on the combustion and performance of a natural gas direct ... · plug fuel injector which combines...

European Journal of Scientific Research ISSN 1450-216X / 1450-202X Vol. 147 No 4 November, 2017, pp. 446-459 http://www. europeanjournalofscientificresearch.com

On the Combustion and Performance of a Natural Gas Direct

Injection Spark Ignition Engine with Spark Plug Fuel

Injector at Near Idling Condition

Taib Iskandar Mohamad

Department of Mechanical Engineering Technology

Yanbu Industrial College, Yanbu Alsinaiyah, Saudi Arabia

Centre for Automotive Research, Universiti Kebangsaan Malaysia

43600 Bangi, Selangor, Malaysia

E-mail: [email protected] Tel: +966-56-7123774; Fax +966-14-3946144

Abstract

Direct injection in natural gas engine offers a potential of improved performance

and thermal efficiency due to increased cylinder charge heating value and possibility of fuel stratification. An engine was converted to direct injection of natural gas using spark plug fuel injector which combines fuel injector and spark plug for easy conversion. The effect of injection timing on the performance and combustion in a 0.507L single-cylinder engine fuelled with natural gas substitute, methane was investigated. Engine was run at of 1100 rpm, 25oCA BTDC ignition and stoichiometric mixture. Start of injection were varied between 160o and 200o BTDC. The optimal start of injection, SOI timing is 170oCA BTDC at the test conditions. Higher burning rate at rapid combustion phase and longer ignition delay are evidenced over port injection. Total combustion and rapid combustion duration are longer with early injection. At the optimal SOI, combustion analysis indicated that long ignition delay, short rapid combustion duration and phasing angle of 1.5oCA ATDC produced highest heat release rate and engine performance. DI operation improves volumetric efficiency by up to 13% especially at low engine speeds. At the optimal SOI, volumetric efficiency, MEP and fuel conversion efficiency are 83.4%, 625 kPa and 21.8% respectively.

Keywords: Natural Gas Engine, Engine Fuel Conversion, Direct Fuel Injection, Burning

Rate, Combustion Characteristics, Indicated Performance.

Introduction Increasing environmental pollution and uncertainty in fossil fuel reserves are two major concerns related to energy generation and utilization. Transportation is mainly powered by internal combustion engines (ICE). Energy use by transportation accounts for nearly 20% of world consumption while contributes to 18% of greenhouse gas emission. Land transportation consumed 72% of the total energy use within this sector and released 80% of the total CO2 emissions. During the 1990–2005 period alone, the energy use for transportation increased by 37% thus making it the fastest growing energy sector in the world [1, 2]. The number is increasing with the continuous growth of vehicle numbers.

Natural gas (NG) is the most widely used alternative fuel in ICE mainly due to resource availability [3], easy adaptability to current ICE and attractive price over conventional fuels. In the recent years, the number of natural gas fuelled vehicles has increased significantly and the share in

On the Combustion and Performance of A Natural Gas Direct Injection Spark Ignition Engine with Spark Plug Fuel Injector At Near Idling Condition 447 global scale is expanding [4]. NG is used either as supplement or sole fuel in these engines, and is usually supplied using conventional fuel systems (i.e. carburetor, port or manifold injection) [5-14]. Natural gas is considered the cleanest, most promising and very practical alternative fuel. Engine CO2 emission with NG is significantly lower than conventional fuels for the same amount of energy supplied due to high hydrogen-to-carbon ratio [2, 15, 16]. The composition of NG, which is mainly methane, has significant effects on engine performance and emissions [8]. High Wobbe number which reflects high hydrocarbon content results in higher power density, better fuel economy, lower CO2, CO and HC but higher NOX emissions [6]. The calorific value of NG is significantly superior to gasoline and diesel due to high hydrogen fraction in the hydrocarbon. High auto ignition temperature and high octane number make it more suitable for high compression ratio operation, leading to higher thermal efficiency potential. However due to its gaseous nature with low density, volumetric energy content is less than liquid fuels but this can partially overcame with pressurized (CNG) or liquefied (LNG) storage. Despite various advantages, when an engine is fuelled with NG with externally mixing fuel systems, engine power is reduced and maximum engine speed is limited. These are due to reduced volumetric efficiency and lower flame speed of NG-air combustion [17-19]. Direct injection (DI) can mitigate these problems because it reduces volumetric efficiency loss, and the high velocity gas jet intensifies in-cylinder turbulence.

