Experimental study of an automotive Diesel engine efficiency when running under stoichiometric conditions

Applied Energy 105 (2013) 116–124

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Experimental study of an automotive Diesel engine efficiency when running under stoichiometric conditions Xavier Tauzia ⇑, Alain Maiboom Ecole Centrale de Nantes, LUNAM Université, LHEEA UMR CNRS 6598, 1, rue de la Noë, BP 92101, 44321 Nantes Cedex 3, France

h i g h l i g h t s " A Diesel engine can run under Stoichiometric Diesel Combustion (SDC). " SDC increases BSFC between 5% and 10% as compared to conventional lean Diesel. " BSFC can be improved with a specific injection strategy and exhaust gas recirculation. " There is a trade-off between combustion efficiency and thermal efficiency.

a r t i c l e

i n f o

Article history: Received 11 September 2012 Received in revised form 27 November 2012 Accepted 17 December 2012 Available online 21 January 2013 Keywords: Stochiometric Diesel Combustion Engine efficiency Energy balance Combustion efficiency Indicated thermal efficiency Brake thermal efficiency

a b s t r a c t If run under stoichiometric air–fuel ratio a Diesel engine could use a simple three way catalyst for NOx after-treatment instead of complex and expensive devices. This concept of Stoichiometric Diesel Combustion (SDC) has been experimentally tested on a modern automotive Diesel engine. Injection strategy (injection pressure, phasing, with or without pilot, multi-injection) and intake strategy (exhaust gas recirculation rate, swirl level) have been studied, for three operating points. It appears that, for the best strategies, brake thermal efficiency drops between 5% and 10% as compared with conventional lean Diesel. For each operating point, this drop is analysed with energy balance charts, and combustion rate of heat release. In particular the evolutions of combustion efficiency and gross indicated thermal efficiency are studied and it appears that there are some trades-offs between these two parameters. The evolution of particulate emissions and exhaust temperature are also described and commented. Finally, these results are used to propose some hardware modifications to improve SDC engine efficiency. Ó 2013 Elsevier Ltd. All rights reserved.

1. Introduction Diesel engines are widely used for automotive propulsion, mainly in Europe where their higher efficiency (when compared to petrol engines) is appreciated to reduce vehicle fuel consumption and CO2 emissions. However the recent and up-coming regulations (such as Euro6 or Tier2, Bin 5 in the US [1]) require drastic decrease of pollutant emissions, in particular NOx emissions [2]. Some studies try to reduce pollutant emissions by incylinder methods [3–8] or by developing new Diesel combustion concepts such like Homogenous Charge Compression Ignition (HCCI) [9–11], but complex and expensive after-treatment devices seem to become unavoidable. Diesel Particulate Filter (DPF) is used to remove Particulate Matter (PM) from the exhaust gases, Diesel Oxidation Catalyst (DOC) treats carbon monoxide (CO) and

⇑ Corresponding author. Tel.: +33 2 40 37 68 80; fax: +33 2 40 37 25 56. E-mail addresses: [email protected] (X. Tauzia), alain.maiboom@ ec-nantes.fr (A. Maiboom). 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.12.034

unburnt hydrocarbons (HC), while Selective Catalytic Reduction (SCR) or Lean NOx Trapp (LNT) is often necessary to reduce NOx emissions. These equipments notably increase the cost of Diesel engines and may also increase its fuel consumption (during DPF regeneration for instance). A few years ago a new concept emerged: if a Diesel engine which traditionally runs under a lean equivalence ratio could run under stoichiometric equivalence ratio, then a cheap Three Way Catalyst (TWC) could be used to treat gaseous emissions. Obviously, reaching stoichiometric air fuel ratio has some potential drawbacks regarding brake thermal efficiency, in particular if the engine has to be throttled like traditional SI engine are. However, the higher compression ratio remains an advantage when compared to SI engine. Moreover pumping losses due to throttling can be avoided or at least limited by the use of EGR and eventually a Variable Valve Timing/Lift System (VVT/VVL) [12]. There are relatively few papers about SDC, which is also referred sometimes as Stoichiometric Compression Ignition (SCI). Some are relative to a heavy duty engine at maximum rated power [13,14], while most of them, produced by researchers from Engine Research Centre (University of Wisconsin-Madison) and Hanyang

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Nomenclature Subscripts 1 Before compressor 2 Post compressor/before intercooler 20 Post intercooler 200 Post HP EGR 3 Before VGT 4 Post VGT before TWC 5 Post TWC before DOC+DPF 6 Post DOC+DPF Acronyms 2SA 2-spray-angle BMEP Brake Mean Effective Pressure BSFC Brake specific fuel consumption BTE Brake thermal efficiency CA Crank angle CE Combustion efficiency CFD Computational fluid dynamics CLDC Conventional Lean Diesel Combustion DME dimethyl ether DOC Diesel Oxidation Catalyst DPF Diesel Particulate Filter EGR Exhaust Gas Recirculation GHN Group-Hole Nozzle

