Prechamber Optimization

In our efforts, the prechamber volume was kept constant at 1600 mm³ corresponding to 0.08%

of the combustion chamber clearance volume, and four different designs were tested. Table 2.12 presents the diameter, number of holes, and the surface area of each design, while a schematic representation is shown in Figure 2.41. Designs D1 and D3 have the same total nozzle cross-sectional area. Designs D2, D3, and D4 have the same configuration, except for the fact that hole sizes were different. Only few tests conducted with D4, three holes 2 mm diameter, but they provided a result similar to those attained by D3. Hence, the fourth design has been neglected in this evaluation.

Table 2.12. Prechamber geometries.

Design Number of holes Diameter of holes [mm] Total surface area [mm²]

D1 6 2 & 1 11.78

D2 3 1.6 6.03

D3 3 2.23 11.72

D4 3 2 9.42

Figure 2.41. Schematic representation of three prechamber geometries.

Even though that the optimum IT might be different for each ignition system or prechamber configuration, here a first comparison was made under fixed IT and λ to understand the benefit of spatial distribution and nozzle diameter. After that, the full analysis is shown for the most promising design. Figure 2.42 and Figure 2.43 present the pressure traces and rate of heat release for λ = 1.58 with two IT. It is evident that D1 which has six holes in staggered configuration gave

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the best performance, higher and faster pressure rise. The rate of heat release also provided the same conclusion with the most rapid combustion in D1 compared to D2 and D3 (Figure 2.43 illustrates faster rate of heat release with D1, shown in red color).

Design 2 which has three holes, 1.6 mm diameter, gave a slightly higher rate of heat release (maximum value) because of its optimum IT in that case. The full thermodynamic analysis of D1 revealed that the IT was advanced beyond the MBT timing as shown in Table 2.13. Alternatively, D2 at IT 15° CA BTDC and D3 at IT 14° CA BTDC attained the MBT timing according to the maximum acceleration of heat release criterion.

Table 2.13. Combustion parameters for the three prechamber designs at λ = 1.58.

Design D1 D2 D3

Maximum pressure [bar] 75.07 71.44 68.84

Location of Max. pressure [CA ATDC] 13.4° 14° 14.7°

Mass fraction burned 50% [CA ATDC] 8.5° 9.65° 10.6°

Maximum acceleration of heat release [CA ATDC] -3° 0 0

The cross-sectional area of D1 and D3 are the same, but the spatial distribution of flame jets (six compared to three) in D1 provided a shorter ignition delay and combustion duration (see Figure 2.44). Similarly, comparing D2 and D3 that has the same hole configuration but with different diameters reveals a definite advantage. D2 provided shorter combustion duration because the flame jets velocity increases as the nozzle diameter decreases, thereby, deeper penetration and better spatial distribution were accomplished. However, the ignition delay shown in Figure 2.44 increased slightly in D2 because of higher flow restriction. In that design, engine knock was detected for λ ≥ 1.65 after reaching the target engine load (10 bar BMEP). Hence, D2 was tested only at λ = 1.58, and the full range was evaluated for D1 and D3. In case of engine knock, the weakest spot is the focusing lens that cracks after few knocking cycles.

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Figure 2.42. Pressure traces of natural gas reciprocating engine equipped with several designs of PCLI at λ = 1.58 and 10 bar BMEP.

Figure 2.43. Rate of heat release of natural gas reciprocating engine equipped with several designs of PCLI at λ = 1.58 and 10 bar BMEP.

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In lean combustion mode, the spatial distribution of jets has a dominant effect by minimizing the flame travel path which in turn leads to faster combustion. Therefore, instead of advancing IT with lean mixtures, IT was retarded in PCLI relative to SI and 2P-LI systems. Ignition delay and combustion durations for the three geometries at  = 1.58 are shown in Figure 2.44. It is obvious that D1 provided a shorter ignition delay, the difference was approximately 1-1.5° CA compared with the other designs. The MFB 10% was almost constant with IT, however, the MFB 10-90%

decreases as the IT advanced because of a better charge utilization. As seen, the combustion durations are somewhat similar at λ = 1.58.