Direct injection of natural gas has been studied in recent years where higher power can be realized compared to port injection gasoline at part loads [16, 20-27]. Shiga et al. investigated the stratified combustion and resulting emissions of CNG DI engine using a rapid-compression machine (RCM) simulating a disk-shaped combustion chamber with compression ratio of 10. They compared the effects of three injector locations and varied mixture equivalence ratio. High combustion efficiency up to 95% can be achieved by carefully positioning the fuel injector and flammability limits were extended to equivalent ratio of 0.02 [15]. Huang et al. studied the effects of ignition position with respect to injector nozzle tip in a CNG DI engine. The study which was carried out using a RCM revealed that a very lean equivalence ratio of 0.1 was achieved, but the flammability range narrowed down as ignition electrode gap move nearer to injector tip. Also there were direct relation between early injection to longer early combustion and late injection to longer late combustion [28, 29]. Two studies by Bo Yang et. al and Zeng et al. investigated the effect of injection timings on the combustion characteristics of a CNG DI engine. They found that late injection has significant impact on combustion quality and optimal injection timing correspond to maximum cylinder pressure, maximum rate of pressure rise, maximum heat release rate (HRR), and shortest combustion duration [30, 31]. In a full load operation with natural gas direct injection at optimal after intake-valve-close fuel injection timing can speed up the early flame development process by nearly 18°CA, allows for up to a 10% improvement in full-load power density and up to a 30% reduction in carbon dioxide emissions over liquid and gaseous port-fuel injection for a naturally aspirated engine [32]. The effects of compression ratio (CR) on the performance and combustion of a CNG DI engine was investigated by Zheng et al. [26]. Increasing CR results in stronger mixture stratification, shorter injected gas jet penetration, faster burning rate and higher thermal efficiency up to CR of 12 especially with low load and lean mixture. Kalam et al. converted a four-cylinder gasoline engine to CNG DI operation with the same engine geometry. The brake power, brake thermal efficiency and brake specific fuel consumption was improved with CNG Di especially at part load conditions but with penalty of increased hydrocarbons and CO emissions [16]. The dynamics of cycle-to-cycle variation (CCV) of indicated mean effective pressure (IMEP) signals of a CNG DI engine were investigated using continuous wavelet transform technique by Sen et al. They found that varying CR and engine speed significantly influence the CCV spectral power. In addition, there is a strong interdependency between IMEP and combustion duration [24]. CNG injection spray characteristics and jet development in DI environment was studied using a fixed-volume combustion chamber showing high sensitivity to injection pressure, air velocity and ambient temperature [33, 34]. Yiu et al. reported that decreasing the distance between spark plug and fuel injector is favorable in achieving effective fuel stratification [22]. Yadollahi et al. numerically

448

investigated the effect of combustion chamber geometry on injection and mixture preCNG DI engine using AVL Fire software. They discovered that geometrical arrangement, injection parameters, injector type, cylinder head and piston crown shapes have significant effects on engine performance of natural gas engine even with the presence of diesel pilot injection of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing al. supported this fineffects towards mixture formation in a CNG DI engine. They discovered two stages of mixture evolution during the injection event and compression stroke Table

This study

Aljamali [48

Zeng

Mohammed [49

Kalam

engines meant to explain the novel aspect of this study. In all listed previous studies dedicated direct fuel injectors were used where natural gas was injected towards igInjection pressure were tested at lower range between 1.4 to 2 MPa except in the work by Zeng where 8 MPa injection was used with low compression ratio of 8 [30]. All other works took advantage of high octane number by operCR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to achieve direct fuel injection is by combining fuel injector with spark plug but wiinjecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder internal flow and wall impingement.

methane thermodynamic properties due to varying composition based on production area properties that affect emissions of NG engines are the methane number (MN) and the Wobbe number (WN) be avoided by using methane NG DI with technical simplicity and minimal cost by avoiding device called spark plug fuel injector (SPFI) was developed and described previously by combining a fuel injector and a spark plug injected from a single 0.7mmpartially expanded before entering the combustion chamber. The SPFI injection nozzle equivalent diameter is