University, report experiments and CFD calculations carried out on automotive engines running at low load [15–24]. In all cases, a SFC increase is observed when a Diesel engine is run under SDC as compared with a conventional lean run. Winsor et al. [14] measured a 15% BSFC increase at full load. It is attributed to combustion efficiency dropping to 97% and to an increase of pumping losses. Lee et al. [15] indicate a 7% penalty on ISFC at low load. The various studies therefore focus on reducing the fuel consumption penalty. In most studies a single injection strategy is used. In that case, at low loads, an early Start of Injection (SOI) leads to highly premixed combustion, often referred as Premixed Charge Compression Ignition (PCCI) combustion. This results in lower Indicated Specific Fuel Consumption (ISFC) and better combustion efficiency than a later SOI enabling a standard Diesel combustion [15,16,19]. The PCCI combustion phasing, however, does not seem to have a significant effect on ISFC, because there is a trade-off between combustion phasing and combustion duration [15,16]. At full load, with a standard Diesel combustion, the optimal SOI regarding BSFC produces high levels of PM emissions. Advancing injection reduces these PM emissions, but then BSFC deteriorates [14]. CFD calculations by Kim et al [20] showed that injection phasing can also change the way the spray targets the piston. This, in turn, can modify local equivalence ratio and combustion efficiency. Split injection was also studied [16] but no significant improvement was observed, when compared to single injection. Apart from eventually non-optimal combustion phasing, the main reason for engine efficiency deterioration under SDC is the insufficient use of surrounding air by the fuel spray. Thus several attempts have been made to improve the air–fuel mixing process. The increase of injection pressure usually enhances air entrainment by the fuel spray [16] and was found to reduce ISFC at low loads [16,18,19]. However, at full load a BSFC increase was observed, due to fuel pump efficiency deterioration with higher injection pressure [14]. Increasing boost pressure (with EGR) has a positive effect on ISFC [16,18,19], attributed to a better air entrainment by the fuel

GITE HCCI HP HSDI IMEP ISFC LHV LNT LP NEDC NITE PCCI PM ROHR SCI SCR SDC SFC SOI TDC TWC VGT VVL VVT

Gross Indicated Thermal Efficiency Homogenous Charge Compression Ignition High pressure High speed direct injection Indicated mean effective pressure Indicated Specific Fuel Consumption Lower heating value Lean NOx Trapp Low Pressure New European driving cycle Net Indicated Thermal Efficiency Premixed Charge Compression Ignition Particulate Matter Rate Of Heat Release Stoichiometric Compression Ignition Selective Catalytic Reduction Stoichiometric Diesel Combustion Specific Fuel Consumption Start of Injection Top dead center Three Way Catalyst Variable Geometry Turbine Variable Valve Lift Variable Valve Timing

spray [16]. A 30% reduction of fuel consumption with EGR when compared to throttled intake is reported by [18]. Increasing the swirl ratio from 1.85 (baseline) to 3.3, however, tends to increase ISFC. This is explained by secondary flow structure induced by an increased swirl that could deteriorate the fuel–air mixing [16]. This means that the optimal swirl ratio under SDC is lower than the engine baseline value. Using KIVA CFD code, Park [19] showed that spray included angle and piston shape can have a great influence on ISFC. Kim et al. [18] experimentally tested a Group-Hole Nozzle (GHN) which improves air–fuel mixing and reduces ISFC when used with throttled intake but which has no significant effect when used in conjunction with EGR. Using CFD, Park and Reitz [21] studied the impact of a 2-spray-angle (2SA) GHN on fuel/air mixture for SDC. They showed that the 2SA GHN improves ISFC (reduction around 15%) in comparison with a standard nozzle. A parametric study on the two angles characterizing the nozzle indicates that the lowest ISFC is obtained when one group of spray targets the bowl while the other is split between the bowl and the squish region. In addition, the angle between the two groups of spray must be large enough to prevent plume-to-plume interaction. Recently, Cha et al. [22] showed that an engine fuelled with dimethyl ether (DME) can run under SDC with very small soot emissions (due to DME chemical structure). The fuel consumption penalty when compared to lean operation is around 10%, with a trade-off between combustion efficiency and combustion phasing. The penalty can be slightly reduced by adding 10% to 20% of ethanol to the DME, which increases the ignition delay. Concerning emissions, acceptable level of raw PM emissions (lying between 0.2 and 0.4 g/kW h) can be obtained with SDC. They could be managed by a Diesel Particulate Filter (DPF) for which regeneration could be continuous, due to the high exhaust temperature, according to [13]. NOx and CO raw emissions produced under SDC were also measured by several authors [14,15] and the values are such that they can be treated by a TWC, as experimentally demonstrated by Sung et al. [23] and Kim et al. [20].

