7.9 Mol of Helium Are in a 18 L Cylinder. The Pressure Gauge on the Cylinder Reads 65 Psi .

The Combustion Process and Combustion Analysis

A.J. Martyr , M.A. Plint , in Engine Testing (Fourth Edition), 2012

Engine Indicating Pressure Transducers (EIPTs)

Cylinder pressure measurements play the key role in whatsoever engine indicating and combustion analysis work, but since modernistic engine enquiry requires more than peak cylinder pressure to be measured against crank angle, pressure transducers are also made for the following roles in engine indicating work:

•

Cylinder pressure for R&D work, optimized for precision of measurement.

•

Cylinder pressure for the in-service monitoring of non-automotive engines, optimized for durability.

•

Intake and exhaust manifold pressure for R&D work, optimized to handle the temperature and force per unit area ranges of each work surroundings.

•

Fuel line pressure for R&D work for pressure pulse measurement up to 4000 bar.

Cylinder Pressure Transducers

Nigh transducers are of the piezoelectric type and require their signals to be processed past matched accuse amplifiers. However, there are now dynamic pressure level transducers available that work by the deflection of an optic cobweb, in the range of temperatures and pressures that make them suitable for CA work. Although these devices are not yet used in mainstream CA piece of work they tin can be made downwardly to diameters of effectually ii mm so they may find a place in "hard to locate" cylinder-caput installations.

Up to the early 1990s commercially available force per unit area transducers were more than temperature sensitive, both in terms of cyclic effects and absolute maximum working temperature, than units available today. The temperature effects are due to exposure to the combustion process and most would fail if required to operate above 200 °C, measured at the diaphragm; consequently, nearly EIPTs were water cooled.

Water-cooled EIPTs are even so required for some situations, including turbocharged engines, and it is vital that the cooling system meets the following weather condition:

•

The transducer cooling organisation, while in use, must exist integrated with the control arrangement shutdown circuit to ensure that the cooling is running during the consummate testing cycle, including engine showtime and cool down. Transient dips in the coolant supply pressure volition take chances permanent loss of the transducers.

•

The water should be distilled (deionized) and filtered. The transducers have very small passages that will become ineffective if blocked by scale.

•

The deionized h2o should be supplied at a abiding pressure level that is as low as possible to ensure flow; this is to avoid changes in internal transducer pressure that could corrupt output signal.

Uncooled EIPTs are continually being developed; currently models will operate to 400 °C, which covers the bulk of naturally aspirated automotive work outside specialist and motor sport development work.

An important characteristic of whatever EIPT is its circadian temperature drift, which is the range of measurement fault due to the heating of the transducer'due south diaphragm over the cylinder'southward working bike. The lower the drift value, the higher the unit accuracy; the figures quoted are typically improve than ±0.three to ±0.6 bar depending on the blueprint and size. Since at that place is no standard process to define how circadian drift has to exist measured, it makes a direct comparison betwixt the values measured by transducers from different manufacturers almost incommunicable, which is why almost all examination facilities utilize all their CA equipment in the complete transducer–cable–amplifier–analysis chain from just i manufacturer. At that place is some similarity with employ of exhaust emission equipment, where direct result comparisons across manufacturers can be hard in detail.

The choice of force per unit area transducer from the wide range now commercially available [4,5] needs to exist made with corking intendance, as the correct mounting in the engine and its integration inside the calibrated organisation are vital ingredients in obtaining optimum results.

Mounting of the Cylinder Pressure Transducer

The precision of pressure measurement is critically dependent on the location and mounting of the probe in the cylinder-head. The key selection facing the exam engineer is whether the cylinder-caput can exist machined with a hole terminating at a suitable location; if not, then the simply option is to use either a transducer designed to supersede the glow plug in a diesel or a special combined transducer–spark plug in an SI engine (see Figure 15.13).

Figure 15.xiii. Role department through a typical combined spark plug and cylinder pressure transducer.

If the head tin can exist machined and there is space and material in the engine head construction to support the transducer, at that place are iii types that might be used: probe types that are usually the thinnest, plug types, and the most common, threaded types. Note that some pressure transducer manufacturers supply the cut tools designed to machine the recesses required to firm the specific model.

Threaded EIPTs are made in various sizes, usually ranging from 5 mm upward to 18 mm. They are normally inserted into the combustion chamber so that their tips are flush with the parent combustion bedroom cloth.

