Espacios académicos

In this part, the results from the experimental and simulation studies are presented and analyzed. First, a basic comparison between gasoline and methanol was performed. Then, the maximum torque of the gasoline-fueled engine was found.

Later on, the measurement results with variable valve timing and boost pressure were analyzed to understand the change in engine BMEP. The downsizing factor of the methanol-fueled engine is then calculated. Finally, the comparison between methanol and gasoline is performed at the same load (maximum BMEP for gasoline) with boost and valve timing controls.

Comparison between methanol and gasoline

Figure 2.3 presents BTE versus engine BMEP at 2500 rpm, exhaust lambda of 1.

The measurements were done with the standard valve timing. At two similar loads (equivalent to BMEP of 7 and 11 bar), the MBT ignition timings for each fuel were applied. To achieve the same BMEP, the throttle opening is smaller for methanol.

For the comparison at wide open throttle (WOT) and boost pressure of 1.34 bar, the same ignition timing (MBT ignition timing for gasoline, 15 CAD bTDC_f) was used.

25 28 31 34 37 40

6 8 10 12 14 16 18

Brake thermal efficiency (%)

BMEP (bar)

Methanol Gasoline WOT 2500 rpm, λ = 1

Figure 2.3: Comparison of brake thermal efficiency versus engine BMEP, with gasoline and methanol, at 2500 rpm and exhaust lambda of 1. At BMEP of 7 and 11 bar, MBT ignition timing was used; at WOT, ignition timing was 15 CAD bTDCffor both fuels.

As can be seen in this figure, the BTE of methanol is higher. This can be explained by a variety of interesting properties of methanol compared to gasoline

METHANOL AS A FUEL FOR A DIRECT-INJECTIONSIENGINE 31

[57]. With a higher heat of vaporization, the in-cylinder mixture is cooler, so the volumetric efficiency increases. Furthermore, the combustion temperature of methanol is lower than gasoline, the heat losses therefore decrease. Another reason is a faster burning velocity of methanol compared to gasoline; the combustion is more isochoric, closer to the ideal Otto cycle. At the same load, the difference between BTE of gasoline and methanol is smaller as the BMEP increases. At WOT condition, not only the BTE, but the maximum BMEP of gasoline is also lower than that of methanol, by around 2 bar. At that point, if the MBT ignition timing for methanol is used, the maximum BMEP and BTE of the engine can be further increased.

Figure 2.4 illustrates the in-cylinder pressures at BMEP of 11 bar and WOT condition, two conditions which were presented in Figure 2.3. At the same engine settings (WOT condition), the in-cylinder pressure at ignition timing of methanol is a bit higher than gasoline due to the fuel evaporation. Methanol has a higher heat of vaporization, the volumetric efficiency increases; therefore the in-cylinder pressure at the end of the compression stroke improves. Because the flame speed of methanol is faster than gasoline, both ignition delay and combustion duration are shorter. Therefore, the combustion starts earlier and the pressure reaches peak sooner. Due to the combustion being more isochoric, the peak in-cylinder pressure increases, the engine BMEP improves. Thanks to these properties, to maintain the BMEP (11 bar) with methanol, a smaller throttle opening angle was used.

Therefore, the pressure at the sparking condition is lower for methanol. However, due to the faster burning rate of methanol, the peak in-cylinder pressure is still higher.

Optimum point of gasoline

To find the optimum point of gasoline, first the S-VVT and boost pressure sweep tests were done to know how BMEP changes at a lower engine speed, 1650 rpm. The baseline point (without valve shifting and boost control) is used as the reference. The BMEP at that point (with knock limit spark advance-KLSA ignition timing, 9 CAD bTDC_f) is 13.72 bar. In all of these sweep tests, this ignition timing was maintained. The valve shifting angle and boost pressure for these sweep test points are presented as the triangular symbols in Figure 2.5. The maximum valve shifting angle and boost pressure is limited at 25 CAD and 49 kPa due to knock, respectively.