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Although a lot of interesting information is available in previous papers, most studies were performed at low loads (IMEP around 5 bar), mainly with single injection, and on a research engine so that the evolutions of pumping work and mechanical efficiency are difficult to investigate. Moreover a detailed comparison between SDC and conventional Diesel combustion regarding engine efficiency is rarely proposed. Thus, the purpose of this paper is to experimentally compare the efficiency of an engine running under Stoichiometric Diesel Combustion (SDC) and under Conventional Lean Diesel Combustion (CLDC). SDC is studied on an automotive multiple cylinder engine and a broad range of load conditions is investigated, including high loads. An in-depth analysis of efficiency differences between SDC and CLDC is proposed. 2. Material and methods The experiments are conducted on a 2.0 L water-cooled HSDI Diesel engine which originally conforms to Euro 4 standards. It is equipped with a Variable Geometry Turbine (VGT) turbocharger, an intercooler, a Diesel Oxidising Catalyst (DOC) and a DPF (see Fig. 1). Engine specifications are given in Table 1. A TWC has been added downstream the turbine and upstream the DOC + DPF to verify its ability to treat NOx emissions. The standard high pressure (HP) Exhaust Gas Recirculation (EGR) system is supplemented by a Low Pressure (LP) EGR loop. The advantages of the LP configuration are the suppression of cylinder-to-cylinder EGR unequal distribution that can happen with HP EGR [25–27], and the possibility to reach lower temperatures and higher boost pressures [28,31]. On the contrary, for a given EGR rate, HP EGR may generate lower pumping losses [30,32], lower PM mass flow through the DPF (for a given PM concentration at the cylinder exhaust) [29] and if transient operations are considered LP EGR has a longer response time [27,32]. An independent water circuit through the LP EGR cooler is used, in order to control the temperature of recirculated exhaust gases before reintroduction. The LP EGR route is used for all tests in SDC configuration, in order to be able to reach higher EGR rates. The HP EGR route is used for engine tests under CLDC configuration, so as to scrupulously keep the engine in its standard configuration, even if the comparison between SDC and CLDC may appear unfair due to the different EGR routes. Though, when in-cylinder processes are considered, in particular combustion and heat losses through the cylinder walls, the key points regarding EGR are the EGR rate, its composition and its temperature. The fact that the exhaust gases are recirculated through a LP or HP route does not matter. HP EGR is cooled with the standard

Table 1 Engine specifications. Compression ratio Number of cylinders Number of valves per cylinder Bore Stroke Combustion chamber Injection system Maximum injection pressure Number of injection holes Injector nozzle diameter

18:1 4 4 85 mm 88 mm Re-entrant bowl-in-piston Common-rail piezoelectric 1600 bar 7 0.150 mm

heat exchanger which uses engine coolant. The EGR temperature is thus not controlled in this case. The EGR rate is defined as follows:

X EGR ð%Þ ¼ 100 

X CO2 X CO2

in

ð1Þ

ex

where XCO2_in and XCO2_ex are measured CO2 concentrations in the inlet and exhaust manifolds respectively. PM emissions are measured with an AVL 415S smoke-meter, while two gas analyzers are used to measure exhaust CO, CO2, O2 and UHC concentrations, as well as intake CO2. The Rate Of Heat Release (ROHR) is calculated from in-cylinder pressure measurements. Details about measurements can be found in [33,34]. On ROHR figures (Figs. 5, 7 and 9) the arrows indicate the SOI (eventually for pilot and main injection). They are based on injection rate measurements made on a specific test bench, while SOI values given in the tables correspond to the electric signal generated by the ECU. As usual, the actual injection process is slightly delayed from the electric signal. The combustion efficiency (CE) has been defined here as follows [35]:

CE ¼ 1 

P _ ex  i xi  LHVi m FFR  LHVf

ð2Þ

_ ex is the exhaust mass flow rate (sum of fuel flow rate and where m intake air mass flow rate), xi and LHVi are the concentration and the lower heating value of species i respectively, FFR is the fuel flow rate, and LHVf is the lower heating value of fuel. LHV values have been taken from [35] and are listed in Table 2. The Gross Indicated Thermal Efficiency (GITE) is defined by:

GITE ¼

W gross W gross ¼ Q comb LHVf  mf  CE

Fig. 1. Engine configuration.

ð3Þ

X. Tauzia, A. Maiboom / Applied Energy 105 (2013) 116–124

Fig. 5. Rate Of Heat Release for CLDC and SDC configuration – OP05.

Fig. 2. Simplified representation of pilot spray location as a function of pilot SOI.

Fig. 6. Energy balance for CLDC and SDC configuration – OP10.

Fig. 3. Trade-off between CE and GITE for EGR sweeps.

where mf is the total mass of fuel injected per cycle and Wgross is the gross indicated work per cycle calculated with measured incylinder pressure during the working loop (closed loop). GITE is

Fig. 4. Energy balance for CLDC and SDC configuration – OP05.

Fig. 7. Rate Of Heat Release for CLDC and two SDC configurations – OP10.

Fig. 8. Energy balance for CLDC and SDC configuration – OP15.

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Table 4 Maximum relative error. XEGR PM(g/h) CE GITE BTE

2.6% 2.3% 2.6% 2.8% 1.05%

Table 5 Operating points.

Fig. 9. Rate Of Heat Release for CLDC and SDC configuration – OP15.

Species

LHV (MJ/kg)

CO UHC (=fuel) PM

10.1 42.8 32.8

thus independent of pumping losses. The later are integrated in the Net Indicated Thermal Efficiency (NITE), which is defined by:

W net Wnet ¼ Q comb LHVf  mf  CE

ð4Þ

where Wnet is the net indicated work per cycle, calculated with measured in-cylinder pressure during the whole cycle.