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The combustion process and combustion analysis

Anthony J. Martyr , David R. Rogers , in Engine Testing (Fifth Edition), 2021

Cyclic-based work

Cylinder pressure level data tin can be used to calculate the work transfer from the gas to the piston. This is mostly expressed as the indicated mean effective pressure level (IMEP) and is a mensurate of the piece of work output for the swept book of the engine. The result is a fundamental parameter for determining engine efficiency as it is independent of speed, number of cylinders, and deportation of the engine. The gross IMEP adding is basically the enclosed surface area of the high-pressure part of the P–5 diagram and tin can be calculated as the integral of the pressure level volume curve divided by the swept volume [Eqs. (xvi.half-dozen) and (xvi.7)]. IMEP tin besides be presented in net grade. The gross vales only accounts for the work done on, or by, the piston during the compression and expansion strokes, the cyberspace value also includes the piece of work washed by the piston in the frazzle and inlet strokes [Eq. (16.eight)].

The IMEP for a single cylinder can be computed thus:

(16.6) $\text{IMEP}\left(\mathrm{Gross}\right)=\frac{1}{{\mathrm{Five}}_s}{\oint }^{}pdV$

This can exist explained by the following equation:

(xvi.7) $\mathrm{IMEP}\left(\mathrm{Gross}\right)=\frac{\left(\text{area of high pressure loop}\right)}{V_s}$

This finer gives energy released or gross MEP (gross piece of work over the compression and expansion wheel, GMEP).

Net IMEP for a single cylinder is derived by subtraction in the following equation:

(16.viii) $\mathrm{IMEP}(\mathrm{Net})=\frac{\left(\text{area of high force per unit area loop}\right)-\left(\text{area of low pressure loop}\right)}{V_s}$

Integration of the depression-pressure level (or gas exchange) role of the cycle gives the work lost during this office of the process (pumping losses, also known as PMEP). Subtraction of this value from the gross IMEP gives the net IMEP (NMEP) or bodily piece of work per bicycle [Eq. (sixteen.9)]. Therefore it tin can be stated that:

(16.nine) $\mathrm{NMEP}=\mathrm{GMEP}-\mathrm{PMEP}$

where GMEP is the gross indicating mean effective pressure level (area of power loop), PMEP is the pumping indicating hateful constructive pressure (area of pumping loop), and NMEP is the net indicating mean effective pressure level.

The virtually important factor to consider when measuring IMEP is the TDC position (see the "'Exact' Determination of True Height Expressionless Center Position" section in this chapter).

Note that, in guild to optimize the calculation, information technology is not necessary to learn data at high resolution. Measurement resolution of a maximum of i degree crank angle is sufficient for MEP calculations; higher resolution than this does not improve accurateness and is a waste material of system resource. Also of import for accurate IMEP calculations are the transducer properties, mounting location, and stability of the transducer sensitivity during the engine cycle.

If several cycles are analyzed, the coefficient of variation (CoV) of IMEP tin be determined as shown in the following equation:

(xvi.10) $\mathrm{IMEP}\cdot \mathrm{CoV}=\frac{\sqrt{{\sum }_1^n{\left(\mathrm{imep}-\overline{\mathrm{imep}}\right)}^2/(\mathrm{due\; north}-1)}}{\overline{\mathrm{imep}}}$

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24th European Symposium on Figurer Aided Process Applied science

Katarzyna Bizon , ... Bianca M. Vaglieco , in Computer Aided Chemical Engineering science, 2014

1 Introduction

In-cylinder pressure is perhaps the most important parameter for engine diagnosis and control equally it contains useful information on the phenomena taking identify in the combustion chamber. Direct measurement of in-cylinder pressure requires expensive pressure probes ( Docquier and Candel, 2002) which are able to resist astringent weather condition and besides represent a substantial intrusion. A better solution would certainly exist the application of non-intrusive sensors placed exterior of the combustion chamber, that could indirectly give information on the quality of the combustion procedure. Good candidates should exist able to mensurate quantities which strongly correlate with the in- cylinder pressure. Particularly, combustion force per unit area betoken and engine block vibrations have been found as beingness well related to each other, both for unmarried (Bizon et al., 2011) and multi-cylinder (Chiavola et al., 2010, ii cylinders, and Taglialatela et al., 2011, 4 cylinders) engine applications Information technology is thus logical to try and correlate combustion-related quantities with the vibration bespeak coming from accelerometers placed externally on the engine block. However, the transformation of the vibration signal into in-cylinder pressure is non straightforward. This is not only due to the strongly nonlinear character of this relation but also, plain, to the fact that the vibration signal contains some noise introduced, among the others, past piston slaps and valve impacts. In this view, artificial neural networks (ANN) are recognized as a tool having a great flexibility in the approximation of not-linear mappings and capable to learn both the associations and patterns in the measured data, even in presence of noise and uncertainty. Earlier studies accept shown the potential of neural networks in engine diagnostics. Unlike types of ANN models have been used to model, for example, the relationship betwixt crankshaft speed and parameters derived from in-cylinder pressure. Particularly, Saraswati and Chand (2010) proposed a recurrent ANN for reconstruction of in-cylinder pressure of a spark ignition engine; Gu et al., 1996, used a non-parametric RBF ANN to predict force per unit area in the cylinders from the instantaneous angular velocity of the crankshaft, whereas in (Taglialatela et al., 2013) the same issue was approached by a multilayer perception ANN. Other applications include prediction of diesel/biodiesel fuel mixture properties (Kologeras et al., 2010) and exhaust NO₁₀ emission prediction based on engine operating variables (Krijnsen et al., 2000). Recently, Wai and Vishy (2013) investigated the employ of ANN as virtual sensors for emissions prediction and control.

