Gasoline port fuel injection on a heavy-duty diesel engine

Gasoline port fuel injection on a heavy-duty diesel engine

van den Berge, B., Leermakers, C. A. J., Luijten, C. C. M., Somers, L. M. T., Albrecht, B. A., & Goey, de, L. P. H. (2011). Gasoline port fuel injection on a heavy-duty diesel engine. In R. F. Cracknell (Ed.), proceedings of the 8th International Symposium: towards clean diesel engines (TCDE 2011) Shell Global Solutions.

Document status and date: Published: 01/01/2011

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Gasoline port fuel injection on a heavy-duty diesel engine

C.A.J. Leermakers

, B. van den Berge, C.C.M. Luijten, L.M.T. Somers, L.P.H. de Goey

www.combustion.tue.nl Abstract

To be able to adapt fuel reactivity to operating conditions during transient engine operation, one dedicated test cylinder of a modified six-cylinder 12.6 liter heavy duty DAF engine has been equipped with port fuel gasoline injection complementing the stand-alone diesel injection system, EGR circuit, and air compressor. For future use, this real time control of timing and rate of heat release, by varying the balance between direct injected diesel and port fuel injected gasoline, cycle by cycle. The current study presents the very first tests of this dual fuel system.

Introduction

Current diesel combustion technology results in Ni-tric Oxides (NOx) and soot levels much higher than the limits to be imposed by the forthcoming emissions legis-lation (EURO VI, US 2013). Consequently, after-treatment systems based generally on Selective Catalytic Reduc-tion (SCR, [1, 2]) and Diesel Particulate FiltraReduc-tion (DPF, [2]) technologies have to be used for the reduction of NOx and soot, respectively. The development of com-bustion technologies with intrinsically lower emissions of smoke and NOxcan minimize the aftertreatment sys-tem required, and thus reduce the related costs. Premixed Charge Compression Ignition (PCCI) is such a combus-tion concept that holds the promise of combining emis-sion levels of a spark-ignition engine with the efficiency of a compression-ignition engine [3, 4, 5, 6].

PCCI combustion is characterized by low temperature partially homogeneous combustion by using early injec-tions, large ignition delays and large quantities of Exhaust Gas Recirculation (EGR) [7]. The most common prob-lems of PCCI combustion are wall-wetting, resulting in high Unburned Hydro Carbon (UHC) emissions, and lack of control of combustion phasing, possibly resulting in high pressure rise rates [7, 8]. Tie Li and coworkers [8] found that for PCCI combustion at higher loads, the pro-motion of fuel-air mixing at relatively higher intake oxy-gen concentration is necessary. He proposes to use low reactivity fuels and low compression ratios to expand the operating load range of smokeless low temperature com-bustion. Kalghatgi and co-workers [3, 7, 9, 10] showed that low reactive fuels elongate the mixing time of fuel with air and can be used for PCCI combustion at higher compression ratios. To achieve this kind of PCCI com-bustion practical fuels were used by Kalghatgi, like gaso-line and diesel.

In the investigations mentioned above, fuel was pre-blended before injection. Compression ignition based on fuel reactivity depends heavily on load and therefore fuel reactivity should be controlled per cycle. Hanson and Reitz used port gasoline injection and early cycle, direct

∗_{Corresponding author: C.A.J.Leermakers@tue.nl}

Proceedings of the European Combustion Meeting 2011

injection of diesel fuel for in-cylinder fuel blending and combustion phasing control. Their engine experiments were conducted with a conventional common rail injec-tor and demonstrated control and versatility of dual fuel PCCI combustion with the proper fuel blend, injection timings and compression ratio [11, 12].

One of the most promising features of dual fuel PCCI is an increase in thermal efficiency. In almost all pub-lished investigations, thermal efficiencies of over 50% are stated. According to Reitz and co-workers [11, 12, 13, 14] high thermal efficiencies can be achieved due to low temperature combustion in combination with highly ac-curate combustion timing and therefore minimized heat losses. Yet, a more detailed explanation of these high ef-ficiencies remains to be established.

