fig. 1) at the ankle, knee and hip were determined using inverse dynamic techniques (Elftman,
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Performance, Combustion & Emission Analysis of the Single Cylinder CI Engine using Karanja Oil Methyl Ester Blends

Performance, Combustion & Emission Analysis of the Single Cylinder CI Engine using Karanja Oil Methyl Ester Blends

The Fig-11 shows the variation of heat release rate with crank angle at full load for the pure diesel & the KOME blends. From the above Fig, the rate of Net Heat Release for the pure diesel (B0) is more & reduces for the KOME biodiesel blends. This may be due to non-viscous and better atomization property of diesel fuel (higher hydrocarbon fraction for the diesel fuel) as compared to the B10 & B20 [18]. The Peak Heat release values are 48.62 at 358°, 49.23 at 360° & 46.52 J/Degree at 361° for Diesel fuel, B10 & B20 respectively. The peak value is Maximum for B10 with respect to the diesel fuel. This is due increased delay period of fuel during the relatively longer delay period caused in higher rate of heat release. On the other hand, the peak heat release rate is lower for B20with respect to B10 and pure diesel. This may be due to the lower volatile, high temperature flash point and higher viscosity of KOME blends, that leads to a reduction in fuel-air mixing ratio, causing in smaller content of fuel is ready for premixed combustion stage after ignition delay [18].It can also be witnessed that the diffusion burning indicated by the area under second peak is dominant for B20. This may be due to the complete combustion occurs at the late combustion phase. Hence for the B20 Heat release is increased at End of the Power Stroke.
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Cylinder Pressure Variation Modeling inside a Diesel Engine

Cylinder Pressure Variation Modeling inside a Diesel Engine

There is no measurement available to validate the ability of the engine model to describe a diesel engine. J. Scarpati et al. in 2007, however used the same model to describe the cylinder pressure in straight 6-cylinder diesel engine, [5]. the comparison between the model output and measured cylinder pressure is done for three different engine speeds. There are different approaches and ways of modeling ranging from easy to compound method and may vary in both structure and correctness. Several approaches of recording engine data are possible. In conventional applications data is logged with a fixed acquisition rate, whereby the time interval between two following recordings is fixed. However, because an engine runs in a cycle dictated by a set of mechanical mechanisms slider-crank, poppet valves, etc. – and because these mechanisms have fundamental consequences to how combustion takes place, it is necessary to record data at known crank angle intervals.
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Performance and Emission Characteristics of Diesel Engine with Exhaust Gas Recirculation (EGR) on Diesel and Neem Biodiesel Blends with Ethanol as Additive

Performance and Emission Characteristics of Diesel Engine with Exhaust Gas Recirculation (EGR) on Diesel and Neem Biodiesel Blends with Ethanol as Additive

The variation of Cumulative Heat Release Rate with respect to crank angle for diesel and different blends B20% and B40% neem biodiesel at constant pressure of 180 bar is as shown in figures 24-25, At full load condition without EGR cumulative heat release rate of 20% and 40% blends with 5% ethanol values found to be 1.54 kj and 1.49 kj respectively. The blends B20% and B40% with ethanol 5% with EGR 5% the values found to be 1.51 kj and 1.5 kj respectively. The blends B20% and B40% with ethanol 5% with EGR 10% the values found to be 1.53 kj and 1.46 kj respectively. The pure diesel value found to be 1.06 kj. It was observed from the figure that there is an increase in the ignition delay for the blends. From figure Peak cumulative heat release rate was found to be 1.54 kJ for without EGR B20% with 5% ethanol.
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CFD Simulation of In Cylinder Gases of Multi cylinder Diesel Engine for Estimation of Liner Temperature from Gas Side

CFD Simulation of In Cylinder Gases of Multi cylinder Diesel Engine for Estimation of Liner Temperature from Gas Side

Various Models are used to account for the fuel spray,toxic waste generation, and combustion. For direct injection engines, the fuel drizzle from the tip of the nozzle injector is introduced at the fix crank angle and time using a spray exemplary. For port fueled engines, it is considered that the combustion charge is well mixed, uniform and homogenyeous. A chemical mechanism narrating the reaction of vapor fuel with air is used to detailise the combustion, and models for turbulence-chemistry synergy are specified. Sub-models for NOx and soot formation are taken into consideration to estimate pollutant formation, which can be conjoined with the combustion calculation or calculated as a post processing step. Within cylinder combustion, the main challenge comes when we deal with physics for spray modeling and combustion. The spray is composed of a column of liquid getting in the domain at great speed which eventually split into droplets due to aerodynamic forces. These droplets break into smaller droplets or can even combine into larger droplets, all while exchanging mass with the surrounding gases. Sub-models are used for estimation of coalescence and breakup, heat and mass transfer computation, for capturing spray dynamics. The CFD mesh has to be adequately resolved to seize the connection among/admist the liquid droplets and the gases in the cylinder exactly. If fuel splash hits on the cylinder walls, it may form a fragile liquid film which undergoes its own processes of motion and vaporization and requires a separate processing.
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Design of Alpha Stirling Engine in Conjunction with Solar Concentrator

Design of Alpha Stirling Engine in Conjunction with Solar Concentrator

Chart -3: Regenerator Effectiveness Vs. Heat Supplied. IV. Crank Angle Vs. Volume Variation Volume variation of expansion Cylinder, compression cylinder and the total volume has been plotted with respect to Crank Angle. It has been found that the volume shows sinusoidal variation with respect to crank angle. The graph of Crank angle Vs. Volume variation is shown below.

