Results and Discussions - Discussions and Conclusions

Discussions and Conclusions

4.1 Results and Discussions

The sequence of operations which takes place inside the Diesel engine cylinder is illustrated in Fig 4.1, where the estimated values of cylinder gas pressure and bulk gas temperature are plotted against engine crank angle through the entire four-stroke cycle. Ambient air at BS.

5514 standard condition (l bar pressure, 300 K temperature and with a relative humidity of 60%) is inducted during the intake stroke and is compressed to a high pressure and temperature originated from the high compression ratio of 21, employed in the present study. At 12.5 deg before TOC, diesel injection into the engine cylinder commences and after the consumption of the ignition delay period, assumed 3 deg crank angle in the present study, the heat release from the fuel starts to cause the rapid rise both in cylinder gas pressure and bulk gas temperature.

Both of these quantities reach their maximum values some degrees after the TOC, but not at the same crankshaft angles, and beyond the maximum values these start to fall because of the increase of cylinder volume in the expansion stroke. Hence, gas pressure approaches very close to the atmospheric pressure while the bulk gas temperature remains high at the end of

,j the expansion stroke. Beyond the expansion stroke, exhaust stroke begins where the bulk gas temperatures decline slowly and this stroke is followed by the intake stroke where fresh air

c;,

is inducted into the engine cylinder. It is observed that at the beginning of the intake stroke, bulk g',ls temperature is significantly higher than the ambient temperature. This is due to the

30 presence of hot residual gas in the cylinder and due to the hot intake manifold and the engine cylinder.

In Diesel engines, engine load is varied by varying the amount of fuel injection or heat input, into each cylinder on each cycle. Shown in Figs. 4.2 and 4.3 are the cylinder gas pressures and bulk gas temperatures, respectively, for different amount of diesel injection per cycle into an engine cylinder. It is seen that both of these quantities increase significantly with increase in fuel injection per cycle. Similar cylinder gas pressures at the end of the expansion stroke are observed, as shown in Fig. 4.2. However, bulk gas temperatures are significantly higher for higher fuel injection per cycle, as shown in Fig. 4.3. Values of maximum bulk gas temperatures and cylinder gas pressures for different amount of fuel injection per cycle are shown in Fig. 4.4, where both of these quantities are found to increase linearly with increase in the fuel injection per cycle.

Shown in Fig. 4.5 are the air-fuel ratio, A,and equivalence ratio,

<p,

plotted as a function of diesel fuel injection quantity per cycle. It is observed that, values of the air-fuel ratios decrease with increase in the diesel-fuel injection because of the injection of more fuel while keeping the air flow rate essentially constant. However, values of equivalence ratios are found to increase nearly linearly with increase in the fuel injection rate. Values of equivalence ratio provide more generalized information than the air-fuel ratio data as different fuels require different amount of air for complete combustion, e.g., the stoichiometric combustion of methane requires air-fuel ratio of 17.12, while stoichiometric combustion of diesel requires air-fuel ratio of 14.30 (Fer- guson and Kirkpatrick 2001). In Fig. 4.5, it is observed that only lean combustion is employed in diesel engine because of the very nature of the Diesel engine combustion system where fuel is directly injected into the cylinder and heterogeneous mixture is formed and in case of overall stoichiometric mixture some portion of the charge will starve from oxygen to complete the combustion.

Shown in Figs. 4.6 and 4.7 are the indicator diagrams on a linear and logarithmic basis, respectively, for the same conditions shown in Figs. 4.2-4.4. The area enclosed by the Pressure- Volume trace on the indicator diagrams yield indicated work, Pi' done by the cylinder gas on

31 the piston. From the indicated work, indicated mean effective pressure (imep)can be estimated which is the average pressure that results in the same amount of indicated work. Indicated mean effective pressure is a measure of indicated work per unit swept volume, in a form independent of the engine size and the number of engine cylinders and the engine speed (Lumley 1999).

Plotting the pressure-volume diagram on the logarithmic basis, as in Fig. 4.7, provides several insights, generally not clearly demonstrated in the normal plot in the indicator diagram:

• The pumping loop can be seen more clearly.

• As the compression process is a straight line, it demonstrates that the polytropic compression assumption in this phase is a reasonable one.

• The departure from a straight line just before TOC indicates the start of combustion.

• The combustion is fairly symmetric about TOC, but the end of the combustion is in fact much later and less clearly defined.

Shown in Fig. 4.8 are the indicated work, Pi and brake work, Pb, plotted as a function of diesel injection quantity per cycle and heat input. Brake power is the power available at the flywheel, as compared to the indicated work which is generated inside the engine cylinder. Both power quantities are found to increase with heat additions; however, available break power is always lower than the indicated work because of the effects of energy required in pumping the gases into or out of the cylinder, diving some accessories, turning the camshaft, rubbing the piston rings up and down against the cylinder walls, and turning the crankshafts in the bearing (Lumley 1999). Indicated mean effective pressure, break mean effective pressure are also plotted in Fig 4.8, while these quantities follow the same trend to those of the respective power quantities.