The nowadays study aims at edifice an efficient and robust radial basis functions (RBF) – ANN model able to reconstruct existent-time the in-cylinder force per unit area signal starting from the vibration signal.

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An optical investigation of a cold-kickoff DISI engine startup strategy

P. Efthymiou , ... J. Harris , in Internal Combustion Engines: Performance, Fuel Economy and Emissions: IMechE, London, 27–28 November 2013, 2013

3.one In-Cylinder Pressure level Measurements

In-cylinder pressure traces were captured across 300 engine cycles for all test points. Hateful in-cylinder pressures and a histogram of the mean peak pressures are illustrated in Figure half dozen (relative to After Top Expressionless Centre (ATDC) of the compression stroke). The upper graph highlights the distinct differences and distribution of peak in-cylinder pressures achieved inside the first few seconds of the startup bike with a range from 3.2   bar at Point 3 to 16.0   bar at Point 2.

Figure six. In-cylinder pressure profiles for points testing along the cold-startup cycle at 23   °C cylinder head, liner and intake conditions

There is also a notable departure in the location of the peak pressure which is due to parameter scheduling likewise as the nature of combustion structures. The first firing cycle (Point 1) exhibits an interesting pressure trace in which two peaks are seen occurring at TDC and 30°CA ATDC which is similar to the premixed and main combustion peaks ordinarily seen in diesel engines (28). It is probable that the very low engine speed and resulting flame speed and construction produced a burn charge per unit lower than that of typical premixed turbulent combustion. The second elevation however, does show that some useful piece of work has been extracted from the cylinder which is required to initiate successful starting of the engine after cranking.

Point 2 has parameter scheduling which is more typical of the Jaguar Land Rover Standard test point optimised for homogeneous stoichiometric combustion which has been used on previous research on the engine (19,22,24). A shine rise in pressure during the compression stroke is continued after an optimised spark timing, this delivers a maximum in-cylinder height force per unit area (located at 30°CA ATDC) and extracts the almost corporeality of work from the cylinder with a about-stoichiometric λ ratio recorded at 1.i. During the initial acceleration ramp-upward of an engine, maximising work from the engine is particularly important. Once a faster speed is achieved, mechanically driven pumps such as the fuel pump for instance, can deliver the optimum pressure level to the fuel runway required for correct fuel atomisation and subsequent idealised combustion. Despite the fuel pressure level not being fully developed, resulting in poor atomisation and sub-optimal combustion (discussed in next section), piece of work extraction has been maximised for this indicate.

The pressure distribution of the final three points of the startup are like in the fact that less work is extracted from the cylinder due to the injection and spark scheduling. Point 3 produced a pinnacle pressure at TDC of 3.ane   bar, similar to mechanical piston pinch and was observed to combust in a lean and unstable style when run continuously on the optical engine. The increment in engine speed to 2000   rpm for a brief period is known to be customer-driven, then it can be assumed to provide no specific benefit to engine emissions or frazzle temperature. A recorded λ ratio of 1.5, however, outlines why combustion instability could take occurred, excess air may have provided an overly lean charge mixture which prevented the flame from successfully propagating. The flame may be lean when run in steady-state but the presence of excess fuel is likely from the preceding engine run-up conditions. This was confirmed when testing Betoken ii as skip-firing was required. The in-cylinder optical analysis will provide a farther insight into this.

