Engine characteristics

Figure 7.7 shows the molar fraction of key species of the unburned gas in the combustion chamber as a function of EGR ratio for both conventional EGR and R-EGR cases. As can clearly be seen, the concentration of H₂and CO is very small in the conventional EGR cases (top graph). The molar fraction of methanol decreases and molar concentration of CO₂and H₂O increases as EGR ratio increases. Due to the presence of in-cylinder residuals, the molar fraction of CO₂and H₂O does not equal 0 at an EGR ratio of 0%.

Figure 7.7: Molar fraction of in-cylinder reactants as a function of EGR ratio.

For the R-EGR cases (bottom graph), the molar fractions of H₂and CO do not equal 0 beyond 6.5% EGR rate. The molar fraction of H₂equals ∼0.05 for different EGR ratios. At an EGR ratio of 6.5%, the CO molar fraction equals ∼0.0075, higher than for the other cases due to a high CO selectivity (see Figure 7.6). The molar fraction of water at this EGR ratio is similar to the case without dilution (0% EGR), i.e. it is due to internal EGR. This means 100% of the water in the EGR loop was consumed by the steam reforming process. Compared to the case without dilution, the molar fraction of water slightly decreases due to the reduction of residual mass fraction with a larger throttle position.

ENGINE SIMULATION 163

Thanks to the presence of H₂in the reactants, the combustion efficiency increases.

Figure 7.8 illustrates the combustion efficiency for two cases against the EGR ratio. The combustion efficiency was predicted using energy left at EVO (from unburned fuel, H₂and CO) and the inlet fuel energy. The fraction of unburned fuel, H₂and CO was calculated using the equilibrium method developed by Olikara and Borman [213]. As can be seen, the combustion efficiency in the R-EGR cases is higher than in the conventional EGR cases, especially at high EGR ratio. Also thanks to the formation of H₂with fuel reforming, LBV increases in the R-EGR cases. This causes a difference in the MBT ignition timing between the two cases.

Figure 7.9 presents the ignition timing for the two cases as a function of EGR ratio. As can clearly be seen, the R-EGR cases have a later ignition timing than the conventional EGR cases.

98.5 99 99.5 100

0 5 10 15 20 25 30

Combustion efficiency (%)

EGR ratio (%) Conv. EGR R-EGR

Figure 7.8: Combustion efficiency.

With the presence of H₂, the LBV increases which leads to a change in the flame development period (CA0-10) and the combustion duration (CA10-90). Figure 7.10 shows the CA0-10 (top graph) and CA10-90 (bottom graph) as a function of EGR ratio for both conventional EGR and R-EGR cases. CA0-10 and CA10-90 of the R-EGR cases are shorter than the conventional EGR cases, especially the flame development period. This is due to the increase in LBV. There are two main reasons for the increase of LBV: the presence of H₂ and a later ignition timing (higher unburned gas temperature). In SI engines, the combustion is first initiated by a laminar flame before it is wrinkled by the in-cylinder turbulence to form a turbulent flame. Therefore, the impact of a difference in LBV on CA0-10 is considerable. The CA10-90 is strongly influenced by the total (turbulent plus laminar) flame speed.

-30 -25 -20 -15 -10 -5

0 5 10 15 20 25 30

Ignition timing (CAD aTDCf)

EGR ratio (%) Conv. EGR R-EGR

Figure 7.9: MBT ignition timing for two cases as a function of EGR ratio.

10 15 20 25 30

0 5 10 15 20 25 30

CA0-10 (CAD)

EGR ratio (%) CA0-10 = 26.5 CAD

15 20 25 30 35

0 5 10 15 20 25 30

CA10-90 (CAD)

EGR ratio (%) Conv. EGR

R-EGR

Figure 7.10: Comparison of flame development period (CA0-10) and combustion duration (CA10-90) between the conventional EGR and the R-EGR at a BMEP of 7 bar, 1500 rpm.

ENGINE SIMULATION 165

To define the combustion stability limit, a CA0-10 limit of 26.5 CAD was applied (see Chapter 2). This corresponds to 5% coefficient of variance of IMEP (CoV_imep). As shown in Figure 7.10, the EGR limit for the conventional EGR is

∼24% and ∼30% for the R-EGR (CA0-10 of 26.5 CAD at these EGR ratios). Note that the combustion becomes unstable not only due to the presence of external EGR gases but also because of the presence of internal residual. At this EGR ratio, the residual mass fraction equals ∼5% (in the conventional EGR cases). Thus the total burned gas mass fraction is ∼29% for the EGR case.

The relationship between gross ITE and BTE with the change of EGR ratio in the conventional EGR and the R-EGR cases is presented in Figure 7.11. As can be seen in the top graph, the BTE in the R-EGR cases is higher. The absolute difference in BTE between the two cases becomes larger with increased EGR ratio, except for the case with EGR ratio of 6.5%. At this EGR ratio, the CO selectivity is high (see Figure 7.6). Higher CO fraction leads to an increase in the engine efficiency. At higher EGR ratios, the absolute difference is small, around 0.3%.

At an EGR ratio of 25%, the difference is larger, ∼0.7%. As explained in section 4.2.2, the improvement increases as EGR increases due to a smaller reduction of molar expansion ratio. Furthermore, the BTE decreases with an EGR ratio > 25%

due to increased combustion duration. Thanks to a shorter combustion duration (see Figure 7.10), the BTE of the R-EGR cases keeps increasing continuously, so the absolute difference between the two cases will be larger at high dilution rates.