The engine BMEP increases with a larger valve shifting angle and a higher intake boost pressure. It can be explained by the reduction of burned gas mass fraction in the combustion chamber due to scavenging and a richer combustion when positive

0

Figure 2.4: In-cylinder pressure curves of methanol and gasoline at BMEP of 11 bar and WOT condition, engine speed of 2500 rpm and exhaust lambda of 1.

valve overlap is applied [80]. With a positive valve overlap, the fresh air from the intake manifold can push the burned gases out of the cylinder and follow the flow to go to the exhaust manifolds during the valve overlap period. As the lambda sensor is an oxygen sensor, therefore it will show λ > 1. The injection duration is then prolonged to achieve the overall exhaust lambda of 1, i.e. the in-cylinder mixture is richer than what is indicated by the lambda sensor. With a lower mass fraction of burned gases and a richer combustion (faster burning velocity, more heat is absorbed during the fuel evaporation), the BMEP improves. The BMEP is also raised with a higher intake boost pressure due to the increase of the engine volumetric efficiency.

Then the measurement matrix following the DoE approach was tested to understand the interaction between the two settings. The boost pressure varied from ∼38 kPa to ∼50 kPa, and the valve shifting angle from 5 to 25 CAD; with steps of 5 CAD. The conditions of these 13 test points are presented in Figure 2.5 (diamond symbols). At each point, the KLSA ignition timing is used. A similar trend was found: BMEP improves with a larger valve overlap period and a higher boost pressure. However, the improvement of the engine BMEP is less significant with the increase of boost pressure when S-VVT is applied. Furthermore, a significant improvement of engine BMEP is found when the valve shifting angle is larger than 20 CAD. Therefore, the test was continued with a higher value of valve shifting angle and higher boost pressure to find the maximum BMEP. Three extra points were tested, located at the top right corner of the test matrix.

METHANOL AS A FUEL FOR A DIRECT-INJECTIONSIENGINE 33

16.29

16.21

15.57

13.72

0 10 20 30

30 35 40 45 50 55

Valve shifting angle (CAD)

Boost pressure (kPa)

BMEP increase

BMEP increase baseline

point

Figure 2.5: The valve shifting angle and boost pressure of the matrix test points at WOT, 1650 rpm, and exhaust lambda of 1. The values present the BMEP (in bar) at that point.

Square symbol: baseline point; triangular symbols: sweep test points; diamond symbols:

DoE matrix test points; circular symbols: extra points.

The first one is at S-VVT of 15 CAD, and a boost pressure that is raised to achieve the maximum BMEP. The maximum BMEP with S-VVT of 15 CAD was 15.57 bar with a boost pressure of 55 kPa. The second point is at a boost pressure of 44 kPa, and the valve shifting angle increase before knock onset. The maximum valve shifting angle at this boost pressure is 30 CAD with KLSA of -0.5 CAD.

The BMEP at this point was ∼16.3 bar. The third point is at the maximum boost pressure and valve shifting angle in the test matrix, 50 kPa and 25 CAD, respectively. The BMEP at this point is ∼16.2 bar. In theory, the BMEP can be further increased if we increase both valve shifting angle and boost pressure;

however, heavy knock will occur. Therefore, no more extra points were tested.

The conditions at these three points present the knock/super knock border for this engine speed. Finally, the maximum BMEP achievable for gasoline at this speed was ∼16.3 bar, improved by ∼19% compared to the baseline case. In order to achieve this BMEP, valve shifting angle of 30 CAD (positive valve overlap of 30 CAD) and boost pressure of 44 kPa were employed. Thus, the in-cylinder mixture is richer than the stoichiometric, and engine efficiency decreases. The potential of achieving higher BMEP with methanol will be tested in the following section.

