Gaseous and particulate matter emissions of biofuel blends in dual-injection compared to direct-injection and port injection

Applied Energy 105 (2013) 252–261

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Applied Energy journal homepage: www.elsevier.com/locate/apenergy

Gaseous and particulate matter emissions of biofuel blends in dual-injection compared to direct-injection and port injection Ritchie Daniel a, Hongming Xu a,b,⇑, Chongming Wang a, Dave Richardson c, Shijin Shuai b a

School of Mechanical Engineering, University of Birmingham, Birmingham, UK State Key Laboratory of Automotive and Energy, Tsinghua University, Beijing, China c Jaguar Advanced Powertrain Engineering, Coventry, UK b

h i g h l i g h t s " We investigate the gaseous and PM emissions of various DMF blends and compare them with the results of ethanol. " We compare dual injection with DI as baseline with reference to port injection. " Dual-injection effectively reduces mean PM diameter and gives rise to negligible accumulation mode. " It is identified that dual-injection has higher NOx emissions due to reduced charge-cooling.

a r t i c l e

i n f o

Article history: Received 3 April 2012 Received in revised form 30 August 2012 Accepted 7 November 2012 Available online 4 February 2013 Keywords: Dual-injection Direct-injection 2,5-Dimethylfuran Ethanol Particulate Matter

a b s t r a c t To meet the needs of fuel security and combat the growing concerns of CO2 emissions, the automotive industry is seeking solutions through biofuels. Traditionally, when supplying biofuel blends to the combustion chamber, the blend is mixed externally prior to its injection in one location. This location occurs either before the cylinder (port-fuel injection, PFI), or directly into the cylinder (direct-injection, DI). However, the use of dual-injection allows the in-cylinder blending of two fuels at any blend ratio, when combining the two locations (PFI and DI). This injection strategy offers increased flexibility as the blend ratio can be changed instantaneously according to engine speed and load demand and fuel availability. Previous work by the authors has reported the improved combustion performance of dual-injection with 25% blends (in gasoline) of a new biofuel candidate: 2,5-dimethylfuran (DMF). This current investigation extends the analysis to include the gaseous emissions of various DMF blends (25%, 50% and 75%) from 3.5 bar to 8.5 bar IMEP and the particulate matter (PM) emissions of similar fraction ethanol blends at a selected condition of 5.5 bar IMEP. Compared to DI, dual-injection offers reduced CO and CO2 emissions and comparable HC emissions. The mean PM diameter is decreased and the accumulation mode particles are negligible compared to DI. However, the implication of the higher combustion pressures is an increase in NOx due to reduced charge-cooling. Ó 2012 Elsevier Ltd. All rights reserved.

1. Introduction There is an increasing concern that the world’s hydrocarbon fuels might be depleted within the next 40 years if the current Abbreviations: aTDC, after top dead centre; bTDC, before top dead centre; CAD, crank angle degrees; CO, carbon monoxide; CO2, carbon dioxide; DISI, directinjection spark-ignition; DMF, 2,5-dimethylfuran; ETH, ethanol; HC, hydrocarbon; IMEP, indicated mean effective pressure; LHV, lower heating value; MBT, maximum brake torque; MFB, mass fraction burned; NOx, nitrogen oxides; PM, particulate matter; RPM, revolutions per minute; SI, spark-ignition; TDC, top dead centre; ULG, unleaded gasoline. ⇑ Corresponding author at: School of Mechanical Engineering, University of Birmingham, Birmingham, UK. E-mail address: [email protected] (H. Xu). 0306-2619/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.11.020

trend of energy use continues [1,2]. Therefore, it is important to search for alternative and sustainable energy sources in order to alleviate environmental concerns with respect to CO2 and confront the ballooning energy demand. For large energy consuming nations such as China and India there is an increasing gap between energy demand and supply. Therefore, biofuels can provide improved energy security and reduced dependency on imported oil. It is believed that the biofuel route offers the most viable mid-term supplement or substitute for gasoline, compared to technologies which are in their infancy (hydrogen fuel cells and full electric platforms) [3]. Currently, ethanol is the most widely adopted biofuel [4,5]. In 2007, ethanol accounted for 80% of the world’s total biofuel

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R. Daniel et al. / Applied Energy 105 (2013) 252–261

production [6]. However, a new biofuel candidate, 2,5-dimethylfuran or DMF (as it has become known [7]) is receiving increasing interest, due to the recent breakthrough of its manufacturing techniques [7,8]. DMF has a higher energy density than ethanol and it is also insoluble in water. Its superior advantages have spurred research into its fundamental combustion characteristics [9,10]. The first publications on the engine combustion and emissions characteristics of DMF were recently reported by the authors of this paper [11,12] and were followed up by the publication of a series of papers [13–17], including those from Xi’an Jiao Tong University [18–21]. This has encouraged others to investigate furan derivatives [22–25]. In the most recent work of [26] a dual-injection strategy has been researched for 25% blends of DMF (D25) in gasoline (by volume) and the results show competitive performance to the equivalent blend in DI and gasoline in homogenous DI. For instance, the fuel consumption was up to 3.2% lower than D25 blends in DI and up to 1.2% lower than gasoline in homogenous DI mode. The difference is attributed to the reduced piston wetting, improved mixture preparation and accelerated combustion rate. This current paper is a follow-up of the previous publication [26] for comparing the emissions of blends in dual-injection to DI. Traditionally, the method of fuel injection of blends has mirrored that with gasoline; the blend is either injected using either port fuel injection (PFI) or direct-injection (DI). However, the two modern injection methods (PFI and DI) have differing emissions benefits depending on the current operating mode [27]. Therefore, in order to minimize the emissions of gasoline–biofuel blends, the authors have investigated the possible leveraging of both injection systems simultaneously and directly compared the emissions of the equivalent blends in DI. The high potential of the dual-injection technology has been proposed and explored in [28]. Other researchers have examined the potential of ethanol (hydrous and anhydrous) boosted directand dual-injection engines to help cool the charge and suppress knock with only modest hardware modifications [29–31]. An exploitation of this technology has been found in the Ford ‘Ecoboost’ gasoline turbo-charged direct-injection (GTDI) engines.