Points 4 and v, from their parameters, can be identified as 'catalyst warmup' points with different strategies for high exhaust temperature generation. Both points use a large quantity of fuel injected twice per bike and in different ratios to combust at a very rich λ ratio of 0.vii. The pressure level trace for Bespeak 4 demonstrates this late burning with a characteristic depression-gradient force per unit area trace 'tail-off' during the expansion stroke. Betoken 5 also exhibits an interesting pressure trace peak during expansion which occurs much later. This is due to farther retarded spark scheduling which causes combustion to occur after TDC and into the expansion and frazzle stroke. Point 5 forms the master period of the catalyst heating phase and entails a higher manifold pressure and a further retarded spark than Point iv to generate a larger quantity of college temperature gas during the exhaust stroke.

The lower graph in Figure six illustrates the top pressures on a histogram and is overlaid with their corresponding coefficients of variation (COV). The COV parameter is a mensurate of the bike-to-cycle variability (divers as the ratio of the standard departure to the mean) and gives a practiced indication to engine stability. A distinct correlation is seen betwixt the fourth dimension from engine first and COV with the final 3 points having lower COV values. A possible reason for this is that the last 3 data points were run at college engine speeds and had more established and developed in-cylinder turbulence structures. These small-scale structures, peculiarly at TDC, have been shown to be critical to engine operation and stability (eighteen,22). Some other trend of the latter points is a fully developed fuel rail pressure level, assuasive the injector to work within its operating range and atomise the fuel correctly.

Figure 7 presents MFB times for Betoken 1, 2, 4 and 5 which represent the fourth dimension taken for each change in force per unit area with crank bending (dp/dθ) to achieve either ten%, l%, or ninety% of the total mass burned. Point 3 has been omitted as the unstable combustion produced times exceeding reasonable tolerances, this indicates that the ignited charge in the cylinder had very little effect over the natural in-cylinder pressure level compared to when the engine is motored. The graph shows that Points i and ii, despite using a fuel pressure that is non fully developed, produced burn down times faster than those of Points 4 and 5 (up to 50°CA difference in the 90% MFB example) and similar to those in a similarly configured engine at 23   °C with a fuel pressure of 150   bar (24). This highlights the dominant effect that injection and ignition scheduling has on the latter points with the aim of generating high exhaust temperature. Information technology is also reasonable to presume that the effect on the burn rate is non-detrimental to the engine's primary aim during this position in the common cold startup bike. The COVs of the presented data show the adverse effect of cycle-to-cycle variability on the burn rate of the Signal 1. The detrimental effect of scheduling parameters to generate exhaust heat can too be seen in Points iv and 5, despite having a fully developed fuel rail pressure and well established in-cylinder turbulence at these speeds (19). The effect of parameter scheduling on flame kernel cosmos, growth and speed has been investigated in the adjacent section of this paper and aims to build a farther agreement on what has been observed from the in-cylinder pressure information.

Figure 7. Bar graphs outlining peak pressure and corresponding COVs for startup points 1, two, 4 and five at 23   °C cylinder caput, liner and intake conditions

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Utilization of biofuels in diesel engines

T. Le Anh , ... K. Wattanavichien , in Handbook of Biofuels Production (2nd Edition), 2016

Engine indicating data

Comparison of in-cylinder pressure level, fuel line force per unit area, fuel injection rate, oestrus release rate, net oestrus release, and mass fraction burned is shown in Fig. 23.12. The measurement of in-cylinder force per unit area and fuel injection line force per unit area has indicated that 10% CPO diesel fuel has approximately 1 degrees of early injection timing compared with diesel. The 10% CPO diesel also has longer ignition filibuster and higher amount of fuel injected mass (grand _f) due to its lower free energy density. The maximum in-cylinder pressure of 10% CPO diesel fuel is like to diesel. Net heat release and mass fraction burned of 10% CPO diesel are likewise lower than diesel fuel.

Figure 23.12. Comparing of in-cylinder pressure level, fuel line pressure, fuel injection rate, heat release rate, net heat release, and mass fraction burned equally engine operate with diesel and 10% CPO diesel at 2000   rev/min, 30   Nm (Wattanavichien and Traiphopphoom, 2006).

Comparison of maximum in-cylinder pressure (P _max), SOI, ignition delay, and fuel injected mass (m _f) every bit engine operated with diesel fuel and 10% CPO diesel are summarized in Table 23.iv.