Methanol sweep tests

In the previous section, the increase of valve overlap, as well as the intake boost pressure, have a positive effect on the improvement of engine BMEP. A similar approach was applied for methanol: three valve strategies (S-VVT 10, S-VVT 25, and S-VVT 40), and three boost pressure (40 kPa, 60 kPa, and 70 kPa) were tested.

When one of these two techniques is applied, the second one is kept at its original setting (boost pressure of 34 kPa for S-VVT sweep tests and valve overlap of -30 CAD for boost pressure tests).

Figure 2.6 presents the BTE as a function of engine BMEP with different valve strategies and boost pressures. The MBT ignition timing of the baseline case (10 CAD bTDC_f) is used for all measurements. The baseline and the optimum points for gasoline (from the previous section) are also illustrated in this Figure. As can be seen, the BMEP improves with a longer valve overlap duration and a higher boost pressure, similar to the findings on gasoline in the previous section. In the S-VVT sweep tests, the BTE is maintained at around 37% in the first three cases (baseline, S-VVT 10 and S-VVT 25), but a significant reduction of BTE is found when the valve shifting angle increases to 40 CAD (down to ∼34%). In the boost control sweep tests, the BMEP increases from 15.5 bar (baseline BMEP for methanol) to over 20 bar. The baseline BMEP for methanol is lower than the maximum BMEP for gasoline. With higher boost pressure or more extended valve overlap period, the engine BMEP increases to over the maximum value for gasoline. No knock was detected during the measurement. Therefore, the maximum BMEP of the engine can be extended with methanol.

However, due to a faster and more advanced combustion, the peak in-cylinder pressure is over 100 bar in some cases. As the engine was designed to work with gasoline, if the engine is to operate with methanol at higher BMEPs a stronger engine block, cylinder head, piston, etc. are needed. Figure 2.7 presents the relationship between peak in-cylinder pressure and engine BMEP. With boost control, the peak in-cylinder pressure increases linearly with the improvement of BMEP. Therefore, the maximum BMEP can be easily estimated, ∼18.2 bar if the peak in-cylinder pressure is to be constrained to 100 bar. The maximum BMEP with valve timing control is ∼17.5 bar. Therefore, the boost control has a higher downsizing potential.

Because the change of maximum torque (or BMEP) is inversely proportional to the change of engine displacement, the downsizing factor (DF) can be calculated as [67]:

METHANOL AS A FUEL FOR A DIRECT-INJECTIONSIENGINE 35

Figure 2.6: Brake thermal efficiency of methanol engine as a function of BMEP with different valve timings and boost pressure. Test conditions for methanol: WOT, 1650 rpm,

exhaust lambda of 1, ignition timing of 10 CAD bTDCf.

R² = 0.9988

Figure 2.7: The relationship between the peak in-cylinder pressure and engine BMEP for boost control and S-VVT control sweep tests. Test conditions: WOT, 1650 rpm, exhaust

lambda of 1, ignition timing of 10 CAD bTDC_f.

DF =Vswept,gasoline−Vswept,methanol

Vswept,gasoline

≈

1/BMEPgasoline,max−1/BMEPmethanol,max

1/BMEPgasoline,max

=1/16.29 − 1/18.244

1/16.29 ×100% ≈ 10.7%

(2.3)

The maximum pressure also constrains the compression ratio. Based on the relationship between boost pressure versus peak pressure, the maximum boost pressure is 57 kPa if the peak pressure is limited to 100 bar. A basic calculation to predict the maximum compression ratio for a NA SI engine is performed. The end pressure P_endand the initial pressure P_iniis related as a function of the compression ratio CR and the specific heat ratio γ, as in equation (5). The subscript “NA”

indicates the naturally aspirated condition.

P_end

P_ini =CR_NA^γNA (2.4)

Assuming the initial pressure is 1 bar, and the specific heat ratio remains constant during the compression stroke. The value of γ at boost pressure of 57 kPa and 0 kPa (intake boost pressure in NA SI engines, WOT condition) is predicted through the relationship between the intake stroke averaged γ versus boost pressures. The intake-stroke averaged γ at each boost pressure is derived from a TPA simulation.