Fuel Accumulator

Compressed Nitrogen Cylinder

Horiba MEXA7100DEGR Emissions Analyser (HC, CO, CO2, O2, NOx)

Pressure Gauge (150bar) to DI Injector

Exhaust VVT

Exhaust Plenum Chamber

Here, PFI gasoline and DI E85 (15% gasoline and 85% ethanol, by volume) has been used to improve the engine efficiency and therefore reducing CO2 emissions by suppressing knock at high loads [32]. Others such as Toyota and Audi have also demonstrated its advantages in achieving improved engine performance (fuel economy and torque) and reduced emissions for certain engine conditions [33,34]. Until now, the direct comparison of the emissions of dual-injection compared to DI has not been publicized. This paper compares the gaseous emissions of various blends of DMF in dual-injection to equivalent blends in DI as well as the particulate matter (PM) emissions when using ethanol. The experimental work is performed on a single cylinder, 4-stroke spark-ignition (SI) engine at 1500 rpm and k = 1 with loads varying from 3.5 bar to 8.5 bar IMEP. In the following sections, the engine setup, experimental results and finally conclusions are discussed. 2. Experimental setup 2.1. Engine and instrumentation The experiments were performed on a single-cylinder, 4-stroke SI research engine, as shown in Fig. 1 [19,26,28]. The 4-valve cylinder head includes the Jaguar spray-guided direct-injection (SGDI) technology used in their V8 production engine (AJ133) [35]. As well as firing under high pressure (150 bar) SGDI conditions, a low pressure (3 bar) PFI system is available. Early start of injection (SOI) timing (280° bTDC) was optimized for the highest volumetric efficiency and homogeneity [36], which occurs close to the location of the highest piston speed. The two fuelling modes (PFI using gasoline and DI) can be used independently or simultaneously. Although gasoline is the only fuel used in PFI, the DI fuel can be varied as required. The engine was coupled to a DC dynamometer to maintain a constant speed of 1500 rpm (±1 rpm) regardless of the engine torque output. The in-cylinder pressure was measured using a Kistler 6041A water-cooled pressure transducer which was fitted to the side-wall of the cylinder head. The signal was then passed to a

VAF Meter

Throttle

Intake VVT

PFI Injector

Fuel Cooler

Lambda Meter

100 Litre Intake Damper

Intake Plenum Chamber

Intake Filter

Pressure Regulator (3 bar) Fuel Pump

Fuel Pump (5 bar) Kistler Pressure Sensor Mass Balance

Crank Angle Encoder

Oil/Water Cooler Scanning Mobility Particle Sizer (SMPS)

Exhaust

Control Tower

Fig. 1. Schematic of engine and instrumentation setup.

High/Low Speed Data Acquisition

Fuel Tank

Air In

R. Daniel et al. / Applied Energy 105 (2013) 252–261

Kistler 5011 charge amplifier and finally to a National Instruments data acquisition card. Samples were taken at 0.5CAD intervals for 300 consecutive cycles, so that an average could be taken. The crankshaft position was measured using a digital shaft encoder mounted on the crankshaft. Coolant and oil temperatures were controlled at 85 °C and 95 °C (±3 °C) respectively using a Proportional Integral Differential (PID) controller. All temperatures were measured with K-type thermocouples. The engine was controlled using software developed in-house written in the LabVIEW programming environment. High-speed, crank-angle-resolved and low-speed, time-resolved data was also acquired using LabVIEW. This was then analyzed using MATLAB developed code so that an analysis of the combustion performance could be made.

100 DI Injector Calibration DI Injector Offset

Injection Mass (mg/cycle)