Specific Objectives

The dual-fuel PCCI concept thus efficiently reduces emissions of smoke and nitric oxides, while still match-ing or even exceedmatch-ing a typical compression-ignition en-gine’s efficiency. However, as the fundamentals of the concept are still largely not understood, the Eindhoven Combustion Technology group has started to investigate this concept both through modeling and through dedi-cated experiments. The dual fuel PCCI concept will be studied on a dedicated test cylinder of a modified six-cylinder 12.6 liter heavy duty DAF engine.

For these investigations, this test cylinder has been equipped with a port fuel gasoline injection complement-ing the stand-alone diesel injection system, EGR circuit, and air compressor. Timing and rate of heat release can be directly controlled by varying the balance between di-rect injected diesel and port fuel injected gasoline, cycle by cycle. The current study presents the very first tests of the system of this dual fuel system and some initial results of what will be possible in the near future.

Experimental Apparatus

The CYCLOPS is a dedicated engine test rig, see Ta-ble 1 and [5, 10], based on a DAF XE 355 C engine. Cylinders 4 through 6 operate under the stock engine con-trol unit and together with a Schenck W450 dynamometer they are merely used to control the crankshaft rotational speed of the test cylinder, i.e. cylinder 1.

Table 1: Cyclops specifications Base engine 6 cylinder HDDI diesel

Bore [mm] 130

Stroke [mm] 158

Compression ratio [-] 15:1

Once warmed up and operating at the desired engine speed, combustion phenomena and emission formation can be studied in the test cylinder. Apart for the mutual cam- and crankshaft and the lubrication and coolant cir-cuits, the test cylinder operates autonomously from the propelling cylinders.

Fed by an Atlas-Copco air compressor, the intake air pressure of the test cylinder can be boosted up to 5 bar. Non-firing cylinders 2 and 3 function as EGR pump cylin-ders, see Figure 1, the purpose of which is to generate ad-equate EGR flow, even at 5 bar charge pressure and recir-culation levels in excess of 70%. The EGR flow can be cooled both up- and downstream of the pump cylinders and several surge tanks and pressure relief valves have been included in the design.

M

Resato fuel pump HPU200-625-2

EGR surge tanks

Gasoline tank T 2 3 4 5 6 Schenck W450 Dynamometer Atlas Copco air compressor Air filter Exhaust Back pressure valve EGR cooler EGR valve Micromo"on flow meter Common rail Port fuel injector

Figure 1: Cyclops engine test rig schematic Fuel to cylinder 1 is provided by a Resato HPU200-625-2 pump, which can deliver a fuel pressure up to 4200 bar, but in this case 1500 bar is used. An accumulator placed near ( 0.2 m) the fuel injector is to mimic the vol-ume of a typical common rail. The prototype common rail injector used has a nozzle with 8 holes of 0.151 mm diameter with a cone angle of 153 degrees. Like for fresh air and EGR, a Micro Motion mass flow meter is used to measure the steady state fuel mass flow. Time resolved

fuel pressure and temperature are measured in the fuel line with a Kistler pressure sensor and the injector actua-tion current is measured with a clamp meter.

For the purpose of the current investigations, port fuel injection of gasoline is provided by a Vialle LPG tank with an internal fuel pump. In the fuel line a filter is placed and the flow through the gasoline fuel line is mea-sured with a Micromotion mass flow meter. The injector used is a Vialle28 LPG injector placed on the intake man-ifold with an angle of 120 degrees resulting in an injec-tion spray posiinjec-tioned on the intake valve for better vapor-ization of the gasoline. The gasoline injector is controlled by a MoTeC M400 engine management system. With this system injection timing and duration can be controlled, as is done with the direct diesel injection. The port fuel in-jection system pressure is set to ca. 3 bar by controlling the rotational speed of the fuel pump in the gasoline tank. For measuring gaseous exhaust emissions a Horiba Mexa 7100 DEGR emission measurement system is used and exhaust smoke level (in Filter Smoke Number or FSN units) is measured using an AVL 415 smoke-meter. All quasi steady state engine data, such as intake and exhaust pressures and temperatures, and oil and water tempera-ture are recorded by means of an in house data acquisi-tion system. Finally, a SMETEC Combi crank angle re-solved data acquisition system is used to record and pro-cess cylinder pressure (measured with an AVL GU12C pressure transducer), intake pressure, fuel pressure and temperature and injector current.