The γ for the boost pressure of 57 kPa and 0 kPa is ∼1.3059 and 1.3 respectively.

Hence the maximum compression ratio for the NA SI engines when it works with methanol is calculated as follows:

CR_NA= ( P_end

P_ini)

1/γ_NA

= (

P_boost×CR_boost^γboost

P_ini )

In conclusion, if the maximum pressure is limited to 100 bar, using methanol the engine could be further downsized by ∼10.7%. If the engine hardware would be stronger, the downsizing factor could be further improved. However, the weight of the engine would then increase. For NA SI engines, the maximum compression ratio when the engine works with methanol (at λ = 1) is around 14:1.

In this calculation, the influence of CR on turbulent flame speed and residual mass

METHANOL AS A FUEL FOR A DIRECT-INJECTIONSIENGINE 37

fraction is ignored, assuming the flame speed is similar to the case which has an intake boost pressure of 57 kPa. To explain the improvement of BMEP by changing the boost pressure and valve timing, a more in-depth investigation into the operating cycle will be presented in the following sections.

S-VVT sweep tests

Figure 2.8 shows the in-cylinder pressure profiles in log(P)-log(V) diagram at different valve strategies. As is shown clearly, the peak in-cylinder pressure increases with a more prolonged valve overlap duration, especially with a positive valve overlap. There are two reasons for that change: lower residual gas mass fraction and a richer in-cylinder mixture. Both improve the burning velocity, so the pressure rises more sharply. As can be seen, the crank angle when the in-cylinder pressure reaches peak gets closer to TDC with a longer valve overlap period.

1 10 100

40 400

In-cylinder pressure (bar)

Cylinder volume (cm³)

Baseline S-VVT 10 S-VVT 25 S-VVT 40

1650 rpm, λ = 1 IT = 10 CAD bTDCf

Figure 2.8: In-cylinder pressure profiles in log(P)-log(V) diagram with different valve strategies. Test condition: WOT, 1650 rpm, λ = 1, IT = 10 CAD bTDCf, boost pressure of

34 kPa.

When the valve shifting angle further increases, the residual mass fraction decreases due to the scavenging resulting from positive valve overlap. The intake air pushes the burned gases out of the cylinder at the end of the exhaust stroke.

A part of the intake air follows the flow to the exhaust manifold. Therefore, the lambda sensor will show a value greater than 1. The injection duration is then controlled to be longer to increase the amount of fuel. Figure 2.9 presents the instantaneous exhaust lambda in the S-VVT 40 case from a TPA simulation. The exhaust lambda increases during the valve overlap because of the presence of fresh

air in the exhaust. The in-cylinder lambda should be rich, λin−cylinder = 0.95, to achieve the overall exhaust lambda of 1. In the S-VVT 25 case, the exhaust lambda also reaches a peak during valve overlap. However, only a small amount of fresh air is then present in the exhaust, so the overall exhaust lambda does not change significantly, remaining around 1. For the four cases, only the S-VVT 40 case has a significant rich combustion. This explains the reduction of BTE in that case, as shown in Figure 2.6.

0.9 1 1.1 1.2 1.3

0 180 360 540 720

Exhaust lambdaλ(-)

Crank angle (CAD aTDCf)

Exhaust Intake

λexhaust_stroke= 0.962

1650 rpm, WOT, S-VVT 40 λin-cylinder= 0.95

Figure 2.9: Instantaneous exhaust lambda in the case S-VVT 40, WOT, engine speed of 1650 rpm, overall exhaust lambda of 1.

Figure 2.10 presents the ignition delay ID, crank angle when 50% mass is burned (CA50), combustion duration CA10-75 and CA10-90, and residual gas mass fraction in the four cases. These parameters are derived from the TPA simulation in GT-Power. As can be seen, the residual gas fraction first increases and then decreases with the prolongation of the valve overlap period. ID, CA50, CA10-75 and CA10-90 decrease when the valve shifting angle increases from 0 to 40 CAD.