Table 23.iv. Comparison of maximum in-cylinder pressure (P _max), SOI, ignition delay, and fuel injected mass (one thousand _f) every bit engine operate with diesel and 10% CPO diesel (Wattanavichien and Traiphopphoom, 2006)

Test betoken	P max (bar)		SOI (°CA)		Ignition delay (ms)		chiliad f (mg/cycle)
	Diesel fuel	ten% CPO diesel	Diesel fuel	ten% CPO diesel fuel	Diesel	10% CPO diesel	Diesel	10% CPO diesel
Idle	53.26	53.31	−iv.0	−iv.0	ii.08	2.two	6.22	7.04
1000   rpm, 30   Nm	58.45	59.45	−ten.5	−11.5	2.08	two.17	nine.63	10.77
2000   rpm, thirty   Nm	61.48	61.84	−11.0	−11.five	1.54	1.l	ix.99	10.88
2000   rpm, fifty   Nm	61.72	61.74	−x.0	−x.0	0.46	0.46	12.64	13.97
2250   rpm, xx   Nm	64.98	64.97	−10.v	−11.0	0.78	ane.04	viii.72	nine.81
2750   rpm, xx   Nm	63.90	64.66	−9.0	−9.0	0.21	0.21	9.56	10.46

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Eco friendly biofuels for CI engine applications

B. Ashok , K. Nanthagopal , in Advances in Eco-Fuels for a Sustainable Surroundings, 2019

In-cylinder peak pressure

Higher peak in-cylinder pressures are attained in the case of booze blends when compared to diesel because of the availability of more than oxygen in the fuel. The longer ignition delay period also contributes to this higher meridian pressure considering the fuel accumulated during the delay undergoes rapid combustion and adds to the pressure in the sleeping room [nine, 25]. As the engine load increases, more than fuel is injected into the cylinder and the in-cylinder pressure increases. Consequently, the pressure in the cylinder increases with increase in engine load and reaches a peak at loftier engine loads. Amongst alcohol blends, combustion of college alcohols resulted in college cylinder pressures every bit compared to lower alcohols. This can be attributed to the reduction in viscosity and increased volatility that result with the addition of higher alcohols; these changes lead to enhanced atomization and hence improve fuel air mixing, improving fuel combustion.

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In-cylinder studies of split injection in a single cylinder optical diesel fuel engine

1000.R. Herfatmanesh , ... 50. Ganippa , in Internal Combustion Engines: Improving Performance, Fuel Economy and Emissions: IMechE, London, 29–thirty November 2011, 2011

4.2 Effect of Dwell Angle on Combustion and Emissions

The in-cylinder pressure and heat release rate data for the separate injection strategies are depicted in Figure 3. The peak in-cylinder pressure increased as the beginning injection timing was farther retarded. Although improved fuel evaporation was expected closer to TDC, the trend observed was mainly attributed to the inconsistency in the total injected fuel quantity equally illustrated in Figure 2.

Figure 3. In-cylinder Pressure and Estrus Release Charge per unit for Split up Injection Strategies

The rut release rate curves exhibited a drop due to accuse cooling issue shortly after the onset of the first injection. Subsequently, the rate of estrus release rapidly increased due to the premixed combustion. The top heat release rate increased every bit the timing of the first injection was further retarded, this was mainly due to improved fuel evaporation and mixing processes with the exception of F4 strategy whereby the greater fuel quantity injected was the primary cistron. Nonetheless, the reverse was observed during mixing controlled combustion phase, this was attributed to the limited time available for the combustion procedure to complete.

F2 strategy was selected for further assay using optical diagnostic techniques since the desired fuel quantity was injected for this strategy. The loftier speed image sequence for F2 strategy is presented in Effigy 4. The offset 4 frames bear witness the fuel spray jets evolution from the outset to the stop of the offset injection where in the last frame, fuel sprays were almost fully evaporated. The fuel spray impingement on the piston basin wall was evident in frames iii and four where the tip of fuel sprays spread forth the piston wall upon touch on, generating a mushroom blazon structure. The frames presented at TDC, 0.nine^◦ CA ATDC, 1.8 ^◦ CA ATDC and three.6 ^◦ CA ATDC testify the paradigm sequences from the commencement to the end of the 2nd injection.

Figure 4. Combustion Image Sequence for F2 Strategy

The first combustion phase which was credible from the heat release rate information at - 6.viii^◦ CA ATDC, Effigy 3, resulted in no visible combustion until -3.vi^◦ CA ATDC where the first visible combustion was observed. The flame was propagated at the tip of the fuel sprays where the fuel vapour was highly concentrated, spreading along the periphery of the fuel sprays where the fuel evaporation rate was considerably college as shown in frames seven and 8. In the next frame at three.half dozen^◦ CA ATDC, the flame at the tip of the sprays spread outwards, following the mushroom similar structure previously created due to fuel spray impingement. In the subsequent frame at 8.1^◦ CA ATDC, the flame was fully propagated covering the entire combustion bedroom. From the epitome sequences at 19.8^◦ CA ATDC and 31.5^◦ CA ATDC, information technology was evident that the flame intensity was diminishing during the expansion stroke and moving clockwise due to the swirl motion.