For more information on the design of the experimen-tal apparatus and measurement procedures, the reader is referred to [10].

Measurement matrix

In this first investigation conventional diesel combus-tion timings (Start of injeccombus-tion near TDC) are used to ig-nite a mixture of gasoline and air.

To investigate the potential of online control of fuel reac-tivity, first the maximum allowable mixing ratio of gaso-line via port fuel injection is determined. This is done at low loads when running in conventional diesel combus-tion mode. Start of gasoline injeccombus-tion is placed at intake valve closing (IVC, 153◦CA bTDC) to spray on the hot intake valve for better vaporization of the fuel.

The measurements are performed using a geometric com-pression ratio of 15:1 and the following conditions:

• Two loads: 4 and 6 bar IMEP. • Two EGR levels: 0% and 40% EGR. • Two intake pressure levels: 1 and 1.25 bar. • Injection timings: -5, -10, -15◦_{CA aTDC.}

• Replacement of diesel by port fuel injected gaso-line up to 40wt%.

All possible combinations are tested, taking engine hardware limitations and combustion quality targets into account.

Fuel Characteristics

For auto-ignition control based on fuel reactivity a low reactivity fuel is necessary to extend ignition delay. For this investigation a regular RON95 gasoline is used in combination with an EN590 diesel to create the desired mixture reactivity. By choosing these two practical fuels, the experiments can be highly relevant to future applica-tions. Table 2 shows the specifications of the two fuels.

Table 2: Fuel specifications

Properties Diesel EN590 Gasoline RON 95

T10 210◦C 65◦C T50 268.5◦C 115◦C T90 333.3◦C 185◦C Cetane Number 55.9 14.7 Density [kg/m3_{] 825 @300K 725 @300K} LHV [MJ/kg] 41.54 44.40 Experimental procedure

Prior to a measurement series, the engine speed is set to 1200 rpm and the engine is warmed up until lubrica-tion oil and coolant fluid are 90 and 80◦C respectively. All operating conditions are set to the desired values and the test cylinder is fired at a conventional timing to heat the combustion chamber and exhaust. Measurements are only started after the CO2content of the exhaust flow has stabilized.

For each target load, the fuel mass flow is determined which is necessary in the conventional combustion regime. This fuel mass flow is then kept constant, while varying start of actuation and other operating conditions. When diesel is replaced by gasoline, the mass fractions are com-puted accordingly. The calculated total injected fuel en-ergy is used together with the resulting IMEP to compute the thermal efficiency.

Data analysis and definitions

When comparing emission levels and fuel consump-tion for heavy duty engines, it is common practice to calculate these brake specific, i.e. with respect to the power output at the crankshaft. In this test setup, this is not possible. Therefore the IMEP as calculated from the in-cylinder pressure signal is used. To be able to evalu-ate the combustion performance also with different intake pressures and varying exhaust back pressure, in all results presented the gross IMEP has been used to calculate in-dicated fuel consumption and emissions.

CA5 is used to indicate start of combustion (SOC), because of its considerably higher stability compared to CA2. Analysis of logged injector actuation current and injection pressure data furthermore shows a constant 4 ◦_{CA lag between start of actuation (SOA) and start of} in-jection (SOI) at the engine speed set at this study. Given the assumptions made above the following definitions are used to characterize combustion. Here CD is used to de-scribe the average mixing time, while ignition dwell

rep-resents the separation between injection and combustion events.

• Start of Combustion = CA5 • Start of Injection = SOA + 4◦_CA

• Combustion Delay = CA50 - SOI • Ignition Dwell = SOC - EOI • Ignition Delay = SOC - SOI

Furthermore, a fixed injection duration is assumed to represent a constant fueling rate, for a certain fuel and fuel temperature. Specific emissions are computed from their respective concentrations by using molecular weights. Only NOxis treated as NO2, in line with European leg-islation. PM emission is computed from measured FSN, using an empirical correlation.