Although the residual fraction is higher in the S-VVT 10 case, a small reduction of ID, CA50, CA10-75 and CA10-90 can be observed when the valve shifting angle increases from 0 to 10 CAD. Higher mixture temperature can explain this behavior (see Figure 2.11). In the case of S-VVT 25, less residual gas is the main reason, whereas, in the S-VVT 40 case, the combination of a lower residual gas fraction and a richer combustion leads to the decrease of ID, CA50, and combustion duration.

The ignition timing is maintained at 10 CAD bTDC_fin this sweep test, this is the MBT timing of the baseline case. With different valve timings, the MBT timing will actually change. As shown in Figure 2.10, the CA50 is closer to the TDC with

METHANOL AS A FUEL FOR A DIRECT-INJECTIONSIENGINE 39

Valve shifting angle (CAD) Ignition delay

Figure 2.10: The ignition delay, CA50, CA10-75, CA10-90, and residual gas mass fraction versus valve shifting angle at 1650 rpm, IT = 10 CAD bTDCf.

a larger valve shifting angle, this means the combustion phasing is too advanced using the fixed ignition timing. A later spark timing is required to keep the CA50 at around 8 CAD aTDC_fas with the baseline case. Therefore, in these cases, the ignition timing is too advanced, compared to the MBT timing.

With a shorter combustion duration, the unburned mixture has less time for auto-ignition. However, it also leads to a sharp increase of pressure as well as unburned gas temperature, reducing the auto-ignition delay time of the mixture.

Therefore there is a trade-off of faster combustion versus knock. Figure 2.11 presents the pressure-unburned gas temperature trajectories for different valve strategies. The grey lines show the constant auto-ignition delay contours of the undiluted stoichiometric methanol-air mixture which is derived from a simulation with Li’s mechanism [70]. Because the amount of burned gas in the mixture is small, and the difference in residual mass fraction between the four cases is limited, the results from the simulation of an undiluted methanol-air mixture are used to present the auto-ignition delay contours for all situations. The “knock limit” line is the auto-ignition delay contour of 1.36 ms, which was calculated from equation 2.6 [81].

τRES=

13.5 × 1000

6 × RPM (2.6)

This is a very rough approach to define the knock limit in an SI engine. At 1650 rpm, the residence time is 1.36 ms. As can be seen, the end gas state is still far

0

500 600 700 800 900 1000

In-cylinder pressure (bar)

Unburned gas temperature (K) Baseline

Figure 2.11: In-cylinder pressure-unburned gas temperature trajectories for different valve strategies at 1650 rpm, IT = 10 CAD bTDCf. Grey lines: simulated auto-ignition delay

contours for stoichiometric methanol-air mixture.

from the “knock limit”. From this figure, the S-VVT 25 is the case which has the highest chance to experience knock. In the S-VVT 40 case, due to the cooling effect of the fuel evaporation (rich mixture), the peak in-cylinder temperature is lower than the S-VVT 25 and S-VVT 10 cases. These peak temperatures are still far from the knocking temperature, ∼955 K, for methanol (see Chapter 6).

Intake boost pressure sweep tests

Figure 2.12 presents the in-cylinder pressure curves with different boost pressures at 1650 rpm; the ignition timing is maintained at 10 CAD bTDC_f. With a higher intake boost pressure, the pressure is higher over the entire cycle. The crank angle when the in-cylinder pressure reaches the peak is not advanced as was the case in the variable valve timing sweep test. This means the combustion speed does not change much when the intake pressure increases.

Four combustion parameters, the ID, CA50, CA10-75, and CA10-90 are presented in Figure 2.13. As can be seen, there is an insignificant change in these parameters

Four combustion parameters, the ID, CA50, CA10-75, and CA10-90 are presented in Figure 2.13. As can be seen, there is an insignificant change in these parameters