448

investigated the effect of combustion chamber geometry on injection and mixture preCNG DI engine using AVL Fire software. They discovered that geometrical arrangement, injection parameters, injector type, cylinder head and piston crown shapes have significant effects on engine performance [of natural gas engine even with the presence of diesel pilot injection of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing al. supported this fineffects towards mixture formation in a CNG DI engine. They discovered two stages of mixture evolution during the injection event and compression stroke

Table 1 Test conditions of some experimental studies of direct injection NG/CH

This study

Aljamali 48]

Zeng [30]

Mohammed 49]

Kalam [16]

Table 1 shows a summary of some experimental studies of CNG/methane direct injection engines meant to explain the novel aspect of this study. In all listed previous studies dedicated direct fuel injectors were used where natural gas was injected towards igInjection pressure were tested at lower range between 1.4 to 2 MPa except in the work by Zeng where 8 MPa injection was used with low compression ratio of 8 [30]. All other works took advantage of high octane number by operCR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to achieve direct fuel injection is by combining fuel injector with spark plug but wiinjecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder internal flow and wall impingement.

In this work, a unique approach to achieve direct injection of NG is further elaborated. Pure methane was used as it is the main constituent of NG to avoid uncertainty of natural gas’s thermodynamic properties due to varying composition based on production area properties that affect emissions of NG engines are the methane number (MN) and the Wobbe number (WN) [39]. Therefore, significant effects of NG composition on engine performance and emissions can be avoided by using methane NG DI with technical simplicity and minimal cost by avoiding device called spark plug fuel injector (SPFI) was developed and described previously by combining a fuel injector and a spark plug injected from a single 0.7mmpartially expanded before entering the combustion chamber. The SPFI injection nozzle equivalent diameter is significantly larger than typical fuel injector nozzle but smaller than the quenching distance

investigated the effect of combustion chamber geometry on injection and mixture preCNG DI engine using AVL Fire software. They discovered that geometrical arrangement, injection parameters, injector type, cylinder head and piston crown shapes have significant effects on engine

[35]. Bin Wang et.al cof natural gas engine even with the presence of diesel pilot injection of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing al. supported this finding in his numerical study on nozzle type, injection pressure and injection timing effects towards mixture formation in a CNG DI engine. They discovered two stages of mixture evolution during the injection event and compression stroke

Test conditions of some experimental studies of direct injection NG/CH

Engine

parameters

1-cylinder, 0.5L, CR 10.5

4-cylinder, 1.6L, CR 14




Table 1 shows a summary of some experimental studies of CNG/methane direct injection engines meant to explain the novel aspect of this study. In all listed previous studies dedicated direct fuel injectors were used where natural gas was injected towards igInjection pressure were tested at lower range between 1.4 to 2 MPa except in the work by Zeng where 8 MPa injection was used with low compression ratio of 8 [30]. All other works took advantage of high octane number by operating at CR 14. In this work, 6 MPa injection pressure was used and a typical CR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to achieve direct fuel injection is by combining fuel injector with spark plug but wiinjecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder internal flow and wall impingement.

In this work, a unique approach to achieve direct injection of NG is further elaborated. Pure was used as it is the main constituent of NG to avoid uncertainty of natural gas’s

thermodynamic properties due to varying composition based on production area properties that affect emissions of NG engines are the methane number (MN) and the Wobbe number

. Therefore, significant effects of NG composition on engine performance and emissions can be avoided by using methane NG DI with technical simplicity and minimal cost by avoiding device called spark plug fuel injector (SPFI) was developed and described previously by combining a fuel injector and a spark plug injected from a single 0.7mmpartially expanded before entering the combustion chamber. The SPFI injection nozzle equivalent

significantly larger than typical fuel injector nozzle but smaller than the quenching distance