The flame temperature and KL factor image sequences for the F2 strategy are shown in Effigy 5. Modest combustion at relatively low temperature was detected in the get-go frame at 0.9^◦ CA ATDC, which was in good agreement with the high speed image taken at this creepo angle. The subsequent frames at ane.8^◦ CA ATDC and 3.6^◦ CA ATDC showed flame propagation around the tip and periphery of the sprays where high concentration of premixed fuel vapour and air were nowadays. As the combustion developed, the flame temperature and soot formation increased until 8.i ^◦ CA ATDC, due to premixed combustion. Subsequently, the flame temperature and soot concentration steadily decreased during the expansion stroke. Although a luminous flame was detected in high speed images at creepo angles subsequently in the expansion stroke, no information could be detected at these crank angles due to inadequate sensitivity of the detection arrangement at the selected wavelengths.

Effigy 5. Flame Temperature and KL Factor Images for F2 Strategy

The IMEP, soot, Nox and uHC emissions results are depicted in Figure 6. The substantial variations in the engine output and frazzle emissions for the investigated dissever injection strategies were primarily acquired past the inconsistency in the total fuel quantity injected. However, F2 and A3 strategies tin can be compared since the total fuel quantity injected was almost identical. In comparing, F2 strategy resulted in higher IMEP values with considerably lower uHC, indicating improved fuel evaporation and mixing processes with significantly lower combustion noise. All the same, this injection strategy led to considerably college soot and NOX emissions. The onetime was mainly attributed to the injection of fuel sprays into burning regions during the second injection, resulting in the propagation of highly luminous flame within the combustion chamber as evident in Figure 4 while the latter was due to relatively higher in-cylinder temperature.

Effigy 6. Engine Output and Emissions Values for Carve up Injection Strategies

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Human B&W low-speed engines

Malcolm Latarche , in Pounder's Marine Diesel Engines and Gas Turbines (10th Edition), 2021

Cylinder pressure measuring system

Reliable measurement of the cylinder pressure is essential for ensuring sustained 'as new' engine performance. A conventional mechanical indicator in the hands of skilled engineroom staff can provide reasonable data on this parameter, merely the process is quite time consuming, and the cylinder pressure data derived are not bachelor for analysis on a computer. Some valuable information is therefore less likely to be used in a further analysis of the engine condition.

A computerized measuring system with a high-quality force per unit area pick-upward connected to the indicator diameter—pressure measuring organisation (PMI) offline—was developed past Man Diesel for application to its MC engines. For the ME, however, online measurements of the cylinder pressure are necessary, or at least highly desirable. In this case, the indicator cock cannot exist used because the indicator diameter would clog up after a few days of normal operation.

The strain-pin blazon of pressure sensor was applied instead. Hither, the pressure level sensing element is a rod located in a bottom hole in the cylinder cover, in shut contact with the bottom of the hole, and close to the combustion chamber surface of the cylinder cover. The sensor thus measures the deformation of the cover caused by the cylinder pressure without beingness in contact with the ambitious combustion products. The position of the sensor also makes it easier to prevent electrical racket from interfering with the cylinder pressure betoken.

The pressure transducer of the above-mentioned offline arrangement is used for taking simultaneous measurements for calibrating the online arrangement. A calibration curve is determined for each cylinder by feeding the 2 signals into the computer in the calibration mode. The fact that the same high-quality pressure transducer is used to calibrate all cylinders ways that the cylinder-to-cylinder balance is not at all influenced past differences between the individual pressure level sensors.

Both online and offline systems provide the user with valuable assistance in keeping the engine functioning at 'as new' standard, extending the TBOs and reducing the workload of the crew. The systems automatically identify the cylinder being measured without any interaction from the person carrying out the measurement (because the system contains data for the engine'southward firing lodge). Furthermore, compensation for the crankshaft twisting is automatic, exploiting proprietary data for the engine design. If there is no such compensation, the hateful indicated pressure volition be measured wrongly, and when the figure is applied to accommodate the fuel pumps, the cylinders will not have the same true compatible load later the adjustment, although it may seem then. (Crankshaft twisting may atomic number 82 to errors in mean indicated pressure of some 5%, if not compensated for.)