W net ¼ W gross þ W pump

ð5Þ

where Wpump is the indicated work (usually negative) corresponding to the pumping loop. Pumping losses depend on engine settings (rotation speed, load, boost pressure and EGR rate, etc.) but also on back pressure in particular the DPF generated back pressure which increases with soot loading. Since the soot loading varies between the various measurements (it would be too tedious to regenerate the DPF before each measurement) a correction procedure is used. The pressure drop associated with an empty DPF has been measured for several volume flow rates through the DPF. Then a model was built, based on the following relationship proposed by [36]:

DPDPF

empty

¼ aQ þ bQ 2

OP05

OP10

OP15

Approx. BMEP (bar) Engine speed (rpm) Total fuel quantity (mg/str)

5 1450 18.38

10 2050 32.48

15 2050 46.75

DPDPF

Table 2 Lower heating values for combustion efficiency calculation [33].

NITE ¼

Operating point

ð6Þ

For each measurement, the actual pressure drop in the DPF is measured. The pressure drop due to DPF soot loading is calculated using the aforementioned model with:

loading

¼ DPDPF

measured

 DP DPF

ð7Þ

empty

It was also observed that the pressure drop across the TWC increases with time, probably due to PM fouling. As a consequence, for a future set-up it would be better to place the TWC downstream the DPF, provided that the exhaust gas temperature at this location is high enough for a good TWC efficiency. Since the initial TWC back pressure was low compared to the DPF back pressure, the choice was made to remove the TWC back pressure in pumping work calculations, in order to avoid to bias results with TWC fouling. The additional pumping work corresponding to the additional back pressure due to DPF soot loading and to TWC is removed from the actual (measured) pumping work, thus defining a corrected pumping work, which is used in Eq. (5) in lieu of the measured one. It must be noted that this correction is approximate, since the change in TWC and DPF back pressure also modifies the turbocharger behaviour, but this is believed to be a second order effect. The same correction is applied for brake work estimation, and then the brake thermal efficiency is calculated as follows:

BTE ¼

Pbrake _f LHVf  m

ð8Þ

where Pbrake is the brake power. The additional back pressure due to DPF soot loading is also removed from raw inlet turbine pressure measurements to calculate corrected values which are presented in Tables 6–8 (see Section 3). Some energy balances are performed throughout the paper to help analyse the results. They are established considering engine cylinders as the control volume. The chemical energy contained in the fuel is converted into heat by the combustion process. However, the combustion is incomplete, so that some chemical energy is available in the exhaust stream, contained in species such as CO, UHC or soot. This remaining energy appears on the graph as ‘‘combustion inefficiency’’ (see above section for calculation details).

Table 3 Relative measurement error. Instrument

Calibrated range

Accuracy

Relative error

Inlet gas temperature (k-type thermocouple) Inlet gas pressure (2 bar piezoresistive relative pressure sensor HCS Sensor Technics) Air mass flow (hot wire air flow meter) Fuel consumption (PIERBURG PLU 401/121) Smoke (AVL 415S) Inlet CO2 (CAPELEC CAP 3200) Exhaust CO2 (SIEMENS ULTRAMAT 23) Exhaust CO (CAPELEC CAP 3201) Exhaust UHC (CAPELEC CAP 3201) In-cylinder pressure (Kistler 6055BB)

0–1000 °C 0–2 bar 0–800 mg/str 0.05–23 kg/h 0–10 FSN 0–21% 0–20% 0–10% 0–20,000 ppm 0–200 bar

±1 °C (at engine inlet) ±5 mbar ±4 mg/str ±37 g/h ±0.1 FSN 0.5% ±0.1% 300 ppm 10 ppm ±0.5 bar

±0.75% ±0.25% 1% ±0.16% 2% 2.4% 1% 0.3% 0.05% 1%

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X. Tauzia, A. Maiboom / Applied Energy 105 (2013) 116–124 Table 6 Engine settings and measurements for CLDC and SDC – OP05.

Injection pressure (MPa) SOI pilot (CAdeg ATDC) SOI main (CAdeg ATDC) Pilot mass (mg/stroke) EGR rate (%) EGR loop P 002 (absolute mbar) T 002 (°C) CE (%) GITE (%) NITE (%) IMEP-pumping (bar) BTE (%) PM (g/h) Soot (FSN) k Dair (mg/stroke) T3 (°C) P3 (absolute mbar) VGT closure (%)

Table 8 Engine settings and measurements for CLDC and SDC – OP15.

CLDC

SDC-a: Best BTE & CE

SDC-b: Best GITE

68.4 7.43 3.4 1.46 26 HP 1050 68 99.6 40.5 39.7 0.13 32.7 6.2 3.2 1.32 365 456 1148 76

100 36 8 6 31 LP 997 36 96.3 38.0 37.2 0.13 29.3 6.8 3.9 1.00 265 362 1126 59

100 – 9 0 58.7 LP 1143 38 94 39.3 36.9 0.35 28.4 8.6 4.3 1.00 265 324 1505 95

Injection pressure (MPa) SOI pilot (CAdeg) SOI main (CAdeg) Pilot mass (mg/stroke) EGR rate (%) EGR loop P 002 (absolute mbar) T 002 (°C) CE (%) GITE (%) NITE (%) IMEP-pumping (bar) BTE (%) PM (g/h) Soot (FSN) k Dair (mg/stroke) T3 (°C) P3 (absolute mbar) VGT closure (%)