. Bin Wang et.al confirmed the importance of cylinder geometry on the performance of natural gas engine even with the presence of diesel pilot injection of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing

ding in his numerical study on nozzle type, injection pressure and injection timing effects towards mixture formation in a CNG DI engine. They discovered two stages of mixture evolution during the injection event and compression stroke


parameters

Injection

pressure

cylinder, 0.5L, 6 MPa



cylinder, 0.4L, 1.4 MPa


Table 1 shows a summary of some experimental studies of CNG/methane direct injection engines meant to explain the novel aspect of this study. In all listed previous studies dedicated direct fuel injectors were used where natural gas was injected towards igInjection pressure were tested at lower range between 1.4 to 2 MPa except in the work by Zeng where 8 MPa injection was used with low compression ratio of 8 [30]. All other works took advantage of high

ating at CR 14. In this work, 6 MPa injection pressure was used and a typical CR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to achieve direct fuel injection is by combining fuel injector with spark plug but wiinjecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder internal flow and wall impingement.



. Therefore, significant effects of NG composition on engine performance and emissions can be avoided by using methane [6, 40]. This approach was aimed to convert any spark ignition engine to NG DI with technical simplicity and minimal cost by avoiding device called spark plug fuel injector (SPFI) was developed and described previously by combining a fuel injector and a spark plug [17, 21injected from a single 0.7mm-diameter nozzle GDI injector, passes through a minipartially expanded before entering the combustion chamber. The SPFI injection nozzle equivalent

significantly larger than typical fuel injector nozzle but smaller than the quenching distance


onfirmed the importance of cylinder geometry on the performance of natural gas engine even with the presence of diesel pilot injection of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing

ding in his numerical study on nozzle type, injection pressure and injection timing effects towards mixture formation in a CNG DI engine. They discovered two stages of mixture evolution during the injection event and compression stroke


Injection

pressure Injection timing

SOI 160BTDC

EOI 120BTDC

SOI 150BTDC

1.4 MPa SOI 120300o

Varies by ECU for MBT


ating at CR 14. In this work, 6 MPa injection pressure was used and a typical CR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to achieve direct fuel injection is by combining fuel injector with spark plug but wiinjecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder



. Therefore, significant effects of NG composition on engine performance and emissions can . This approach was aimed to convert any spark ignition engine to

NG DI with technical simplicity and minimal cost by avoiding device called spark plug fuel injector (SPFI) was developed and described previously by combining a

21, 33, 41-43diameter nozzle GDI injector, passes through a mini

partially expanded before entering the combustion chamber. The SPFI injection nozzle equivalent significantly larger than typical fuel injector nozzle but smaller than the quenching distance


onfirmed the importance of cylinder geometry on the performance of natural gas engine even with the presence of diesel pilot injection of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing

ding in his numerical study on nozzle type, injection pressure and injection timing effects towards mixture formation in a CNG DI engine. They discovered two stages of mixture evolution during the injection event and compression stroke [


Injection timing

SOI 160o - 200o BTDC

EOI 120o - 360o BTDC

SOI 150o - 210o BTDC

SOI 120o, 180o, o BTDC

Varies by ECU for MBT






NG DI with technical simplicity and minimal cost by avoiding device called spark plug fuel injector (SPFI) was developed and described previously by combining a

43]. Unique in its nature of direct injection, methane is diameter nozzle GDI injector, passes through a mini



onfirmed the importance of cylinder geometry on the performance of natural gas engine even with the presence of diesel pilot injection [36of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing

ding in his numerical study on nozzle type, injection pressure and injection timing effects towards mixture formation in a CNG DI engine. They discovered two stages of mixture

[4].


Fuel

CH4

CNG (94% CH4)

CNG

CNG + 0-8% H2

CNG (94% CH4)

Table 1 shows a summary of some experimental studies of CNG/methane direct injection engines meant to explain the novel aspect of this study. In all listed previous studies dedicated direct fuel injectors were used where natural gas was injected towards ignition region of the spark plug. Injection pressure were tested at lower range between 1.4 to 2 MPa except in the work by Zeng where 8 MPa injection was used with low compression ratio of 8 [30]. All other works took advantage of high





NG DI with technical simplicity and minimal cost by avoiding any modification to engine structure. A device called spark plug fuel injector (SPFI) was developed and described previously by combining a