254

80

mgexp 60

mgcal Offset =

40

mgexp mgcal

20

2.2. Emissions and fuel measurement 0

The gaseous emissions were quantified using a Horiba MEXA7100DEGR gas tower. Exhaust samples were taken 0.3 m downstream of the exhaust valve and were pumped via a heated line (maintained at 191 °C) to the analyzer. Particulate matter (PM) emissions were measured using a 3936 Scanning Mobility Particle Sizer Spectrometer (SMPS) manufactured by TSI. This comprises of a 3080 Electrostatic Classifier, a 3775 Condensation Particle Counter (CPC) and a 3081 Differential Mobility Analyzer (DMA). PM samples were taken from the same position as the Horiba analyzer but measured asynchronously. A heated (150 °C) rotating disc diluter (Model 379020A, supplied by TSI) was used at a dilution ratio of 30:1. The SMPS measured particles from 7.23 to 294.3 nm in diameter and the sample and sheath flow rates were 1 and 10 L/min, respectively. The fuel consumption rate was calculated (and blend ratios controlled) using the volumetric air flow rate (measured by a Romet positive displacement rotary flow meter and stabilized by a 100 L intake plenum), known DI injector calibration curves for each fuel and by inferring the PFI fuel flow. All tests were run at stoichiometric conditions (k = 1), which was controlled using the cross-over of the carbon monoxide (CO) and oxygen (O2) emissions concentrations (verified using a lambda meter), as described in detail in a previous publication by the authors [19]. In order to calculate the blend ratio, the fuel flow rates for the PFI and DI injections are needed, which requires two assumptions. Firstly, the blending AFRstoich and LHV were assumed linear as the blend ratio is varied. Secondly, the actual DI injector mass flow rates were assumed using an offset to the calibration flow rates because, in an experimental situation, the local temperatures and pressures affect the fuel spray distribution and the mass flow will differ from the calibration. Therefore, an offset is required to the 100% DI case in order to accurately estimate the flow, as shown in Fig. 2. For instance, when comparing the actual DI flow rate of ethanol during combustion to the calibration, an offset exists. If this offset is applied to the lower flow rates, then the DI flow rate can be accurately assumed. Using the estimated DI flow rate, along with the volumetric air flow (VAF) rate and k (assumed to be 1 at the cross-over of the CO and O2 emissions concentrations), the gasoline fuel mass in PFI can be inferred using the relative AFR:

k¼

AFRactual AFRscotch to metric

ð1Þ

For a given gasoline-oxygenated fuel blend in dual-injection, this equation becomes:

 h

mf ;DI mf ;DI þmf ;PFI

ma mf ;DI þM f ;PFI

  AFRs;DI þ m

 mf ;PFI

f ;DI þmf ;PFI

 i AFRs;PFI

ð2aÞ

0

1

2

3

4

5

6

Injection Duration (ms) Fig. 2. SGDI injector calibration curve offset.

)k ¼

ma ðmf ;DI ÞAFRs;DI þ ðmf ;PFI ÞAFRs;PFI

ð2bÞ

In Eq. (2), ma and mf denote the mass of air and fuel (in PFI and DI), respectively. Eq. (2b) can then be simplified and re-arranged to obtain the mass of gasoline in PFI (mf,PFI), as shown in Eq. (3), assuming k is equal to 1 for the stoichiometric condition:

Mf ;PFI ¼

ma  ðmf ;DI ÞAFRs;DI AFRs;PFI

ð3Þ

It is now possible to calculate the fuel blend, as both PFI and DI components are known. The resulting fuel properties for the gasoline–biofuel blends in dual-injection were accurately calculated based on the final volume from each injector for each case. Therefore, linear interpolation between the biofuels and gasoline were then used to interpret the individual hydrogen-to-carbon (H/C), oxygen-to-carbon (O/C), stoichiometric air–fuel ratios (AFRstoich) and lower heating values (LHVs). The gasoline–biofuel blends in DI were externally mixed and measured accurately using a graduated measuring cylinder and burette. 2.3. Test fuels The DMF used in this study was supplied by Shijiazhuang Lida Chemical Co. Ltd., China at 99.8% purity. Both 97 RON gasoline and ethanol were supplied by Shell Global Solutions, UK. A high octane gasoline was chosen as this represents the most competitive characteristics offered by the market and provides a strong benchmark to the biofuels. The fuel characteristics are shown in Table 2. 2.4. Fuelling variations For clarity, the fuelling variations used in this work have been abbreviated in the remaining sections. Each gasoline–biofuel blend volume fraction is denoted by the biofuel content. For instance, for blends of 25% DMF in gasoline, the notation D25 is used. When using gasoline in PFI or DI it is referred to as PFI or GDI, respectively. When using DMF, in neat form in DI, the notation DDI has been used. Therefore, when referring to dual-injection, which comprises PFI and DDI, the notation G-DDI is used. However, the gasoline–biofuel blend fraction is subscripted after the DI fuel element in the previous notations. For instance, for D25 in DI, D25DI is used

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R. Daniel et al. / Applied Energy 105 (2013) 252–261 Table 1 Engine specification.

Table 3 Test variations.

Engine type

4-Stroke, 4-valve

PFI fuel

Combustion system Swept volume Bore  stroke Compression ratio Engine speed Injector PFI pressure and timing DI pressure and timing Intake valve opening Exhaust valve closing

Dual-injection: PFI and spray-guided DISI 565.6 cm3 90  89 mm 11.5:1 1500 rpm Multi-hole nozzle 3 bar, 50° bTDC 150 bar, 280° bTDC 16° bTDC 36° aTDC

Gasoline

Table 2 Test fuel properties.

Chemical formula H/C ratio O/C ratio Gravimetric oxygen content (%) Density @ 20 °C (kg/m3) Research Octane Number (RON) Motor Octane Number (MON) Octane index ([RON + MON]/2) Stoichiometric air fuel ratio LHV (MJ/kg) LHV (MJ/L) Carbon intensity (gCO2/MJ) Flash point (°C) Stoichiometric heat of vaporization (kJ/kgair) Initial boiling point (°C) Reid vapor pressure (kPa)

DMF

Ethanol

Gasoline

C6H8O 1.333 0.167 16.67 889.7a 101.3c 88.1c 94.7 10.72 32.89a 29.26a 83.6 1 31 92 3.45