Some of the heat released during combustion is converted into work during the working loop of the cycle (compression and expansion strokes): this is the gross indicated work. Some of this work is consumed for the gas exchange process (exhaust and intake strokes): it is referred as ‘‘pumping loss’’ on the graph (calculations details are given in the previous section). Some work is also dissipated due to mechanical friction and to run engine accessories. It is referred as ‘‘mechanical loss’’. The rest of the work is available on the engine crankshaft. It is referred as ‘‘effective output’’ and is deduced from the torque measured on the test bench. Mechanical losses are obtained by removing effective output from the net indicated work. The heat which is not converted into work leaves the cylinders by two ways. First, some heat is dissipated through the cylinder wall during the working loop. This is referred as ‘‘cooling loss’’ and the cumulative value, Qwall, is estimated as follows:

Q wall ¼ Q gross  Q net

ð9Þ

where Qnet is the cumulative net heat release, obtained from incylinder pressure measurement and thermodynamic analysis, and Qgross is the cumulative combustion gross heat, calculated with: Table 7 Engine settings and measurements for CLDC and SDC – OP10.

Injection pressure (MPa) SOI pilot (CAdeg ATDC) SOI main (CAdeg ATDC) Pilot mass (mg/stroke) EGR rate (%) EGR loop P 002 (absolute mbar) T 002 (°C) CE (%) GITE (%) NITE (%) IMEP-pumping (bar) BTE (%) PM (g/h) Soot (FSN) k Dair (mg/stroke) T3 (°C) P3 (absolute mbar) VGT closure (%)

CLDC

SDC-a: Best BTE & CE

SDC-b: Best GITE

105.7 17 2 1.83 4.2 HP 1298 36 99.8 40.3 39.9 0.12 35.4 4.9 1.7 1.28 606 632 1485 61

150 36 8 6 12.3 LP 1207 55 98.0 38.7 38.2 0.16 33.5 21.3 4.3 1.00 474 587 1381 64

150 36 8 6 26.7 LP 1301 35 95.0 40.4 39.1 0.33 33.2 78.6 6.92 1.00 474 525 1586 74

Standard

SDC-a: Best BTE/GITE

SDC-b: Best CE

125 24 1.6 2.5 0 HP 1888 47 99.9 41.6 40.6 0.36 38.3 2 0.5 1.35 900 640 2417 84

150 40 8 4 6.5 LP 1545 45 94.6 39.1 38.2 0.35 34.1 35.7 4.5 1.00 673 663 1911 80

150 40 8 4 0 LP 781 41 95.3 38.3 37.8 0.19 33.8 16.83 3.4 1.00 673 676 1661 62

Q gross ¼ mf  LHVf  CE

ð10Þ

The remaining heat referred as ‘‘exhaust loss’’ is evacuated from the cylinder through the exhaust. This part is calculated so as to satisfy energy conservation. Table 3 sums up the measurement technique, calibrated range, accuracy and relative error of various instruments involved in the experiments for various parameters. Errors in experiments can arise from instrument conditions, calibration, environment, observation, reading and test planning. The accuracy of the experiments has to be validated with an error analysis. It is performed here using the differential method of propagating errors based on Taylor’ theorem [37]. It gives the maximum error u of a function f(x1, x2, . . ., xn) as follows:

uðf ðx1 ; x2 ; . . . ; xn ÞÞ ¼

qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi X ðci  uðxi ÞÞ2

ð11Þ

As a result, the maximum relative errors for XEGR, PM (g/h), Combustion efficiency, GITE, and BTE are given in Table 4. Three operating points, called OP05, OP10 and OP15 are studied, their respective Brake Mean Effective Pressure (BMEP) being approximately 5, 10, and 15 bar (Table 5). OP05 and OP10 are operating points such as those encountered in the European emissions test cycle (NEDC [1]), in the urban and extra urban part of the cycle respectively, for a typical passenger vehicle (equipped with a moderately downsized Diesel engine). OP15 is representative of operating points that could be encountered in the extra urban driving cycle in the NEDC test cycle with a significantly downsized Diesel engine. Such operating points are becoming particularly critical in terms of NOx and PM emissions. Currently, there is a clear tendency to further downsize automotive Diesel engines, and thus there is a need to reduce pollutant emissions at higher loads. The total injection quantities per cycle (sum of various injections in case of split injection) are held constant for each operating point. The changes in BMEP are thus the result of BTE changes when engine settings are modified. 3. Results and discussion 3.1. Improvement of engine efficiency under SDC The main parameters influencing engine efficiency and emission (PM) under SDC have been experimentally tested, in particular