Shown in Fig. 4.9 are the values of indicated efficiency, l1i' brake efficiency, 1/b,and mechanical efficiency, 11m,plotted as a function of heat input from the diesel fuel. It is seen that, value of indicated efficiency, defined by the ratio of indicated work and heat input, decrease with increase in the heat input Tbis is due to the fact that, increase in heat input results in the increase of exhaust gas temperature and increase of which results in higher heat loss. However,

32 brake thenna! efficiency is found to increase with heat input until it reaches a maximum value to define the corresponding power as the rated one corresponding to that speed, and beyond this value, brake thenna! efficiency decreases. This is due to the increase of mechanical efficiency with increase with heat input, as shown in Fig. 4.9. The increase in mechanical efficiency is due to the fact that some amount of energy is' required to power some of the essential auxiliaries, such as oil pump, cooling water pump, radiator fan etc, and this quantity does not increase significantly with output power increase, and hence mechanical efficiency increases with increase in heat input and output power.

Shown in Fig. 4.10 are the brake power output as a function of diesel injection per cycle per cylinder for various engine speeds. It is observed that for all engine speeds, brake power increases with diesel injection per cycle. It is also observed that, for a given amount of fuel injection per cycle, brake power is higher for higher engine speed due to the fact that, engines with higher speed completes more number of cycles in a given amount of time and therefore both output power and heat input rate is higher although fuel injection per cycle remains the same.

Shown in Fig. 4.11 are the variation of brake specific fuel consumption, bsfe, with brake power at different test conditions. It is seen that bs

f

c is high at light loading and its value decreases with increase in the brake power until the brake power reaches the rated value where the bsfc is minimum and any further increase in brake power results in higher bsfc. The values of bsfc for the slower engine operation showed lower values at part load condition, and the situ- ation is found to be reversed at overload condition. It is observed that the rated power increases with engine speed, within the speed range considered in the present study. This behavior of bsfc vs brake power can be explained by considering the different components involved in frictional losses, fraction of which is directly related to the engine speed and fraction of which is directly related to the peak engine cylinder pressure. At low load conditions, the effect of engine cylinder pressure is not significant, rather the frictional losses are dictated by the energy consumed in moving the shafts, valves and pumps, and these losses increase with speed. There- fore, at low load conditions, bsfc is low for low speed operations. However, at higher load, the effect of peak cylinder pressure becomes more significant and peak engine pressure for slower

33 engine operations are higher for the same power generation and hence higher frictional losses.

Moreover, at higher engine speed less heat is lost through the cooling system. Therefore, at high speed operations, values of bsfc's are lower for rated power conditions. However, beyond the rated power, supply of air is not sufficient to ensure complete burning because of the heterogeneous mixture of air and fuel, and in case of overall stoichiometric mixture some portion of the charge starve from oxygen to complete the combustion. Therefore, at rated load, average mixture is always lean and beyond the rated load conditions, combustion is not completed and some of the energy input is lost in the form of incomplete combustion products and results in higher values of bsfc's.

The bsfc is a measure of engine overall efficiency, theses quantifies are inversely related, so that the lower the bsfc, the higher the efficiency of the engine. Shown in Fig. 4.12 are the variation of brake thermal efficiency plotted as a function of brake power at different test conditions. However, for different fuels with different heating values, the values of bsfc's are misleading and hence the brake thermal efficiency is employed in the analyses of engine performance when the engines are fueled by namraI gas and biogas of different compositions.

Values of brake thermal efficiencies are plotted as a function brake mean effective pressure, 1nnep, in Fig. 4.13. The bmep removes the effect of the engine size, and this quantity is roughly comparable even in very different engines, as these different engines burn the same fuel, necessarily under approximately the same conditions and hence generate similar pressures. Differences is /nnep represent genuine engine design difference and not relevant in difference in size (Lumley 1999). For Sf engines, maximum values of /nneps are in the range of 8.5 to 10.5 bar at the engine speed where maximum torque is obtained and at the maximum rated power,/nnep values are 10 to 15% lower. For naturaIly aspirated diesel engines, the /nnep is in the 7 to 9 bar range, with the /nnep at the maximum rated power of about 7 bar (Heywood 1988). As in Table 4.1, the /nnep of a model airplane of 1.6 x 1O-6m3 displacement volume running at 11400 rpm is. 3.2 bar while the Innep of a typical marine engine of 0.433m3 displacement volume running at 160 rpm is 4.5 bar. The relatively small difference between these two fignres are grossly attributed to the compression ratio (Ferguson and Kirkpatrick 2001).