The computer executes the deadening task of evaluating the 'indicator card' data, which are now in computer files, and the cylinder pressure data can be transferred directly to MAN Diesel's CoCoS-EDS engine diagnostic system for inclusion in the full general engine operation monitoring. The outcome presented to engineroom staff is far more than comprehensive, comprising a listing of necessary adjustments. These recommendations take into account that the condition of the nonadjusted cylinders changes when the adjustments are carried out; it is not necessary therefore to check the cylinder pressure subsequently the adjustment.

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Diesel Engines

M.J. TINDAL , O.A. UYEHARA , in Internal Combustion Engines, 1988

6. Cylinder pressure and dissociation effects

Increment in the peak cylinder pressure has ii important effects: (a) The mechanical stresses in the engine components increase. Nowadays large engines operate with peak pressures upwardly to 170 bar ( Herrman and Magnet, 1985). High gas pressures as well tend to cause a rise in friction losses because of the increased force per unit area backside the peak piston ring. (b) The caste of dissociation is reduced, other things being equal. This is illustrated in Figure 38 which relates to the combustion of ethene (C_twoH_four) with air in stoichiometric proportions. Figure 38 gives an indication of the way in which the equilibrium composition of the products of combustion of ethene varies with pressure.

Effigy 38. Result of pressure on dissociation.

Reference has already been made (Section I) to the fact that the DI engine generally has a lower specific fuel consumption than the IDI engine. This topic volition be further discussed hither with particular reference to the importance of local air/fuel ratios. In general, if fuel aerosol burn in a region where at that place is a shortage of oxygen, big amounts of carbon monoxide volition be formed. If more oxygen becomes bachelor later on in the combustion procedure this carbon monoxide may or may non oxidize to carbon dioxide, depending on the temperature. If, by the time the actress oxygen becomes available, the temperature has fallen below about 1800 K, the carbon monoxide concentration will remain "frozen" at its earlier level and at that place will exist no pregnant conversion to carbon dioxide.

Where the air/fuel mixture is non homogeneous, therefore, a significant amount of carbon monoxide may announced in the exhaust, even though the overall equivalence ratio may be significantly below unity. This is illustrated for the case of a spark ignition engine in Figure 39 (Uyehara, 1980a) where the percentage of carbon monoxide in the frazzle is plotted as a role of equivalence ratio. There are four curves. Curves A and B represent predicted concentrations at the start of the expansion stroke and at exhaust valve opening (EVO), respectively. The calculations were based on a compression ratio of 8 and an air temperature of 830 Grand at the end of the pinch stroke; equilibrium conditions were assumed. Bend C and points D represent measured concentrations of CO. In the test runs to which points D refer, great care was taken to ensure that the induced charge was homogeneous; the fuel and air were thoroughly mixed in a system of tanks and screens before beingness fed to the engine. Information technology will be seen that under these conditions and with a stoichiometric mixture, the proportion of CO in the frazzle was very depression — about 0.3%. Curve C, by contrast, represents data from tests in which the fuel was injected into the inlet manifold, so that it had much less opportunity of mixing properly with the air; thus at the instant of ignition, the equivalence ratio would be expected to vary considerably from point to betoken within the combustion chamber. The overall equivalence ratio was measured; the operating weather corresponded exactly to those of points D. It will be seen that the concentration of carbon monoxide in the frazzle is greater than for the homogeneous mixture; the difference is particularly marked for stoichiometric and weaker mixture strengths. It would announced that, in the case of the homogeneous mixture, the carbon monoxide is formed more or less uniformly over the whole of the combustion bedroom but that when the mixture is not homogeneous, there are local fuel-rich zones in which the production of carbon monoxide is relatively high. As the temperature falls below about 1400°C the carbon monoxide concentration freezes. It would seem that, if the overall equivalence ratio is less than 1, then to proceed down the concentration of carbon monoxide in the exhaust it is desirable to take a homogeneous charge in the combustion bedroom, with no zones of local richness.

Figure 39. Concentration of carbon monoxide — dependence on equivalence ratio. (Uyehara, 1980).