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injection and intake strategies. The most outstanding results are summarized hereafter, additional details can be found in [34,38]. 3.1.1. Injection strategy CLDC injection strategy directly applied to SDC lead to unacceptable results. In particular PM emissions reach a very high level (Smoke above 8 FSN), hardly measurable with our apparatus. Moreover, BTE is low, due to poor combustion efficiency (For instance for OP10, BTE is below 29%). These results are due to a too low injection pressure and a too late injection which cause poor fuel/air mixing, and low in cylinder temperatures unfavourable to fuel and PM oxidation. Regarding injection, the most efficient strategy was found to be a combination of pilot and main injection. Single injection has been investigated too but exhibited poor CE at low loads (at least 2 points below the pilot+main strategy for OP05 and OP10). GITE was sometimes a little improved but not enough to compensate for CE deterioration. Thus BTE was found to be lower with single injection. The BTE reduction was accentuated by an increase in friction losses due to higher in-cylinder pressure and pressure rise rate (PRR) than with pilot injection. At higher loads (OP15) the PRR is such that it could damage the engine, thus single injection must be avoided. On the other hand, multi injection (i.e. splitting the pilot injection or main injection in two separate events) was also tested but was not able to improve neither CE nor GITE as compared with ‘‘pilot+main’’ injection strategy. For the latter, the influence of injection pressure, injection timing, dwell angle and mass repartition between pilot and main was investigated. It was found that increasing injection pressure enhances fuel–air mixing and greatly reduces PM emission. However, for too high an injection pressure wall wetting effects appear and deteriorate CE. Thus injection pressure must be adapted to the in-cylinder gas density, which increases with engine load or when EGR is added to the stochiometric air–fuel mixture. It was set to 100 MPa for OP05 and 150 MPa for OP10 and OP15 when EGR is used. Then the best injection strategy consists in phasing the pilot injection quite early (typically between 40 and 36 CAdeg ATDC) so that it reaches the squish region, while main injection targets the piston bowl (see Fig. 2). In these conditions, CE was improved by an increase of pilot mass from its standard value to 6 mg/stroke (OP5, OP10) or 4 mg/stroke (OP15). This strategy seems to maximise air utilisation, CE and to reduce PM emissions. However, the corresponding combustion is phased early in the cycle (around TDC) so that GITE is not optimal. Thus, there is a trade-off between CE/PM emission and GITE. In the SDC literature, although the global aim of maximizing air utilisation is the same, the strategy is usually different and relies on hardware modifications such as specific narrow angle [15,16,19] or 2SA injector nozzles [21,24] or specific pistons shapes. The usual injection strategy is then early single injection, which yields to PCCI combustion. It is thus limited to low loads. However, Lee et al. [16] showed that, in some circumstances, split injection can lead to ISFC slightly lower than single injection, provided that fuel repartition and injections phasing are adequate. This later result is consistent with the present study. 3.1.2. Intake strategy For the ‘‘pilot+main’’ injection strategy defined above, various intake conditions were tested: pure air and increasing EGR rates, corresponding to increasing boost pressures, the intake temperature being kept constant. It appears that EGR has a positive effect on GITE (the dilution reduces in-cylinder temperatures and consequently heat losses through the walls). However it also tends to deteriorate CE (the trade-off between CE and GITE is illustrated in Fig. 3), to increase pumping losses (the VGT closure must be increased to provide higher boost pressure) and PM emissions (for

two reasons: first the soot concentration (g/Nm3) in the exhaust increases, probably due to poor oxidation caused by temperature reduction; second, the exhaust mass flow rate through the DPF increases with EGR rate due to the LP EGR configuration). In the SDC literature, the best engine efficiencies are also obtained with large amounts of EGR [16,18,19], but EGR drawbacks and EGR optimal rates are seldom discussed. Finally, an increased swirl level (as compared with the standard swirl level) was tested, but negative effects were predominant: GITE deteriorates due to increased wall heat losses, pumping losses increase (reduced cylinder head permeability), CE is reduced while PM emissions increase (for the above mentioned pilot+main injection strategy the pilot/main interaction seems to be altered). It is important to note that these results are related to the swirl levels that were considered in this study and should not be generalized. In fact they show that the optimal swirl value for SDC is not greater but lower than the baseline value of this engine. Indeed, increasing the swirl level has both positive (mixing enhancement) and negative (as described above) effects on combustion, emissions and engine efficiency. As a consequence, an optimal swirl level always exists, but its value depends on many parameters, including combustion mode. 3.2. Comparison between SDC and conventional lean Diesel combustion From the various parametric tests briefly described in the previous section, the configurations corresponding to the best BTE (and eventually best CE or best GITE) were extracted and are now compared to the standard (Euro4) engine configuration, referred as Conventional Lean Diesel Combustion (CLDC). It should be noticed that the standard engine was calibrated with a DPF but does not include any NOx after-treatment device, so that NOx are controlled by in-cylinder countermeasures (EGR, combustion retardation). This means that engine standard settings (referred to as CLDC) may not be optimal regarding BTE. It is also necessary to point here that these ‘‘best’’ points for SDC configuration were defined after an analysis of the various parametric studies and are not the result of a mathematical optimization as usually performed for engine calibration. 3.2.1. OP05 Table 6 gathers engine settings and measured parameters (mainly efficiencies and emissions) for OP05 with three different configurations: CLDC and two different SDC configurations. SDC-a is the Stoichiometric Diesel Combustion configuration which exhibits the best BTE, which coincides in this case with the best CE. SDC-b corresponds to the best GITE. Figs. 4 and 5 show the energy heat balance and the Rate Of Heat Release (ROHR) for these three configurations. The BTE under SDC-a configuration is reduced to 29.3% as compared with 32.7% under CLDC. Several phenomena can explain the BTE deterioration. First, CE under SDC is reduced from 99.6% to 96.3%. The mixing between air and fuel becomes critical under stoichiometric conditions. With the ‘‘pilot+main’’ injection strategy, the pilot timing and mass are defined as to maximise air utilisation and consequently CE. But this results in a very early combustion, with maximum ROHR just after TDC (Fig. 5). This non-optimal combustion phasing has a direct negative effect on GITE which drops from 40.5% to 38.0%. It is also probably responsible for an increase in wall heat losses (Fig. 4), due to higher in-cylinder temperatures, which also contributes to GITE reduction. The use of a single injection strategy (SDC-b) allows reaching the best GITE under SDC conditions, quite close to CLDC configuration (39.3% against 40.5%). This is possible thanks to a better combustion phasing (Fig. 5, the maximum ROHR is around 7 CAdeg