Table 4.1. Comparison of two internal combustion engines (Ferguson and Kirkpatrick 2001) Characteristic

Bore (m) Stroke (m)

Displacement per cylinder (m3) Power per cylinder (kW) Engine speed (rpm) Mean piston speed (m1s)

bmep

Model Airplane 0.0126

0.0l3l 1.6x 10-6 0.1

11400 5.0 3.2

Marine Engine 0.737

1.016 0.433 529 160 5.6 4.5

Values of indicated efficiency, TJi and mechanical efficiency, TIm are plotted as a function of lJmep in Figs. 4.14 and 4.15, respectively, for different test conditions. It is seen tbat tbe values of indicated efficiency decrease witb the increase in the values oflJmep which is in line with tbe observations made in Fig. 4.9. It is also observed that, for a given lJmep, indicated efficiency is higher for higher speed engine operation because of tbe lower time availability for heat transfer losses. It is seen tbat fortbe same lJmep, mechanical efficiency decreases witb increase in engine speed because of tbe associated higher frictional losses at higher engine speeds. It may also be noted that during the idling process, mechanical efficiency reduces to zero value as the engine produces just enough power to overcome frictional losses (Pulkrabek

1997).

Shown in Figs.4.16 and 4.17 are the cylinder gas pressures and bulk gas temperatures, respectively, of tbe diesel engine when fueled by diesel, natoraI gas and biogas of two compositions. In this case, tbe engine is generating similar amount of power consuming similar amount of heat input. It

IS

seen that, cylinder gas pressure is the highest for diesel combustion which is followed by natural gas, BG30 and BG50. However, in case of bulk gas temperatures, natoraI gas fueled case produced the highest temperature and the diesel operations produced the lowest one. In cases of both temperatures and pressures, biogas produced lower value than the corresponding natoraI gas values and the deviation increased witb increase in C02content in biogas. All tbese facts are reflected in tbe brake tbermal efficiencies where diesel engine operations have tbe maximum efficiency which is followed by natoraI gas values and subse- quently by biogas values. The lower pressure and higher temperatures generated in case of

35 natural gas operations are in line with the observations made in some literature (Oguchi and Maita 2(00). Shown in Figs. 4.18 at<! 4.19 are the indicator diagrams on a linear and logarithmic basis, respectively, for the same conditions as shown in Figs. 4.16-4.17. It is seen that, indicator diagrams are very close to each other because of the fact that similar amount of heat addition with different fuels produced similar amount of brake power.

Shown in Fig. 4.20 are the fuel consumption rates plotted as a function of brake power. It is seen that, for all the fuels considered in the present study, fuel consumptions increase with brake power. It is also observed that, natural gas consumption values are lower than those of diesel values because of the fact that natural gas have a higher heating value than that of diesel.

It is also observed that, the fuel consumption rates increase significantly with increase in C02 content in biogas. It is not surprising because of the fact that carbon-dioxide present in biogas does not release any heat, and takes away some heat with the exhaust. Hence, with increase in C02 content in biogas, more fuel is required to produce same amount of brake power.

Shown in Figs. 4.214.24 are the equivalence ratios plotted as a function of bmep for diesel, natural gas, biogas BG30 and biogas BG50, respectively. Experimental results of Rah- man (2003) is also plotted in the figures for comparison. Reasonable agreement between the experimental results and the modeled results is observed for all the cases considered. However, it is also observed that the modeled results are extended to higher bmeps, hence higher brake powers, which is not achievable during the experimentations , and the experimental results suggest rapid rise in equivalence ratio beyond the bmep value of 7 bar. This highlights the shortcomings of the present modeling approach using Weibe function to provide the heat release profile without addressing the detailed combustion processes itself. Hence, the modeled results predict the possibility of higher power generation which is not experimentally achievable.

Figure 4.24 shows the brake thennal efficiency plotted as a function of bmep for straight diesel operation. Brake thermal efficiencies, as shown in Fig. 4.26-4.28, is similar to the diesel fueling as shown in Fig. 4.25, when the diesel is substituted partially by natural gas, biogas containing 70% methane and 30% carbon-dioxide by volume and biogas containing 50%

36 methane and 50% carbon-dioxide by volume, respectively. Experimental results of Rahman (2003) is also plotted in these figures. It is seen that, in case of pure diesel operation as shown in Fig 4.25, experimental values of the brake thermal efficiency reaches its maximum value and the corresponding brake power is the rated power of the engine. Although the modeled results are within 2-3% of the experimental results up to the rated power, it predicts higher thermal efficiency in all cases considered. Moreover, the modeling also failed to show any decrease in thermal efficiency beyond the rated power because of its limitation to model real combustion which causes the efficiency to be reduced beyond the rated power because of the decrease of combustion efficiency resulting from the shortage of oxygen required to bum the fuel completely. However, for,gaseous fuel operations shown in Figs. 4.26-4.28, the reduction in thermal efficiencies beyond the rated power can be anticipated. Hence, the modeled results are within reasonable agreement with the experimental results. However, the modeled results show the reduction of power for higher C02 in biogas which is due to the presence of C02 which do not supply any energy but absorbs some heat when "exhausted in the fonn of raised temperature. The results obtained in the present study, shown in Figs. 4.25-4.28, emphasize the fact that thebmep values effectively remove the effect of engine size and speed, and therefore no speed effect is observed in these results.

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