. Unique in its nature of direct injection, methane is diameter nozzle GDI injector, passes through a mini




onfirmed the importance of cylinder geometry on the performance 36]. Yu et al. studied the impact

of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing


Test conditions of some experimental studies of direct injection NG/CH4 engine

Injector-spark plug

orientation

Combined – axis

Table 1 shows a summary of some experimental studies of CNG/methane direct injection engines meant to explain the novel aspect of this study. In all listed previous studies dedicated direct

nition region of the spark plug. Injection pressure were tested at lower range between 1.4 to 2 MPa except in the work by Zeng where 8 MPa injection was used with low compression ratio of 8 [30]. All other works took advantage of high



thermodynamic properties due to varying composition based on production area [38]properties that affect emissions of NG engines are the methane number (MN) and the Wobbe number


any modification to engine structure. A device called spark plug fuel injector (SPFI) was developed and described previously by combining a

. Unique in its nature of direct injection, methane is diameter nozzle GDI injector, passes through a mini



investigated the effect of combustion chamber geometry on injection and mixture preparation in a CNG DI engine using AVL Fire software. They discovered that geometrical arrangement, injection parameters, injector type, cylinder head and piston crown shapes have significant effects on engine

onfirmed the importance of cylinder geometry on the performance . Yu et al. studied the impact

of wall impingement of mixture formation in a low pressure CNG injection to find that position and direction of fuel injection need to be carefully calibrated to achieved optimal mixing [37]. Keskinen et


spark plug

orientation

parallel

0.99

1.8

Varies MBT



ating at CR 14. In this work, 6 MPa injection pressure was used and a typical CR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to achieve direct fuel injection is by combining fuel injector with spark plug but with the penalty of injecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder


]. The two primary properties that affect emissions of NG engines are the methane number (MN) and the Wobbe number



. Unique in its nature of direct injection, methane is diameter nozzle GDI injector, passes through a mini-channel and



paration in a CNG DI engine using AVL Fire software. They discovered that geometrical arrangement, injection parameters, injector type, cylinder head and piston crown shapes have significant effects on engine

onfirmed the importance of cylinder geometry on the performance . Yu et al. studied the impact

of wall impingement of mixture formation in a low pressure CNG injection to find that position and . Keskinen et


λλλλ

1.0

0.99-1.8

1.8-1.9

1.0

Varies for MBT



ating at CR 14. In this work, 6 MPa injection pressure was used and a typical CR of 10.5 found in many modern gasoline engine was used. But more importantly, a unique way to

th the penalty of injecting fuel away from point of ignition. Therefore, mixture ignition is governed by the cylinder


. The two primary properties that affect emissions of NG engines are the methane number (MN) and the Wobbe number



. Unique in its nature of direct injection, methane is channel and


On the Combustion and Performance of A Natural Gas Direct Injection Spark Ignition Engine with Spark Plug Fuel Injector At Near Idling Condition 449 of NG-air combustion. This is to prevent flame penetration into the mini channel. The direction of the injection is parallel to the spark plug which is theoretically more challenging in controlling the ignitability of the mixture because bringing fuel to the vicinity of spark plug heavily depends on flow characteristics after piston crown impingement. In this work, these challenges were overcame by carefully timing the ignition and injection events until steady operation is achieved. Visualization of gas jet from SPFI was studied in both gas and liquid environments [34]. Experimental work with the SPFI system was done on a single cylinder research engine. The effects of fuel injection and ignition timings on the combustion and performance were studied.

2. Experimental Set Up and Procedures The SPFI was mounted on a single cylinder Ricardo E6 engine with spark ignition head [44]. Table 1 lists the engine specification. A driver circuit was developed to control injection timing and duration using a field effect transistor, a power supply unit and a signal generator. The engine was connected to a DC current dynamometer which can motor and absorb power from the engine. Lubrication and cooling liquids were driven by separate motor and centrifugal pumps. In Figure 1 the plan view of cylinder shows the arrangement of SPFI and pressure transducer with respect to primary axis, intake and exhaust valves.