Results and Discussion

In the sections below, results will be given to compare different gasoline fractions and diesel injection timings. Combustion phasing

Using conventional combustion, start of combustion (here defined as CA5) is controlled by the start of injec-tion. As can be seen from Figure 2, where markers over-lap, start of combustion is not depending of the fraction of gasoline injected via the intake port. This gasoline frac-tion does appear not to auto-ignite due to compression. Apart from CA5, also CA50, where half of the injected fuel has burned, stays constant. Only the last phase of combustion, from CA50 to CA90 is seen to be affected by the gasoline fraction. A larger amount of premixed fuel, is here seen to result in a shorter burn duration. Note that the amount of gasoline injected only 30%, so diesel combustion is dominant resulting in high temperature tur-bulent combustion. 4 6 8 10 12 14 16 18 A 9 0 [ d e g C A a T D C ]

4 bar IMEP, Pin 1.0, no EGR

CA90 Diesel CA90 10% gas CA90 20% gas CA90 30% gas CA5 Diesel CA5 10% gas CA5 20% Gas CA5 30% gas -6 -4 -2 0 2 -20 -15 -10 -5 0 C A 5 a n d C A

SOA [deg CA aTDC]

Figure 2: CA5 and CA90 vs SOA

Performance

The reason for implementing the port fuel injection system is to control the auto-ignition based on fuel re-activity. Here, the system is used in conventional diesel combustion mode, where the diesel injection is used to control start of combustion. In this mode, replacing di-rectly injected diesel for gasoline may cause a worse ef-ficiency, rather than a better one.

At 4 bar IMEP the total injected fuel mass is very low. When for example 20 wt% of the total fuel mass is re-placed by gasoline, creating a premixed mixture before start of combustion, in cylinder air-fuel mixture becomes very lean. When this mixture becomes too lean it may not reach the lower flammability limits of gasoline, and will not burn completely. Secondly, unburned hydrocar-bons, certainly with premixed air fuel mixtures, can get stuck into the crevices of the combustion chamber during compression stroke, which affects the efficiency. In fig-ure 3 it is shown that fuel consumption increases when mixing ratio of gasoline is increased. As mentioned be-fore, incomplete combustion due to lean mixture ratios of the premixed gasoline is probably responsible for this.

280 285 290 295 300 305 r o s s IS F C [ g / k W h ]

4 bar IMEP, Pin 1.0, no EGR

265 270 275 280 -5 0 5 10 G

CA50 [deg CA aTDC]

Diesel 10% gas 20% gas 30% gas

Figure 3: ISFC vs CA50

Emissions

The higher mixing ratio of gasoline not only affects combustion phasing and performance; due to the leaner mixture of gasoline and air, emissions undergo changes too. According to the Shell research group [15] the up-per flammability limit of gasoline is approximately twice the stoichiometric air/fuel ratio. In this investigation, the air fuel ratio of the premixed gasoline and air mixture is higher than this. It is therefore expected that certain changes are noticed when more diesel is replaced by port gasoline injection.

Figure 4 shows that both CO and HC emissions in-crease due to later injection timing and the inin-crease of gasoline addition. In the case of more premixed gasoline, more incomplete combustion occurs due to lean mixtures. Preliminary results have already shown, that higher gaso-line fractions are needed to prevent this overleaning. Also,

15 20 25 30 n d I S H C [ g / k W h ]

4 bar IMEP, Pin 1.0, no EGR

CO Diesel HC Diesel CO 10% gas HC 10% gas CO 20% gas HC 20% Gas CO 30% gas HC 30% gas 0 5 10 -20 -15 -10 -5 0 IS C O a n

SOA [deg CA aTDC]

Figure 4: ISCO and ISHC vs SOA

unburned hydrocarbons can get stuck into inlet manifold or crevices of the combustion chamber. A later diesel injection, more near TDC, can lead to a combustion oc-curring too far in the expansion stroke, at lower temper-atures. While at these lower temperatures combustion is less complete, also less NOxis produced, which can be seen in Figure 5. Slightly more NOxis produced when the mixing ratio of gasoline is increased but this can be considered negligible. At the conditions under investiga-tion, NOxlevels are generally quite high. The dual fuel operation is expected to suppress these when changing from a conventional diesel mode to PCCI, which is the subject of current investigations.