Reprinted with permission © 1980 Order of Automotive Engineers, Inc. Copyright © 1980

Consider an IDI engine in which the volume of the pre-chamber is equal to one half of the total clearance volume, the other half consisting of the clearance infinite in the cylinder together with the connecting passage. At high loads the mass of air trapped in the pre-combustion sleeping accommodation at TDC will be significantly less than half the total mass present. This is considering the air in the pre-sleeping room will have picked upward oestrus from the relatively hot surfaces of the passageway and the pre-chamber and volition therefore be at a somewhat lower density than the air in the space above the piston. All the fuel is injected into the pre-combustion chamber; thus the equivalence ratio there will be loftier and the ratios [CO₂]/[CO] and [H₂O]/[H₂] volition be relatively depression. Although much more oxygen becomes available later in the combustion process, when the partly burned gases sally into the master cylinder, the oxidation of the CO is far from consummate and once the temperature falls below 1800 Yard, the reaction becomes extremely deadening. Thus it is possible for the level of CO in the exhaust to exist quite high even though the overall equivalence ratio is weaker than stoichiometric.

Wherever at that place are pockets of rich mixture in an overall lean mixture, there is likely to be a significant concentration of carbon monoxide in the exhaust gas. Associated with the carbon monoxide will be hydrogen, some of which volition exist unable to oxidize to water when the local equivalence ratio is rich. The relative proportions of CO₂, CO, H₂O and H₂ are governed by the "water gas" equation:

$\text{CO}+{\text{H}}_2\text{O}\rightleftharpoons {\text{CO}}_2+{\text{H}}_2$

The dissociation constant for this reaction is given by

$\begin{array}{c}{K}_{\text{p}}=\frac{p_{{\text{CO}}_{\mathrm{two}}}{p}_{{\text{H}}_2}}{p_{\text{CO}}{p}_{{\text{H}}_2\text{O}}}\\ =\frac{{\mathrm{due\; north}}_{{\text{CO}}_2}{\mathrm{north}}_{{\text{H}}_2}}{{\mathrm{north}}_{\text{CO}}{n}_{{\text{H}}_2\text{O}}}\end{array}$

since the number of moles of each species is the same (one).

From measurements of the composition of the exhaust gas, information technology is possible to work backwards to determine K _p and hence, by reference to tabulated values of K _p as a function of temperature, to decide the temperature at which the reaction became frozen. Figure 40 shows such a plot. In this case the estimated value of the freezing temperature was 1670 K but nigh values quoted in the literature are higher than this — around 1800 1000.

Figure xl. Determination of "freezing" temperature from exhaust gas limerick.

As pointed out earlier, when the mixture is not homogeneous and rich pockets be in some parts of a charge which is lean overall, carbon monoxide and hydrogen volition be present in the products of combustion too as carbon dioxide and water. Efficient combustion requires that the proportions of carbon dioxide and water be equally high every bit possible and that they be formed early in the expansion stroke — which implies a close approximation to constant book combustion. In the IDI engine information technology is inevitable that relatively large amounts of carbon monoxide and hydrogen are formed in the pre-combustion chamber. These gases menstruation through into the space to a higher place the piston where they mix with air. Reactions occur only equally the piston descends on the expansion stroke the temperature falls; when it reaches the disquisitional value, the concentrations of the various species freeze at their current levels.

Watson and Kamel (1979) used a calculator model to compare the rates of combustion in otherwise similar DI and IDI engines. Figure 41 shows plots of the variation in rate of combustion with crank angle for the ii cases. The greater duration of combustion in the IDI engine is conspicuously evident; although the first stage of combustion in the pre-chamber is completed relatively quickly, the second stage (in the cylinder) is comparatively slow.

Effigy 41. Comparison of rates of combustion for DI and IDI engines.

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Assessment of performance, combustion, and emission behavior of novel annona biodiesel-operated diesel engine

Senthil Ramalingam , Silambarasan Rajendran , in Advances in Eco-Fuels for a Sustainable Environs, 2019

14.3.5 Cylinder pressure

Fig. 14.ix shows the variation of cylinder force per unit area with crank angle for different proportions of AME-diesel fuel and diesel fuel at full load. The cylinder height pressure of AME-diesel blends is lower than that of diesel fuel because of high viscosity and poor volatility. It is observed from Fig. 14.9 that the peak cylinder pressure for diesel is 75   bar and A20 is 65   bar. It is found that peak cylinder pressure is higher for diesel fuel than that of AME-diesel blends and decreases with increasing the percentage of proportions of biodiesel. This is due to the shorter ignition delay and less fuel/air mixture present in the combustion at the time of ignition and thereby more called-for in the improvidence-burning phase than that of the premixed burning phase. This leads to a reduction in cylinder elevation pressure level for biodiesel-diesel blends.

Fig. xiv.nine. Variation of cylinder pressure with crank angle for different proportions of AME-diesel and diesel.

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