Figure 2 shows the cross-section of the disk-shaped engine combustion chamber with flat cylinder head and piston crown. The intake and exhaust valves are poppet-type. The intake and exhaust valves arrangement resulted in both tumble and swirl charge motions. SPFI takes the place of spark plug protruding 60o towards the centerline of the cylinder. For comparison, the same engine was run with manifold fuel injection system in which fuel injection takes place 15 cm upstream from the intake valve in the intake manifold using the same fuel injector. Table 2 Engine specifications

Bore (mm) 76 Stroke (mm) 111 Displacement volume (cm3) 507 Compression ratio 10.5: 1 Air intake Naturally aspirated Intake valve open 8o BTDC Intake valve close 33o ABDC Exhaust valve open 42o BBDC Exhaust valve close 8o ATDC Thermal management Water cooled Valve clearance (intake/exhaust) 0.15 mm / 0.20 mm

Figure 1: Plan view of combustion chamber

450 Taib Iskandar Mohamad

Figure 2: Cross-sectional view of combustion chamber

Figure 3: Experimental schematic

The schematic of experimental set up is shown in Figure 3. Ignition timing was varied between 15 and 30o BTDC. Engine speed was set at a typical high idling speed of 1100 rpm. A pressure sensor of Kistler 6121 A1 model was installed through the secondary spark plug hole on the cylinder head to measure in-cylinder pressure. The signal is amplified and synchronized with the crank angle signal at 12 kHz sampling rate. Engine speed and crank angle were determined by a shaft encoder on the crank shaft. Another shaft encoder was mounted to the camshaft. This encoder sent one TTL signal in every two crankshaft rotations, which was set as an input signal to a pulse generator. The output from the pulse generator is set at the desired pulse length and delay, is then sent to a MOSFET which acts as a gate between a power supply unit of 12V and 5A that control the GDI injector.

Methane gas was supplied at 23 MPa from a tank. A pressure regulator reduced the pressure to

6 MPa, which was the set injection pressure. Air-fuel mixture was set at stoichiometric (λ=1.0). The injection and ignition timings were referenced to compression stroke top dead center, BTDC. Error!

eference source not found. shows the injection timings defined by the start of injections (SOI) with reference to cylinder pressure of non-combusting operation and intake valve closing. The injection durations were constant at 12 ms for all injection timings regardless of injection timing, i.e. at the given range of operation; cylinder pressure does not affect fuel delivery timing since gas jet at SPFI nozzle

On the Combustion and Performance of A Natural Gas Direct Injection Spark Ignition Engine with Spark Plug Fuel Injector At Near Idling Condition 451 are at choke flow. Even though SOI occurs at different cylinder pressures, the amount of fuel delivered remains unchanged due to a high injection-to-cylinder pressure ratio. The fuel stoichiometry was found constant over all injection timings which were confirmed by an exhaust lambda sensor during engine running.

Figure 4: Injection parameters

With port/manifold injection, all the operational setting was similar to the ones of direct injection except that the injection pressure was set at a lower injection pressure of 3 MPa and the injection duration was increased accordingly. All data was taken at minimum spark ignition advance for best torque (MBT).

3. Results and Discussion The fundamental advantage of DI over PI operation is increased volumetric efficiency which provides the potential for increased power for the same mixture stoichiometry. Volumetric gain of DI operation compared to PI can be seen in Figure 4. Over the speed range, significant increase in charge efficiency between 8-13% were realized. As a result, similar increased in inhaled energy per cycle was achieved. However, while volumetric efficiency with PI increased with increasing speed, the DI operation showed almost unchanged behavior across the speed range with peak value was achieved at 1550 rpm but deteriorate as speed progressed.

Figure 4: Cylinder volumetric efficiency


In Figure 5 the cylinder pressure of 5 consecutive cycles are plotted for both PI and DI operation at 1100 rpm 25oCA BTDC ignition time and stoichiometric mixture. The peak cylinder pressure with DI reached peak value earlier than those of PI and their value were between 28% and 33% higher. This indicated that higher heat release rate and faster overall burning speed than PI operation. In term of cycle-to-cycle variation, DI operation showed a more uniform behavior over rapid burning period even though minimum-to-maximum peak pressure variation were almost identical. Figure 6 shows the normalized mass burnt fraction. It should be noted that both mixtures were ignited at the same spark time. PI injection showed shorter ignition delay (10% of mass burn fraction duration) but slower rapid burning (between 10% and 90% of mass burn fraction duration) and longer total combustion period. Total combustion duration with DI was reduced by 21% compared to PI operation at the operational conditions.