15 20 25 30 35 40 m is s io n s [ g / k W h ]

4 bar IMEP, Pin 1.0, no EGR

0 5 10 15 -20 -15 -10 -5 0 IS N O x e m

SOA [deg CA aTDC]

Diesel 10% gas 20% gas 30% gas

Figure 5: ISNOx emissions vs SOA

PM emissions are not shown here, because at the low load under investigation no significant amount of soot is produced. It is expected that the premixed gasoline does not produce any soot, even at higher loads.

Pressure rise rate

In Figure 6 it is shown that CA50 is of great influence on MPRR. In the second phase of combustion, a higher mixing ratio of gasoline results in a shorter burn dura-tion. This in its turn, results in an increase of pressure rise per crank angle. Note that when even more gasoline is added, in combination with the compression ratio of 15:1, the risk exists that the premixed gasoline with air forms a mixture which might auto-ignite. When this oc-curs, control of combustion phasing is lost and pressure rise rates can increase dramatically.

20 25 30 35 40 45 50 R [ b a r / d e g C A ]

4 bar IMEP, Pin 1.0, no EGR

Diesel 0 5 10 15 20 -5 0 5 10 M a x P R R

CA50 [deg CA aTDC]

10% gas 20% gas 30% gas

Figure 6: MPRR vs CA50

Timing of port fuel injection

When using a port fuel injection system the timing of injection may influence the vaporization of the injected fuel. Generally, in gasoline engines, fuel is sprayed onto the hot intake valve at cold starts and transient condi-tions [16, 17]. In this case all parameters are constant when a measurement series is started. In this measure-ment series, for 6 bar IMEP, two different positions of injection timing are investigated: start of injection while intake valve is open (IVO), at 300 ◦bTDC, or when in-take valve is closed (IVC), at 140◦CA bTDC. Influences of different injection timings of the gasoline injector are small on combustion phasing, however, unburned hydro-carbons are decreased. In Figure 7 one can see that when injection is timed at IVO, HC emissions decrease. It is be-lieved that, especially at low loads, the gasoline vaporizes better after injection with the help of swirl and in-cylinder conditions.

Conclusions

To be able to adapt fuel reactivity to operating condi-tions during transient engine operation, one dedicated test cylinder of a modified six-cylinder 12.6 liter heavy duty DAF engine has been equipped with port fuel gasoline injection complementing the stand-alone diesel injection system, EGR circuit, and air compressor. For future use, this real time control of timing and rate of heat release, by varying the balance between direct injected diesel and

1,5 2 2,5 3 3,5 S H C [ g / k W h ]

6 bar IMEP, Pin 1.25, no EGR, SOA -10

IVC, SOI 140 bTDC IVO, SOI 300 bTDC 0 0,5 1 0% 10% 20% 30% 40% IS Gasoline frac!on [%]

Figure 7: ISHC vs Gasoline fraction

port fuel injected gasoline, cycle by cycle. The current study presented the very first tests of this dual fuel sys-tem, where a conventional diesel injection strategy has been combined with different fractions of port fuel in-jected gasoline.

As expected, because of the premixed gasoline frac-tion, combustion is generally faster. This would give pos-sibilities for efficiency gains, although these are limited by incomplete combustion. The relatively small fraction of gasoline used in these experiments is premixed to a lean homogeneous mixture which may be outside its flamma-bility limits. Due to this incomplete combustion HC and CO emissions are raised. Therefore, to make this concept more viable, the premixed fraction should be less lean, which can be done in two ways.

First, the amount of gasoline injected could be in-creased. One should however pay attention that auto-ignition limits of the premixed gasoline are not exceeded. Furthermore, there is the possibility of multiple diesel in-jections. This could consist of an early injection of diesel, to stratify fuel reactivity of the mixture and to enrich this global mixture, prior to a second diesel injection near top dead center. This can make combustion more complete and more efficient. Both options are currently under in-vestigation. Multi-zone models will in the near future be applied to scan the possible conditions of interest. Acknowledgements

This project is funded by the Dutch Technology Foun-dation STW (is the applied science division of NWO) and the Technology Programme of the Ministry of Eco-nomic Affairs. DAF Trucks N.V., Shell Global Solu-tions, Avantium Technologies B.V. and Delphi are also acknowledged for their contributions to the project. The authors kindly appreciate the support of the technicians of the Eindhoven Combustion Technology group: Bart van Pinxten, Hans van Griensven and Theo de Groot.

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