C2H6O 3 0.5 34.78 790.9a 107b 89b 98 8.95 26.9a 21.3a 71 13 93.9 78.4 5.83

C2–C14 1.795 0 0 744.6 96.8 85.7 91.25 14.46 42.9 31.9 74.4 40 25.8 32.8 70.6

a

Measured at the University of Birmingham. Heywood, J.B., Internal Combustion Engine Fundamentals. 1988: McGraw-Hill [46]. c API Research Project 45 (1956) and Phillips data. b

and for dual-injection, G-D25DI is used. A summary of this information is found in Table 3. 2.5. Experimental procedure The engine was considered warm once the coolant and lubricating temperatures had stabilized at 85 °C and 95 °C, respectively. All the tests were carried out at the stoichiometric AFR (k = 1) using the cross-over of CO and O2 emissions concentrations, fixed injection timing (280° bTDC) and engine speed (1500 rpm), ambient air intake conditions (approximately 25 °C ±2 °C) and constant valve timing (see Table 1). The averaged in-cylinder pressure data (from 300 consecutive cycles) was then analyzed using MATLAB. Engine load sweeps (3.5–8.5 bar indicated mean effective pressure (IMEP) in 1 bar increments) were performed using gasoline in DI and PFI and DMF in DI and gasoline–DMF splash blends (25%, 50% and 75%, by volume) in direct- and dual-injection modes (gasoline in PFI, DMF in DI). At each load, the spark was advanced in order to find the minimum advance for best torque or MBT timing. If audible knock occurred, the timing was retarded by 2CAD, an arbitrarily safe margin and is denoted KL-MBT, or knock-limited MBT timing. For the particulate matter (PM) emissions testing using ethanol blends, the load was fixed at 5.5 bar IMEP. Repeats were performed for each test and an average was taken. When changing fuels, the high pressure fuelling system was purged using nitrogen until the lines were considered clean. Each blend was mixed vigorously prior to each test, in order to discourage phase separation (no fuel additives were available). Once the line was re-pressurized to 150 bar using the new fuel, the engine

DI fuel

Notation

Gasoline DMF

PFI GDI DDI

DI and dual-injection blends (25% DMF, by vol.) DMF/gasoline Gasoline DMF

D25DI G-D25DI

DI and dual-injection blends (25% ETH, by vol.) ETH/gasoline Gasoline ETH

E25DI G-E25DI

was run for several minutes. This made sure that no previous fuel remained on the injector tip or any combustion chamber crevices before any data was acquired. 3. Results and discussion 3.1. Gaseous emissions The engine-out emissions for the various gasoline–biofuel blends using dual-injection and DI between 3.5 bar and 8.5 bar IMEP are discussed. The legislated gaseous emissions are first evaluated, which include the hydrocarbon (HC), carbon monoxide (CO) and nitrous oxide (NOx) emissions. This is then followed by an analysis of the carbon dioxide (CO2) emissions, which is not a legislated combustion product but is one of the substances responsible for global temperature rises through the greenhouse effect. For each emission, two types of graphs are used. The first type compares the emissions of dual-injection with DI using D25 from 3.5 bar to 8.5 bar IMEP. The results for PFI, GDI and DDI have been added as references. The second graph type compares the normalized emissions, which is the ratio of the emissions with dual-injection to DI. Therefore, values above or below unity represent an increase or decrease in emissions with dual-injection, respectively. Only 3 blends and 6 loads have been tested, representing 18 data points for each injection mode. Therefore, in order to produce a 2D emissions map, the surrounding values have been interpolated. 3.1.1. Hydrocarbons The indicated specific hydrocarbons (isHC) emissions for each fuel combination and injection mode is shown in Fig. 3. As FID analyzers are reported to have a reduced sensitivity to oxygenated HC fuels [37,38] future work will include a detailed HC emissions investigation for more accurate quantification using Fourier Transform Infrared Spectroscopy (FTIR). However, in this study, the results in Fig. 3 provide a reference for determining differences between the dual-injection and DI modes with D25. In previous work, the isHC emissions of DDI have been shown to be lower than GDI from 3.5 bar to 8.5 bar IMEP [11] when using MBT timing. This also extends to PFI, as shown in Fig. 3. Therefore, it is expected that the use of D25DI and G-D25DI will show an increase in isHC emissions from the DDI case up to loads of 7 bar IMEP. The reduction in isHC emissions (increased oxidation of unburned fuel) for both modes (D25DI and G-D25DI) above 7 bar IMEP is believed to be due to the increased combustion temperatures with load. The oxygen molecule contained in the structure of DMF helps to increase the oxidation reaction as the oxygen is more readily available. Between D25DI and G-D25DI there is no discernible difference in isHC emissions. However, the rise in isHC emissions at low load above that with GDI for D25DI is surprising as the supplement of DMF and injection in DI should help in its reduction due to the oxygen content of DMF (see Table 2).

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R. Daniel et al. / Applied Energy 105 (2013) 252–261

1500rpm, λ = 1

8.0

1500rpm, λ = 1

11

D25

D25 D25DI G-D25DI DDI PFI GDI

isHC (g/kWh)

7.0 6.5

10 9

isNOx (g/kWh)

7.5

6.0 5.5

8 7 6

5.0

5

4.5

4

D25DI G-D25DI DDI PFI GDI

3

4.0 3

4

5

6

7

8

3

9

4

5

Fig. 3. Indicated specific HC emissions for D25DI and G-D25DI compared to DDI, GDI and PFI.