Figure 2.44. (A) Ignition delay and (B) combustion duration for the three prechamber geometries for λ = 1.58.

Subsequent tests comparing D1 and D3 (see Figure 2.45) showed that both ignition delay and combustion duration were lower with D1. It has been noticed that the benefit of D1 is more noticeable with a leaner mixture. For example, the difference in ignition delay between D1 and D3 at IT 13° CA BTDC for λ = 1.58 was only 0.65° CA while the difference increases to 1.7° CA at λ = 1.7. Similarly, at the same conditions, the difference in combustion duration was almost negligible for λ = 1.58 while it was 3.1° CA with λ = 1.7. This observation shows how crucial is

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the spatial distribution with very lean mixture combustion. Additionally, there was no incidence of autoignition in the case of D1 even with a very lean mixture (λ = 1.75). The superior performance of D1 is attributable to the fact that it uses 2 and 1 mm diameter holes in a staggered fashion, which in addition to the spatial separation, leads to temporally separated issuance of jets from the larger and smaller holes.

Figure 2.45. (A) Ignition delay and (B) combustion duration for D1 and D3.

Overall design D1 - prechamber volume 1600 mm³, 3 holes 2 mm Dia. and 3 holes 1 mm Dia.

at 45° to the igniter axis - was the optimal geometry that would offer low NOx emissions and at the same time avoid autoignition and soot formation. This prechamber geometry was used in PCLI for subsequent tests comparing the three ignition systems shown in the next section.

A wide range of IT 11-21° CA BTDC, and λ 1.54-1.75 were used for the PCLI D1 to evaluate its performance, and compare it with SI and 2P-LI. The trend of pressure traces and rate of heat release with IT was identical to the results presented earlier with SI and 2P-LI systems; a faster combustion with spark advancement. Figure 2.46 shows the rate of heat release for PCLI D1, λ = 1.58, and several IT. The maximum acceleration of heat release analysis revealed that IT 15° and 17° CA BTDC were advanced beyond the MBT timing. Even though the engine was stable,

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running at those conditions results in a lower brake thermal efficiency due to the negative work against the compression stroke. The reduction of brake thermal efficiency was observed clearly with λ = 1.54 where the value reduced from 36.13% to 35.95% by advancing the IT from 15° to 17° CA BTDC. It is worth mentioning that running the engine with PCLI at λ ≤ 1.58 generates high BSNOx emission, but the author carried out those tests in order to compare the PCLI with SI system at identical conditions IT and λ.

Figure 2.46. Rate of heat release of natural gas reciprocating engine equipped with PCLI D1 at λ = 1.58, several ignition timings, and 10 bar BMEP.

Since the cyclic variations were presented under fixed air to fuel ratio and several IT with SI and 2P-LI systems, the author intended to show the effect of λ for the PCLI system. This illustration does not only demonstrate the capability of PCLI system but also emphasize the effect of λ on cyclic variation Figure 2.47 compares the cyclic variation at IT 15° CA BTDC and a range of 1.58

≤ λ ≤ 1.7. Also, the COV_IMEP, COV_LOMP, and COV_MP are presented in Table 2.14.

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It is clear that under fixed IT all cyclic variation metrics increases by using leaner mixtures because of the reduction in turbulent burning velocity. The COV_IMEP was within the acceptable industrial value for all tests except at λ = 1.75 where a sudden jump of COV_IMEP was noticed as will be shown later. Similarly, the COV_MP increased but at a relatively higher rate. This is an inductive of COV_MP sensitivity to λ which has a drastic effect on combustion duration. The COV_LOMP was almost constant for 1.54 ≤ λ ≤ 1.7 (see Table 2.14), and it was very small compared to SI 38.6% and 2P-LI 8.4% at identical operation condition. Overall, PCLI system improved engine cyclic variation by distributing the ignition sites and minimizing the dependence of flame speed. This feature is crucial for stationary large bore natural gas engines because brake thermal efficiency can be improved by using a lean mixture if the cyclic variation can be maintained within acceptable values.