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Non-Newtonian Effects on Lubricant Hydrodynamic Thin Film Flows in the Initial Engine Start Up

Non-Newtonian Effects on Lubricant Hydrodynamic Thin Film Flows in the Initial Engine Start Up

skirts and cylinder wall is determined as shown in figure 4.The intake stroke has negligible film thickness. This time period becomes very critical. This corresponds to the piston eccentricities towards minor thrust side. In the compression stroke, oil film thickness rises and attains its vertex near the end of that stroke of piston. This means that more film thickness value increases, more are the chances that piston eccentricities will get reduced. Hence, even lesser chances for any solid-to-solid contacts to take place. In the piston expansion stroke, maximum value of combustion gas force due to ignition (firing) is at 372 degree crank rotation. There is corresponding drop in the values of maximum and minimum film thickness. These values reach the minimum at the end of expansion stroke, which means that by that time full impact of combustion gas force on hydrodynamic pressures as well as film thickness has fully materialized. In the exhaust stroke, oil film thickness starts increasing again until it attains significantly high values at the end of exhaust stroke which corresponds to the 720 degree crank angle.
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Effect of Heat Transfer on the Output Performance Characteristics of an Air-Standard Otto Cycle

Effect of Heat Transfer on the Output Performance Characteristics of an Air-Standard Otto Cycle

In order to study the effect of temperature dependent specific heats, Fig. 5.4is presented. It shows variation of gas temperature versus crank angle using variable and constant-average specific heats running at engine speed of 5000 rpm and in- cylinder air–fuel ratio of 15. It is obvious that there is some difference when temperature dependent specific heat is used instead of constant specific heat. Although they have similar trends, the maximum temperatures with constant specific heats are significantly over-estimated in comparison with results obtained with variable specific heats.
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1000 particle at 120° CA ATDC. For the squish region soot particle number started at around 500 particle and dropped significantly to 120 particles at 120° CA ATDC. Fig. 12 explains the soot particle average size in the squish region and in the entire cylinder. At earlier combustion process in the squish region, the soot oxidation rate increases rapidly after 8° CA ATDC [4] to overcome the surface growth of soot. Soot particle average size in the squish region slowly decreased from 25 nm at 8 CA ATDC to 15 nm at exhaust valve opening (EVO). Soot particle size in the whole cylinder displayed a different result, where the soot particle average size at inlet valve closing (IVC) was 25 nm and increase to 40 nm at 30° CA ATDC. Beyond that the soot average size starts to decrease to 30 nm at EVO. The oxidation process started to dominate the overall soot formation process at higher crank angle, namely 30º CA ATDC, thus reduced the overall soot intensity, size and particles.
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Post-ignition knock occurred at various engine test conditions, with all the fuels tested. The knocking cycles were only observed in the fast and medium cycles and none occurred in the slow cycles. Knocking fast cycles, had the crank angles at which knock onset and peak pressure occurred, closer to the Top Dead Center (TDC) compared to that of the knocking medium cycles. The observations made, show that knock in the fast and medium cycles were as a result of compression of end gas from fast propagating flames. The occurrence of the peak pressure at crank angle closer to TDC for the fast cycles compared to the medium, suggests faster mass burning rate for the fast cycles compared to the medium and slow cycles. Higher knock intensities observed at the onset of knock, followed by lower knock intensities in the knocking cycles, suggests the auto-ignition of the region of the end-gas with higher energy level with subsequent auto-ignition of smaller end gas patches. Higher knock intensities observed in the fast cycles when compared to the medium cycles could have been influenced by higher flame compression of the end gas by faster propagating flames. E5 fuel exhibited the best anti-knock behaviour while PRF 95 performed least. This could be attributed to the large constituent of aromatics in E5 fuel and its absence in PRF 95. E5 and ULG 98 fuel were observed to be the most suitable amongst the fuels tested for heavily supercharged SI engines.
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Effect of Spark Timing on Combustion Process of SI Engines using MATLAB

Effect of Spark Timing on Combustion Process of SI Engines using MATLAB

Ignition timing is very important for improving performance of modern SI Engines [5]. In an ideal four stroke SI engine, compression and expansion take place during 180° of crank rotation and combustion takes place instantaneously at TDC. During combustion the volume remains constant and there is a sudden pressure rise [1]. However, in an actual spark ignition engine, combustion does not occur instantaneously. It is initiated by a spark produced before TDC at a definite time which affects the maximum pressure generated and the corresponding crank angle, indicated mean effective pressure, and consequentially the work done in cycle and so the thermal efficiency. Studies have been conducted to demonstrate some of these effects previously [8]. This paper essentially aims at generating MATLAB programs which will take input of spark timing and will give output of the required engine parameters so that salient conclusions can be drawn. Computer models of engine processes are important tools for analysis of engine performance [6].
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Modification and Testing of Manually Operated Rocker
Sprayer

Modification and Testing of Manually Operated Rocker Sprayer

This study was done to modify the existing rocker sprayer. Manually operated lever was replaced with an electric motor and the speed of the motor was regulated by a voltage regulator. For determining the power requirement, the speed of the manually operated sprayer was regulated by subjects’ analysis and was found to work with an average speed of 74 rpm of the piston movement. Modification was done to get the same boom length of the sprayer. The power requirement of the sprayer was determined as 0.016 hp, but 0.25 hp motor was used due to market’s unavailability. A model of the slider crank mechanism was developed to study the effect of connecting rod length, crank angle and rpm on the piston displacement and the linear velocity for the modification of the rocker sprayer. It was found that, the piston displacement and the velocity decreases with an increase in the length of the connecting rod and vice versa. Piston linear velocity was found to be maximum at two crank angles in the range of 90-120° and 260-280°. The weight of the implement was about 4.75 kg. The sufficient pressure of 80-90 psi was obtained for the purpose of spraying, as found in the existing manually operated sprayer. The slider-crank mechanism is useful for the development of the power operated sprayer and for other agricultural machineries.
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