7

1.05

65

1.00

1.0

0.99 0.97 45

0.96

1.0

35

0.98

DMF Volume Fraction (%)

1.01

65

1.15

6.5

7.5

1.11

1.15

55

1.07 1.04 45

1.00

1.15 35

1.15

25 5.5

1.18

1.1

1.02

4.5

1.22

1.1

1.03

3.5

1.05

1.04

1.02

55

9

Normalised isNOx

75

1.0

8

Fig. 5. Indicated specific NOx emissions for D25DI and G-D25DI compared to DDI, GDI and PFI.

Normalised isHC

75

DMF Volume Fraction (%)

6

IMEP (bar)

IMEP (bar)

8.5

IMEP (bar) Fig. 4. Normalized indicated specific HC emissions for DxDI and G-DxDI.

The variation of isHC emissions with load and blend ratio between all dual-injection and DI blends is shown in Fig. 4. Although the difference in isHC emissions for all blends and load combinations is ±4%, the dual-injection mode offers more consistent reductions. For instance, the isHC emissions with D50 are always lower with dual-injection. At 3.5 bar IMEP, G-D50DI (6.8 g/kW h) is 4% lower than D50DI (7.1 g/kW h). This difference decreases as the load is increased but the isHC emissions are consistently lower with G-D50DI. This is likely to be due to the reduced piston impingement with dual-injection as the principle is similar to double pulse or split-injection [39,40]. The lower DI fraction in dualinjection is more likely to avoid impingement than injecting the whole blend in DI. Also, due to the lower LHV of DMF (see Table 2), the blend in DI would have a higher injection duration than GDI, so there is a greater risk of fuel impingement. Any such impingement reduces the fuel droplet vaporization and would result in more unburned fuel and higher HC emissions. At each corner of the emissions map in Fig. 4, the isHC emissions are lower for DI blends. Clearly, as the load is increased, the largest reductions in isHC

25 3.5

4.5

5.5

6.5

7.5

8.5

IMEP (bar) Fig. 6. Normalized indicated specific NOx emissions for DxDI and G-DxDI.

emissions require modifications in blend ratio. This is the advantage of the dual-injection mode as the blend ratio is flexible and can be varied instantaneously. 3.1.2. Nitrogen oxides When analyzing the indicated specific nitrous oxide emissions (isNOx), clear differences develop between the D25 blends in DI and dual-injection as shown in Fig. 5. The isNOx emissions of GD25DI are between 13% and 16% higher than D25DI as the load varies from 3.5 bar to 8.5 bar IMEP. This is due to the offset in the charge-cooling effect by the PFI fraction in dual-injection as opposed to GDI. At 8.5 bar IMEP, the isNOx emissions of D25DI is comparable to PFI and GDI. The addition of 25% DMF has very little effect on increasing the isNOx emissions. However, the isNOx emissions at this load are much higher when using G-D25DI (which is similar to DDI). The consequence of the gasoline fraction in PFI with dual-injection is the increase in combustion temperature [41] due to the loss of charge-cooling, which occurs effectively with DI. The early PFI injection helps to improve vaporization

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R. Daniel et al. / Applied Energy 105 (2013) 252–261

1500rpm, λ = 1

40

D25

PFI GDI

0.98

DMF Volume Fraction (%)

D25DI G-D25DI DDI

37 34

isCO (g/kWh)

Normalised isCO

75

31 28 25 22

1.00

65

0.95

0.8

0.90 0.85

55

0.80

0.9

0.75 45

0.70

0.8 35

0.8

19 25 16

3.5 3

4

5

6

7

8

4.5

9

5.5

6.5

7.5

8.5

IMEP (bar)

IMEP (bar) Fig. 8. Normalized indicated specific CO emissions for DxDI and G-DxDI. Fig. 7. Indicated specific CO emissions for D25DI and G-D25DI compared to DDI, GDI and PFI.

1500rpm, λ = 1

96.5

D25DI G-D25DI DDI

96.0

Combustion Efficiency (%)

[26] but cannot be used to help reduce the charge temperature. In previous studies by the authors, DDI has been shown to burn with much higher combustion temperatures than GDI [11]. Therefore, the combination of PFI and DDI results is even higher isNOx emissions. This is clearly the consequence of dual-injection with the case of 25% DMF. The difference in isNOx between the DMF blends in dual-injection and DI are shown for increasing loads in Fig. 6. Here, the isNOx emissions are consistently lower with DI blends. The largest increase in isNOx occurs with G-D50DI. At 5.5 bar IMEP, the isNOx emissions with G-D50DI are 21% higher than in D50DI. However, as the DMF fraction increases, the difference between the two injection modes decreases. For instance, with D75 blends, the increase with G-D75DI is less than 10% for all loads and at 6.5 bar IMEP is comparable to the D75DI result. This is mainly because the effectiveness of the charge-cooling in DI with the higher fraction of DMF starts to dominate the process. As the fraction of DI reduces the charge-cooling benefit decreases which leads to an increase in isNOx emissions.

D25

PFI GDI

95.5

95.0 94.5

94.0

93.5

93.0 3

4

5

6

7

8

9

IMEP (bar) Fig. 9. Combustion efficiency for D25DI and G-D25DI compared to DDI, GDI and PFI.