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ATDC). Indeed without pilot injection the ignition delay is longer thus allowing more time for fuel–air mixing. However, CE is worse than with a pilot injection (94% against 96.3%) probably due to oxidation reaction freezing when the in-cylinder temperature becomes too low during the expansion stroke. As explained earlier the combustion phasing for the CLDC configuration is not optimal (delayed combustion to reduce NOx emission) but this is partially compensated by reduced wall heat losses that can be related to lower temperatures allowed by the retarded combustion. The advanced combustion and increased wall heat losses under SDC cause a reduction in energy rejection in the exhaust (Fig. 4) and a decrease in measured exhaust temperature (Table 6). This later may have a negative effect on after-treatment device efficiency. The drop in NITE under SDC-b mode is more pronounced than for GITE, because pumping losses raise. Indeed the VGT closure (and thus exhaust pressure) must be increased to compensate for the increase in compressor work (due to the use of LP EGR) and the aforementioned reduction in exhaust energy. Finally, the PM level under CLDC and SDC-a configurations are quite close. These results concerning OP05 are quite consistent with those obtained by Lee et al. [16] approximately at the same load level, on a research engine with a similar displacement. The fuel penalty when switching from CLDC to SDC is around 10%, while CE and NITE under SDC are slightly lower than in the present study, around 93% and 35% respectively. They are obtained with a single injection strategy and a PCCI combustion. However, contrary to the present study, Lee et al. [16] found a CE higher with single injection than with split injection, probably because their engine was specifically designed for PCCI operation.

3.2.2. OP10 Table 7 shows engine settings and measured parameters for OP10 for CLDC and two SDC configurations. SDC-a is the Stoichiometric Diesel Combustion configuration which exhibits the best BTE, which coincides in this case with the best CE. SDC-b corresponds to the best GITE. Figs. 6 and 7 show the energy heat balance and the Rate Of Heat Release (ROHR) for these three configurations. For OP10, BTE is reduced from 35.4% to 33.5% between CLDC and SDC-a. As for OP05, the CE reduction (from 99.8% (CLDC) to 98.0% (SDC-a) for OP10) is one cause of the BTE drop. The ‘‘best’’ injection strategy remains the same (early pilot+main). However, for OP10 the CE reduction is less pronounced than for OP05, maybe due to a slightly increased intake temperature (55 °C for OP10, SDC-a, instead of 36 °C for OP05, SDC-a). In particular, for OP10, in the SDC-a configuration the combustion of the pilot injection is earlier and more complete than in the SDC-b configuration (Fig. 7). For the latter, the injection strategy is identical but the intake charge is diluted with EGR and the intake temperature is lower (36 °C). Trends concerning combustion phasing are the same as for OP05. Combustion is too early regarding GITE which drops from 40.3% to 38.7% between CLDC and SDC-a. This earlier combustion is accompanied with a large increase in wall heat losses. Nevertheless, increasing the EGR rate (from 12.3% to 26.7%) can help reduce in-cylinder temperatures and consequently wall heat losses (Fig. 6): SDC-b configuration exhibits a GITE as high as the CLDC one, although wall heat losses are slightly increased. But the use of large amounts of EGR has several drawbacks, which explain that BTE is finally lower for SDC-b than for SDC-a configuration. The incylinder temperatures reduction has a negative effect on CE and PM emissions (probably due to poor oxidation and maybe also to an increase in equivalence ratio at the spray lift-off length). Besides, pumping losses are also increased because the higher boost pressure combined with a reduced energy rejection in the exhaust (Fig. 6) requires increasing the VGT closure.

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A large increase of PM emissions is observed for SDC-a, PM mass flow rate through the DPF being multiplied by four, and the situation is even worse for SDC-b (Table 7). The FSN increase is less pronounced (4.3 for SDC-a against 1.7 for CLDC) but the LP EGR route inherently increases the exhaust mass flow rate though the DPF as compared with the HP EGR route. Finally a large decrease in exhaust temperature can be observed, that can be attributed to earlier combustion phasing and increased dilution with EGR.