Table 2.14. Cyclic variation of PCLI D1 system at IT 15° CA BTDC and several λ.

λ 1.58 1.65 1.68 1.7

COV_IMEP [%] 2.21 2.7 3.01 3.7

COV_LOMP [%] 6.15 6.15 6.23 6.71

COV_MP [%] 4.95 6.51 6.75 7.52

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(A) Indicated mean effective pressure

(B) Location of maximum pressure

Figure 2.47. Cyclic variation of a single cylinder engine ignited with prechamber D1 at IT of 15° CA BTDC and several air to fuel ratios.

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The average pressure traces for 500 consecutive cycles are illustrated in Figure 2.48 at IT 15°

CA BTDC. The maximum pressure increases and shifts toward the TDC by utilizing richer mixtures due to the increase in combustion rate. However, at IT 15° CA BTDC, only λ ≥ 1.68 provides acceptable BSNOx emission with PCLI because of the rapid combustion process.

Mass fraction burned analysis is shown in Figure 2.49 under several IT and λ. One notices that the ignition delay was almost constant with spark advancement, only at λ = 1.75, the ignition delay was increasing with IT. Conversely, the ignition delay was increasing with λ; for instance, the MFB 10% increased 1.35° CA by increasing λ from 1.68 to 1.7 at IT 15° CA BTDC. The general trend for all data points was an increase in the range of 0.8-1.8° CA in between the considered λ.

Only at λ = 1.75, the increase was 3.65° CA compared to λ = 1.7 at IT 17° CA BTDC. This sudden rise is attributed to the high cyclic variation encountered during those tests which revealed the limit of PCLI system.

The trend of combustion duration with the excess air ratio was identical to the ignition delay.

The MFB 10-90% increased 2.8° CA by increasing λ from 1.68 to 1.7 at IT 15° CA BTDC.

However, the combustion duration decreases with IT as illustrated in Figure 2.49B. This is indicative of burning the charge in a smaller volume which in turn leads to a faster combustion process. For the cases at λ = 1.75 and IT 17-21° CA BTDC, no clear trend was observed due to high cyclic variation as stated earlier.

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Figure 2.48. Pressure traces of natural gas reciprocating engine equipped with PCLI D1 at IT 15° CA BTDC, several excess air ratio, and 10 bar BMEP.

A B

Figure 2.49. (A) Ignition delay and (B) combustion duration for PCLI D1.

Figure 2.50 shows BSNOx emission as well as COV_IMEP for all the data points taken by PCLI D1. The scale of y-axis was increased to accommodate the wide range of engine operating conditions taken with PCLI. The BSNOx emission increases while the COV_IMEP decreases with spark advancement. One notices that all data points taken with λ = 1.54 and 1.58 provide BSNOx

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emission higher than 1.34 g\kW-hr (EPA limit). Also for λ = 1.65, only one data point can be employed with PCLI at IT 11° CA BTDC.

Likewise, all data points taken with λ = 1.75 offers COV_IMEP higher than the industrial accepted value. The sudden rise of COV_IMEP was noticed for all ignition systems but at a different λ. For the case of PCLI system, the COV_IMEP increased from the range of 3.7-5.11%

at λ = 1.7 to 10.24-13.82% at λ = 1.75. Since those data points do not comply with EPA regulation and the industrial accepted COV value, they were disregarded. Therefore, the PCLI system operating range was limited to 1.65 ≤ λ ^≤ 1.7 and IT 11°-17° CA BTDC.

Figure 2.50. Brake specific NOx emission and COV_IMEP for PCLI system at several air to fuel ratio.