3.1.3. Carbon monoxide Conversely to the isNOx emissions, the indicated specific carbon monoxide (isCO) emissions are consistently lower when using D25 blends in dual-injection compared to DI, as shown in Fig. 7. With G-D25DI, the isCO is at least 11% lower than with D25DI. The greatest decrease of 9.4 g/kW h (31%) is seen at 5.5 bar IMEP. Dual-injection produces less isCO emissions than any other fuelling mode. For instance, when using DMF in dual-injection, the isCO emissions are at least 8 g/kW h (28%) lower than PFI, GDI and DDI at 8.5 bar IMEP. Above 5.5 bar IMEP, the use of PFI results in increased isCO emissions due to the onset of knock and need to retard the ignition timing. However, the PFI fraction, which vaporizes quickly, aids the fuel atomization and ensuing combustion of the biofuel fraction in DI [26]. Such positive effects have also been seen in compression ignition (CI) work when supplementing diesel in DI with gasoline in PFI [42,43]. The gasoline component in PFI aids the vaporization of the diesel fuel and increases the combustion efficiency at low loads [43]. This has also been seen with diesel–gasoline blends but to lesser extent [44,45]. The normalized isCO emissions for each fuel blend and load combination is shown in Fig. 8. Following the interpolation, an area of high efficiency (>20% decrease in isCO emissions) exists for

dual-injection fuel fractions below D60 and loads below 6 bar IMEP. However, with higher blends and loads, the difference between the isCO emissions reduces. At 5.5 bar and 6.5 bar IMEP, there is no advantage with G-D75DI because the isCO emissions increase by 4% and 6%, respectively. This is because of the reduced PFI fraction. However, reductions over 15% can be found with lower DMF fractions where the PFI fractions dominate. 3.1.4. Combustion efficiency The analysis of the incomplete combustion products (e.g. unburned hydrocarbons (HC) and carbon monoxide (CO) emissions) is represented by the combustion inefficiency. During the combustion process, not all the chemical energy is released. The fraction that is burned, compared to that which is supplied, is expressed by the combustion efficiency [46] calculated using Eq. (3) [47]:

Combustion Efficiency gc ¼ 1 

Rxi Q LHVi _ fuel =ðm _ air þ m _ fuel ÞQ LHVfuel ½m

ð3Þ

where xi and Q LHVi represent the mass fractions and lower heating values (LHV) of HC, CO, nitric oxide (NO) and hydrogen (H2),

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R. Daniel et al. / Applied Energy 105 (2013) 252–261

1500rpm, λ = 1

(a) 135

105

90 75 60 45

Dual-Injection IMEPMBT = 5.5bar

EDI G-E75DI G-E50DI G-E25DI PFI

120

dM/dLogDp (μg/m3)

105

dM/dLogDp (μg/m3)

Direct-Injection IMEPMBT = 5.5bar

EDI E75DI E50DI E25DI GDI

120

1500rpm, λ = 1

(b) 135

90 75 60 45

30

30

15

15

0

0 10

100

10

100

Particle Diameter (nm)

Particle Diameter (nm)

Fig. 10. PM mass distribution for various ExDI (a) and G-ExDI (b) lends at part load (5.5 bar IMEP) compared to GDI and PFI, respectively.

3.2. particulate matter emissions Currently, particulate matter (PM) number emissions do not form part of the emissions legislations for gasoline SI engines in Europe or the US. However, control of these emissions is expected to commence in European regulation in 2014 [48]. This will require not only the monitoring of PM mass emissions, but also the PM number for all light-duty vehicles. Therefore, an understanding of these emissions will become much more important, especially when using biofuels. In this section, the PM emissions results are discussed for ethanol blends only at a selected part-load condition of 5.5 bar IMEP under the dual-injection and DI strategies. This is because part-load conditions are frequently visited operating points in emissions drive cycle tests. The same definitions as used for DMF in the previous sections and shown in Table 3 are applied

1500rpm, λ = 1

140 120

Mean Diameter (nm)

respectively. For this work, Q LHVHC has been treated equal to Q LHVfuel . In reality, the HC emissions contain different components or species, which have different enthalpies of formation or LHV. In future work, the authors intend to quantify these species in order to better understand the combustion efficiency of each fuel. The combustion efficiency for D25 blends in dual-injection and DI are compared in Fig. 9 with reference to the individual constituents. The typical range of combustion efficiency for SI engines operating under lean conditions, is between 95% and 98% [46] and therefore, the results in this work which were collected under stoichiometric conditions are as expected. For most fuelling cases, the combustion efficiency increases with increasing load, due to decreased emissions of HCs and CO for higher combustion temperatures. The effect of the DMF fraction is to increase the high load combustion efficiency. For D25DI, the combustion efficiency reaches 95% at 8.5 bar IMEP. This represents a relative increase of 0.8% over GDI. However, for G-D25DI the increase over GDI is much higher. At 8.5 bar IMEP, the combustion efficiency with G-D25DI reaches 95.6%, which is almost the relative increase of D25DI over GDI (1.5%). This is due to the increased efficiency of the PFI component over GDI (0.5% higher at 8.5 bar IMEP). However, the synergistic behavior of PFI and DDI results in consistently higher combustion efficiencies for G-D25DI than the constituent parts as seen in previous work by the authors [26]. This is mainly due to the low isCO emissions when using G-D25DI, as shown in Fig. 7, which is caused by the improved mixture preparation of fuel and air prior to combustion with G-D25DI and the improved chemical reaction with ensues during the combustion process [43].