The bottom graph compares the gross ITE of the two cases. As concluded in Section 4.3, the difference in gross ITE between the two cases is trivial. At EGR ratios of 6.5% and 25%, the difference is larger. As explained previously, high CO selectivity is the main reason for the behavior at an EGR ratio of 6.5%. At 25%

EGR, the gross ITE decreases with conventional EGR dilution due to a significant enhancement of pumping work (decrease of absolute PMEP) or gross IMEP (see Figure 7.12). The reduction of gross IMEP is more obvious than the decrease of injected fuel. The reduction rate of injected fuel decreases as EGR ratio increases.

Therefore, the gross ITE decreases at 25% EGR. The increase of BTE is further attributed to the reduction of pumping work. The pumping work decreases as EGR ratio increases, so the absolute difference between gross ITE and BTE becomes smaller at high EGR ratios.

In the conventional EGR cases, the BTE increases by around 1.3% points with 25%

EGR. The R-EGR concept got a slightly higher efficiency versus the conventional EGR, with the absolute difference being larger at higher EGR ratios. Similar to the results of the Otto cycle efficiency calculation (see Figure 4.4), the efficiency increases little with fuel reforming (versus EGR diluted combustion) and the improvement is more obvious at a higher EGR ratios. This can be explained

32 33 34 35

0 5 10 15 20 25 30

BTE (%)

EGR ratio (%) Conv. EGR

R-EGR

38 39 40

0 5 10 15 20 25 30

Gross ITE (%)

EGR ratio (%)

Figure 7.11: The influence of EGR ratio on the gross indicated thermal efficiency and brake thermal efficiency of the conventional EGR and the R-EGR cases at a BMEP of 7

bar, 1500 rpm.

6.5 7 7.5 8 8.5

0 5 10 15 20 25 30

Mean effective pressure (bar)

EGR ratio (%)

BMEP IMEP Gross IMEP

FMEP PMEP

Figure 7.12: Mean effective pressures as a function of EGR ratio for conventional EGR cases.

ENGINE SIMULATION 167

by a small enhancement of the reformate exergy compared to methanol and the reduction in the MER being less significant at high EGR ratios. Compared to the baseline (no dilution), BTE increases 4.04% with EGR and 6.15% with R-EGR at an EGR ratio of 25%. Because the EGR limit for the conventional EGR cases was estimated to be 24%, the efficiency improvement is slightly less. The estimated BTE at EGR ratio of 30% in the R-EGR case is ∼34.3% (at the dilution limit). The relative increase in BTE is 6.9% against the baseline, higher than ∼4%

improvement with the EGR dilution at the same combustion stability.

Figure 7.13 shows an example of the fuel energy distribution at an EGR ratio of 15% in the two cases, conventional EGR and R-EGR. The fuel energy is distributed in 6 parts: combustion loss, heat loss, exhaust loss, pumping loss, friction loss and brake work. The combustion loss represents the unreleased chemical energy in the exhaust gas at EVO (exhaust valve opening). The combustion loss is very small and the difference is almost invisible in the Figure. As in the previous prediction, a larger amount of heat is lost through the cylinder walls in the R-EGR cases. In this simulation, the heat loss increases from 14% to 15.1% with fuel reforming.

The absolute difference in the gross ITE of conventional EGR and R-EGR is very small, 0.1%. It is less than the difference in the BTE, which increases by ∼0.3%.

The absolute difference in friction loss is negligible. This means the improvement of BTE is mainly attributed to the reduction of pumping work. The trend and the absolute change of engine efficiency is similar to the findings in the previous analyses.

Figure 7.13: The fuel energy distribution of the conventional EGR and the R-EGR cases at an EGR ratio of 15%, BMEP of 7 bar, and 1500 rpm.

Figure 7.14 presents the in-cylinder pressure of the conventional EGR and R-EGR

cases at an EGR ratio of 15%. The pressure profiles were plotted in the log(P)-log(V) diagram, so the difference in the pumping work area can clearly be seen. The two cases have a similar exhaust pressure, however, the pressure during the intake stroke of the R-EGR is higher (larger throttle opening). Thus, the pumping work in the R-EGR cases is smaller. Higher intake pressure causes a higher pressure during the compression stroke, thus the compression work increases. Although the MER decreases with fuel reforming, thanks to a higher initial pressure (pre-combustion pressure) and a shorter combustion (see Figure 7.10), the peak pressure in the R-EGR case is slightly higher. This results in a marginal increase of friction work.

In order to further clarify the impact on burning velocities, Figure 7.15 presents the laminar and turbulent flame speeds for the conventional EGR and R-EGR cases at the same EGR ratio (15%). It is clear that the LBV in the R-EGR case is higher than for the conventional EGR case. Thanks to a higher LBV in the R-EGR cases, the flame development period (CA0-10) is shorter, thus the MBT ignition timing is later. Furthermore, the dilution limit was extended to ∼30% instead of 24% as for the conventional EGR dilution cases. Due to a late ignition timing in the R-EGR case, the turbulent burning velocity is slower in the beginning. Then, the turbulent burning velocity increases and reaches a peak before decreasing. The maximum turbulent burning velocity in the R-EGR case is identical to the conventional EGR case.

0.5 5 50

40 400

Cylinder pressure (bar)

Cylinder volume (cm³)

Conv. EGR R-EGR

Figure 7.14: Cylinder pressure in the conventional EGR and the R-EGR cases at an EGR ratio of 15%, BMEP of 7 bar, and 1500 rpm.

The turbulent flame speed depends strongly on the turbulent intensity (u^′) in the combustion chamber [45]. Figure 7.16 shows the comparison of in-cylinder