Pre-versus post-compressor EGR feed - Exhaust gas recirculation (EGR) circuit design .1 EGR ci

SUPERIOR DOWNSIZING

5.3 Exhaust gas recirculation (EGR) circuit design .1 EGR circuit layout

5.3.2 Pre-versus post-compressor EGR feed

For a typical production single-stage turbocharged Di engine, post-compressor eGr routing may be preferred in many cases to reduce excess fuelling at high speed and load. This approach reduces issues with eGr dead volumes and compressor wheel durability. However, successful application is dependent on achieving good cylinder-to-cylinder eGr distribution, maintaining stable combustion and achieving sufficient EGR rate, which is not always readily feasible. simply increasing the back-pressure via some external means is usually unfavourable to fuel economy. A pre/post compressor switching feed may help alleviate such effects but could be very difficult to control on a transient basis and possibly requires a complex setup of eGr valve(s).

regardless of such problems, there has been little detailed published information on the possible thermodynamic differences of pre- and post- compressor eGr for gasoline applications. As such, a comparison has been undertaken at constant high engine speed and load (5500 rpm, 250 nm and 15.8 bar BMeP, BlD – 1.5°). The pre-compressor circuit used was based on the hybrid system first shown in Fig. 5.2. A schematic of the post- compressor gas routing used is set out in Fig. 5.7, with the eGr returned post-intercooler rather than post-intake throttle. This was not ideal but avoided eGr distribution issues. The location of the pre-turbine eGr gas pick-up was similar in each case and is illustrated in Fig. 5.8. Observing Fig. 5.8(b), it was necessary to machine down the original exhaust runner splitter plates to maximise the eGr rate. The fact that the eGr gases were taken before the turbine in either case means that the current comparison cannot identify any benefits of a low-pressure post-turbine EGR pick-up. However, differences in compressor performance would still be identifiable and this comparison was therefore considered to be an appropriate first step.

A dual gas-to-water eGr cooler assembly was employed in both circuits,

where the first cooler was of stainless steel construction and was used to remove the bulk of the heat, and the second was of more conventional multi- tube design and was used to achieve the required eGr temperature set point.

The intake plenum gas temperature had again been closely controlled to 40°C via the intercooler and eGr coolers. All other test conditions remained as described previously at this site (see section 5.2.2).

shown in Fig. 5.9 are the key test results. initially, the engine was operated with the exhaust gas temperature at ~990°C, under slightly rich fuel conditions. The maximum post-compressor eGr rate was ~8%. At this eGr level, both systems could be operated with a stoichiometric fuel–air

EGR valve Intercooler

Throttle Main cooler

Pre-cooler

Comp.

Turb.

5.7 Schematic showing the high-pressure EGR circuit used for the experiments.

EGR extraction Cylinder 2

Cylinder 1

Cylinder 3

Cylinder 4

(a) (b)

5.8 (a) Original and (b) modified exhaust plenum showing the machining of the exhaust runner splitter plates.

Pre-compressor Post-compressor

Pre-turbine gas temp. (°C) Intake throttle duty cycle (%)Boost duty cycle (%)APmax (°aTDC)COV of IMEPg (%)NOx (ppm)

Relative AFR (l)BSFC (g/kWh)Intake plenum pressure (kPa abs)Manifold pressure ratio 1000

980 960 940 920

100 90 80 70 60 50 40 30 10090 8070 6050 4030 2010 –100

4 3 2 1 0

3500 3000 2500 2000 1500 1000 1.101.08

1.061.04 1.021.00 0.980.96 0.940.92 0.90 266 264 262 260 258 256 254 252

180 170 160 150 140

1.30 1.25 1.20 1.15 1.10

0 2 4 6 8 10 12 14 EGR rate (%)

5.9 Key parameters during pre- and post-compressor EGR sweeps at elevated engine speed (5500 rpm and 250 Nm/15.8 bar BMEP).

mixture. However, the post-compressor route appeared to incur a slightly lower exhaust temperature (~10°C), marginally better fuel consumption (254 vs 257 g/kWh), reduced intake plenum pressure (154 vs 158 kPa), lower compressor work (~9.4 vs ~10.5 kW), decreased demand on the turbine and hence lower exhaust-to-inlet manifold pressure ratio. The total increase in intake plenum pressure with 8% eGr was 13 kPa (pre) or 9 kPa (post) respectively. Without eGr at this relatively moderate load site, the engine was still operating with a partially closed intake throttle. As eGr was added, this throttle was fully opened, then, for pre-compressor feed, the wastegate was partially closed.

Given the observed differences in manifold pressure ratio at 8% eGr, it is possible that an increased mass of internal eGr was trapped in the pre- compressor case. such marginal increase in internal eGr cannot be easily computed but is possibly evident in the results, with combustion phasing further advanced and better combustion stability in the post-compressor case. The fact that engine-out emissions of nOx were similar, despite an earlier angle of peak in-cylinder pressure in the post-compressor case, may also be indicative of the small differences in eGr rate incurred. The timing of the spark was 38° bTDC or 39° bTDC for pre- or post-compressor feed respectively. The engine-averaged peak in-cylinder pressure was 74.5 bar (pre) or 76.8 bar (post). The air charge temperature had again been closely controlled to 40°C.

Finally, shown in Fig. 5.10 are the loci of the eGr sweeps superimposed on

Pre-compressor Post-compressor Compressor efficiency

Compressor press ratio

2.00

1.75

1.50

1.25

0.100 0.125 0.150 0.175 0.200

Compressor mass flow (kg/s) 0.72

0.74

0.76

0.76 0.74 0.72 0.70 0.66 0.76

5.10 Loci of operation superimposed on the turbocharger compressor map.

the turbocharger compressor map. in the post-compressor case, the compressor pressure ratio remained around a constant value of ~1.74 due to the intake throttling at the start of each sweep (as the throttle was further opened the compressor pressure ratio would have dropped without eGr being added).

In summary, when sufficient EGR rate could be achieved, post-compressor supply appeared favourable at high speed and moderate load, allowing reduced compressor work and with higher compressor efficiency maintained under the conditions tested. Despite these observations, it is still likely that a revised match of compressor and turbine would be required to achieve the highest engine loads due to the additional boost requirement when using eGr with any EGR circuit design (as quantified over a mapped area later on). Further work is required at different sites to verify the observations of increased internal eGr using pre-compressor supply, but this observation arguably also seems logical given that passing the eGr through the compressor clearly required higher compressor (and hence turbine) work.

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