IMEPMBT = 5.5bar

Direct-Injection Dual-Injection

100 80 60 PFI

40 20 0 E100

E75

E50

E25

ULG

Fig. 11. PM mean diameter for various ExDI and G-ExDI blends at part load (5.5 bar IMEP) compared to GDI and PFI, respectively.

to the ethanol blends. Ethanol has been used to present the differences in PM emissions between the two combustion modes because it is the most widely used commercial biofuel. It is commonly used in blends with gasoline and has received the most attention as a dual-injection fuel [30–32]. 3.2.1. Mass distribution Fig. 10a shows the shift in PM mass distribution from a unimodal distribution with EDI, to a bimodal distribution with GDI and various distributions of the blends in between (E25DI, E50DI and E75DI). As the ethanol fraction decreases, the height of the first peak decreases (associated with volatile particles and known as the nucleation mode) and a second peak appears (associated with solid core (soot) particles and known as the accumulation mode). Therefore, it appears that the effect of the oxygen content in ethanol is to reduce the large soot particles but consequently increase the smaller particle count. Although the first peak with E25DI is positioned as expected (in blend order), the second peak exceeds that with GDI, which is unexpected. This indicates a non-linear relationship with blend ratio in the accumulation mode. In comparison to the DI results, the equivalent dual-injection PM data in Fig. 10b results in a consistent unimodal distribution, where the larger accumulation mode particles present in the DI

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1500rpm, λ = 1

(a) 107

1500rpm, λ = 1

(b) 107

Dual-Injection IMEPMBT = 5.5bar

Direct-Injection IMEPMBT = 5.5bar

10 6

dN/dLogDp (#/cm3)

dN/dLogDp (#/cm3)

10

6

10 5

10 4 EDI E75DI E50DI E25DI GDI

10 3

10 2

10 5

10 4 EDI G-E75DI G-E50DI G-E25DI PFI

10 3

10 2

10 1

10 1 10

100

10

100

Particle Diameter (nm)

Particle Diameter (nm)

Fig. 12. PM number distribution for various ExDI (a) and G-ExDI (b) blends at part load (5.5 bar IMEP) compared to GDI and PFI, respectively.

3.2.2. Number distribution The PM number distribution also reveals differences between the two injection modes and is shown in Fig. 12. The PM number distribution using blends in DI in Fig. 12a shows a high concentration of accumulation mode particles (>50 nm), which is much reduced with the dual-injection blends in Fig. 12b. This trend is clearly shown from the shift from GDI (Fig. 12a) to PFI (Fig. 12b) with gasoline. For instance, the vaporization of fuel droplets is

1500rpm, λ = 1

11

Total Concentration (#/cm3 x105)

blends in Fig. 10a are removed. Nevertheless, the consequence is a higher nucleation mode peak. At this load (5.5 bar IMEP), the PFI fraction encourages fuel vaporization and the increased combustion temperature reduces the accumulation of particles. This ensures that fuel droplets are burned more completely and soot formation is suppressed. The greater the PFI volume fraction, the lower the peak particle mass becomes. This is due to increased mixture preparation time and reduced wetting of the piston and cylinder walls with DI. Similarly, to DI blends, the decreasing ethanol fraction in dual-injection largely results in proportional decreases in the nucleation mode. The vaporization and breakup of large ethanol fuel droplets is evidently less proficient than with gasoline due to the low in-cylinder temperature, which has been shown in previous work [14]. However, the distribution for GE75DI and G-E50DI are inseparable and so the relationship of the PM mass emissions with blend ratio is not strictly linear. Further testing would have to be conducted to confirm this disparity. The injection timing using DI is early and represents a homogenous mode (see Table 1). If stratification was employed, in the absence of exhaust gas recirculation (EGR), the PM emissions would increase due to decreased fuel evaporation time and the potential for greater wall wetting with late injections. A statistical summary of the PM size distributions is shown in Fig. 11, which includes error bars to highlight the repeatability. The mean PM diameters for blends in dual-injection are consistently lower than in DI. Despite the negligible difference between E75DI and G-E75DI blends, decreases in ethanol content results in clear differences and a large increase in mean diameter with EDI towards GDI. Dual-injection consistently results in particles with lower mean diameter which is similar to the level for gasoline PFI. For instance, the mean PM diameter with E25DI (115.3 nm) is 75 nm larger than G-E25DI (40.3 nm) and even greater than GDI (100.3 nm). This quantifies the trends seen in Fig. 10, whereby the PFI component has contributed to a significant reduction in PM mass emissions.

10

IMEPMBT = 5.5bar

Direct-Injection Dual-Injection

9 PFI

8 7 6 5 E100

E75

E50

E25

ULG

Fig. 13. PM total number for various ExDI and G-ExDI blends at part load (5.5 bar IMEP) compared to GDI and PFI, respectively.

improved with PFI due to the increased mixing time and combustion temperatures. This results in more complete combustion and reduces the tendency of particles to accumulate. Although the effect of the ethanol content is to deteriorate the mixing ability, at the same time, it helps to complete combustion due to the oxygen molecule inherent in its chemical structure. The end result of the dual-injection mode is a more effective reduction of the large accumulation particles compared to equivalent gasoline–ethanol blends in DI. Fig. 13 presents the statistical summary of the PM number distributions with error bars included to highlight the repeatability. Here, it is the DI blends which produce the lowest total PM numbers compared to dual-injection. The decrease in number is linear from E100 through to gasoline in DI mode. However, with dualinjection, the concentration is always >230,000/cm3 compared to the equivalent blend in DI. This is clearly the consequence of dual-injection. Although dual-injection helps to lower the mean particle diameter and reduce the soot concentration, it conversely increases the total number concentration compared to gasoline– ethanol blends in DI, due to the increase in the volatile component (nucleation mode particles). The summary of the two PM number modes - nucleation and accumulation - is shown in Fig. 14a and b, respectively. The separation of the two modes is based on the point where they intersect