3.2.3. OP15 Table 8 shows engine settings and measured parameters for OP15 for CLDC and two SDC configurations. SDC-a is the Stoichiometric Diesel Combustion configuration which exhibits the best BTE, which coincides in this case with the best GITE. SDC-b corresponds to the best CE. Figs. 8 and 9 show the energy heat balance and the Rate Of Heat Release (ROHR) for these three configurations. BTE deterioration under SDC is more obvious for OP15 than for OP05 and OP10, with a drop from 38.3% (CLDC) to 34.1% (SDC). The CE reduction is also more pronounced, maybe because the relative increase of injection pressure is lower here, due to limits of the existing hardware. In the meantime the in-cylinder density is increased. Thus it might be more difficult for the pilot spray to reach the squish region. The increased pilot SOI and the reduced pilot mass (when compared to OP5 and OP10) seem to corroborate this hypothesis. Slightly better CE is reached without EGR (SDC-b configuration). In this configuration injection strategy is the same as in SDC-a configuration, and the ROHR is almost identical (Fig. 9). The negative effect of EGR on CE is thus clearly demonstrated. On the other hand EGR is useful for reducing wall heat losses as can be seen in Fig. 8. It should be noted here that EGR rate was limited for OP15 by the permeability of LP EGR loop. However, the high level of PM obtained for SDC-a configuration seems to indicate that a higher EGR rate may lead to unacceptable PM levels. Another consequence of EGR limitation is that boost pressure is lower than in CLDC configuration. This explains why pumping losses do not increase whereas it was the case for OP05 and OP10 in some configurations. Pumping losses are even decreased for SDC-b configuration, although the intake needs to be slightly throttled to reach stoichiometric A/F without EGR. Indeed in this case the VGT closure is low. Finally, for this operating point, exhaust temperature increases when switching to SDC mode, while energy rejection at the exhaust remains almost constant.

4. Conclusions The present experimental study, conducted on a standard automotive Diesel engine, has shown that running it under SDC leads to a relative reduction of BTE, lying between 5% and 10%. These results are roughly consistent with those presented in the bibliography. However they were obtained with a different injection strategy (‘‘pilot+main’’ instead of early single injection and PCCI combustion) and concern a much broader operating range (with BMEP up to 15 bar). Moreover, in the present study, the causes for this BTE deterioration have been deeply analysed. These causes are multiple and their relative weight depends on operating conditions. CE is always reduced when switching to SDC mode (from 2% to 5%). With the existing hardware (Injection device, piston shape) a way to limit the CE deterioration is to inject quite a large mass as a pilot injection early in the cycle so that it reaches the squish region, while main injection is injected later to reach the piston bowl. This seems to maximise air utilisation and consequently CE. But this injection strategy in turn has a negative effect on combustion phasing which is too early regarding GITE.

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There are also pros and cons concerning the use of EGR. The dilution decreases in-cylinder temperature and consequently wall heat losses, with a favourable effect on GITE. But EGR has a negative impact on CE, PM emissions which often become a limiting factor, and pumping losses. Most of these trades-off are specific to SDC. With CLDC, for instance CE is usually very close to unity, so that best BTE often coincides with best GITE and/or NITE. The next challenge for SDC will be to try to get rid of these trades-offs. Some strategies that have not been tested during the present study may help. For instance, single injection strategy seems promising at low loads because it allows a good combustion phasing with low PM emissions; but CE then needs improvements, that may be achieved with higher intake temperature. Another example concerns EGR. Only LP EGR was used during this study, but if it has some advantages, the LP route also has some drawbacks, concerning pumping losses or PM flow through the DPF. Thus a conjunct use of HP and LP EGR routes [32] may improve engine global efficiency. The present study also can help to define hardware modifications aiming at improving engine efficiency under SDC: – An increased maximum injection pressure would probably enhance the fuel–air mixing process. – A different piston shape, with a non-re-entrant, shallow, chamfered bowl may help the spray to reach all the air available in the combustion chamber and thus may allow reaching optimal CE while correctly phasing combustion. – The same result may be achieved with two angle injector nozzles [24]. Finally if BTE is improved under SDC conditions to the same level as in CLDC conditions some additional investigations will be necessary to validate the SDC concept. Points such as transient behaviour or application to full load and high speed operating points should be studied, and a cost comparison with CLDC should be performed. References [1] www.dieselnet.com/standards/. [2] Aithal SM. Modeling of NOx formation in diesel engines using finite-rate chemical kinetics. Appl Energy 2010;87(7):2256–65. [3] Arcakliog˘lu E, Çelıkten I. A diesel engine’s performance and exhaust emissions. Appl Energy 2005;80(1):11–22. [4] Pickett LM, Siebers DL. Soot in diesel fuel jets: effects of ambient temperature, ambient density, and injection pressure. Combust Flame 2004;138:114–35. [5] Tree DR, Svensson KI. Soot processes in compression ignition engines. Prog Energy Combust Sci 2006;33:272–309. [6] Prasad BVVSU, Sharma CS, Anand TNC, Ravikrishna RV. High swirl-inducing piston bowls in small diesel engines for emission reduction. Appl Energy 2011;88(7):2355–67. [7] Al-Hinti I, Samhouri M, Al-Ghandoor A, Sakhrieh A. The effect of boost pressure on the performance characteristics of a diesel engine: a neuro-fuzzy approach. Appl Energy 2009;86(1):113–21. [8] Giakoumis EG, Alafouzos AI. Study of diesel engine performance and emissions during a Transient Cycle applying an engine mapping-based methodology. Appl Energy 2010;87(4):1358–65. [9] Dec JE. Advanced compression-ignition engines—understanding the incylinder processes. Proc Combust Inst 2009;32(2):2727–42. [10] Gan S, Ng HK, Pang KM. Homogeneous Charge Compression Ignition (HCCI) combustion: implementation and effects on pollutants in direct injection diesel engines. Appl Energy 2011;88(3):559–67. [11] Yao M, Zheng Z, Liu H. Progress and recent trends in homogeneous charge compression ignition (HCCI) engines. Prog Energy Combust Sci 2009;35(5): 398–437.

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