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11

(b) 1500rpm, λ = 1 Direct-Injection Dual-Injection

IMEPMBT = 5.5bar

10 9 PFI

8 7 6 5 E100

E75

E50

E25

ULG

Acc. Mode Concentration (#/cm3 x103)

Nucl. Mode Concentration (#/cm3 x105)

(a)

24

1500rpm, λ = 1 Direct-Injection Dual-Injection

IMEPMBT = 5.5bar

20 16 12 8 4 PFI

0 E100

E75

E50

E25

ULG

Fig. 14. PM number concentrations for nucleation (a) and accumulation modes (b), as determined by the points of inflection from the PM Number Distribution, using various ExDI and G-ExDI blends at part load (5.5 bar IMEP) compared to GDI and PFI, respectively.

(rather than a fixed particle diameter), as shown around 50 nm in Fig. 12. The data processing indicates that the nucleation mode counts for >98% of the total PM number concentration with GDI and thus dominates the overall PM concentration as shown by the similarity between Figs. 13 and 14a. Clear differences can be observed between the two injection modes by analyzing the accumulation mode PM concentrations. Fig. 12 has shown that the accumulation mode particles with PFI are significantly lower than with GDI. Fig. 14b reveals that despite an increase in the ethanol fraction in DI using dual-injection, the accumulation mode particles are always lower than 350/cm3. This is due to the effectiveness of PFI to reduce accumulation particles. The G-E25DI strategy results in a total accumulation mode particle concentration of 244/ cm3, which is 1/80 of the level of accumulation mode particles in E25DI (19,550/cm3).

4. Conclusions This study furthers the understanding of dual-injection: the effect of amalgamating PFI and DI technology involving the use of DMF (a new biofuel candidate) and ethanol compared with gasoline (the PM emissions are analyzed using ethanol). Based on these experimental investigations, several conclusions can be drawn. Using dual-injection, the isHC emissions of D25 (G-D50DI) are comparable to DI for the load conditions between 3.5 bar and 8.5 bar IMEP (±4%). The oxygen content of the DMF fraction helps to lower the isHC emissions compared to GDI and PFI. More often than not, dual-injection offers marginally lower isHC emissions than DI. For instance, D50 blends in G-D50DI result in up to 4% lower isHC emission than in DI (D50DI), regardless of the engine load. The lower emissions of isHCs at medium blend ratios has also been reported with gasoline in dual-injection, especially at higher loads [19]. The isNOx emissions of G-D25DI are between 13% and 16% higher than D25DI as the load is varied from 3.5 bar to 8.5 bar IMEP. This is due to higher charge temperatures in dual-injection which results in higher combustion temperatures. However, NOx emissions are higher due to higher combustion temperatures. At 8.5 bar IMEP, the isNOx emissions of D25DI are much lower than G-D25DI and are comparable to PFI and GDI. Nevertheless, the isNOx emissions of gasoline in dual-injection have been shown to consistently lower than with DMF in dual-injection [19], largely due to the greater effect of charge-cooling with gasoline in DI. As the DMF fraction increases, the difference between the two injection

modes decreases. For instance, with D75 blends, the increase of isNOx with G-D75DI is less than 10% for all loads. The isCO emissions are consistently lower when using dualinjection compared to DI. The greatest decrease of 9.4 g/kW h (31%) is seen at 5.5 bar IMEP with G-D25DI. The gasoline component in PFI aids the vaporization and increases the reaction rate helping to consume the DMF fuel droplets. Using ethanol fractions in dual-injection results in a unimodal mass distribution where the larger, accumulation mode particles are removed. Compared to blends in DI, dual-injection therefore helps to lower the mean PM diameter. The mean PM diameter with E25DI at 5.5 bar IMEP is 115.3 nm compared to 40.3 nm with GE25DI. This is due to higher combustion temperatures with dualinjection, which ensures the fuel droplets are more completely burned which helps to suppress the formation of soot. The increased mixture preparation time with dual-injection also helps to reduce wall wetting as associated with DI. Although the PM number concentration in the nucleation mode increases with dual-injection, due to increased PFI component, the accumulation mode particles are almost eliminated. At 5.5 bar IMEP, the accumulation mode particles with G-E25DI are 1/80 of that with E25DI. This is due to the increased mixing time and combustion temperatures. Ultimately, there are clear differences in emissions between gasoline–biofuel blends in dual-injection and DI. The flexibility of dual-injection allows the reduction in emissions to be optimized with changes in engine load by instantaneously varying the blend ratio, which is not available with fixed DI blends. The emissions of CO and PM (accumulation mode) are effectively reduced with dualinjection and the HC emissions are comparable to DI blends. However, the consequence of high combustion temperatures with dualinjection is an increase in NOx emissions. Acknowledgments The present work is part of a 3-year research project sponsored by the Engineering and Physical Sciences Research Council (EPSRC) under the grant EP/F061692/1. The authors would like to acknowledge the support from Jaguar Cars Ltd., Shell Global Solutions and various research assistants and technicians especially Xuesong Wu, Jose Martin Herreros and Shahrouz Norouzi. The authors are also grateful for the financial support from the European Regional Development Fund (EUDF) and Advantage West Midlands (AWM). Finally, the authors would like to acknowledge the support from their international collaborators at Tsinghua University, China.

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