Figure 5: Cylinder pressure comparison over 5 consecutive cycles

Figure 6: Burn rate difference between DI and PI operations

On the Combustion and Performance of A Natural Gas Direct Injection Spark Ignition Engine with Spark Plug Fuel Injector At Near Idling Condition 453

It should be noted that the ignition delay with SPFI operation is significantly longer due to the absence of assistive features to mixture formation and bringing the right stoichiometry to ignition terminals. With SPFI, fuel is injected outwardly from spark plug parallel to its axis. The gas jet develops away from ignition terminals. Bringing methane back to ignition point depends on in-cylinder flow during compression stroke. In comparison to diesel DI engine, ignition delay is mainly influenced by fuel injection and atomization which is assisted by fast penetration of diesel into air and increasing mixture temperature during compression. In gasoline DI engine, ignition delay is influenced by the arrival of gasoline droplet within the vicinity of ignition terminals which are controlled either by wall-guided, spray-guided or air-guided injection strategies.

The effects of ignition timing were investigated with engine running at 1100 rpm and compared to the reference case of the same engine equipped with port/manifold fuel injection [45]. Indicated power and MEP were calculated from the pressure data. Peak cylinder pressure of 46, 44 and 46 bar was achieved with 30, 35 and 40o BTDC but the highest power and IMEP was realized at 25o BTDC ignition. Earlier ignitions lower output due to higher fraction of cylinder negative work before TDC. From this the ignition timing of 25o BTDC was selected for the next steps in this investigation.

The effects of varying injection timing on the performance and combustion of the test engine were investigated. The fundamental benefit of direct injection is the increased of volumetric efficiency as shown in Figure 4, hence for the same mixture stoichiometry cylinder charge heating value is higher compared to PI operations. This provide the potential for higher power with the right mixing and combustion behavior. However, since fuel injection happens after intake valve close, there are challenges from reduced spatial and temporal window for air-fuel mixing. As previously described by Keskinen et al., injection timing is the key parameters affecting charge mixing and the consequent combustion in direct injection gas engine [46]. The effects of varying injection timings by starting the injection between 160 to 220o CA BTDC were investigated with the aim to determine optimal start of injection (SOI) for best performance.

Figure 7: SOI influence on volumetric efficiency

Figure 8 shows the volumetric efficiencies with respect to injection timings. It should be noted that intake valve close (IVC) happens at 147o BTDC, which means all injection starts before IVC event and injection process continued after IVC at different fractions as shown in Error! Reference source

not found.. The combination of fuel injection and in-cylinder flow during IVO and IVC have strong influence on the amount of air inhaled. Advanced SOI, between 200-180 oCA BTDC, resulted in lower volumetric efficiencies due to air displacement by larger amount of methane injected during IVO. The highest efficiency achieved was 83.4% at 170o CA BTDC. After that volumetric efficiency dropped which can be attributed to the fact that as IVC is approaching, momentum of incoming air weakens and displacement by high cylinder pressure dominates the event

Figure 9: SOI influence on mean effective pressure


Figure 9: SOI influence on mean effective pressure

Figure 8: SOI influence on fuel conversion efficiency

Figure 9: SOI influence on maximum heat release rate

Shows the indicated mean effective pressure (IMEP) with respect to injection timings. The highest IMEP of 625 kPa was realized at the highest volumetric efficiency. At earlier injection timings, IMEP are lower but steadily moved towards peak value, while at later timings (after 170o CA), they reduced to constant values. With advanced injection timing, where cylinder pressure is relatively low, the gas jet travels further away from the spark plug position causing less effective initial burning. At

On the Combustion and Performance of A Natural Gas Direct Injection Spark Ignition Engine with Spark Plug Fuel Injector At Near Idling Condition 455 delayed injection, however, cylinder pressure is higher and the gas plume grows nearer to the spark plug point, resulting in more reliable ignition and more susceptible to complete combustion [33, 47]. The optimal injection timing was achieved by balancing these two factors. Figure 8 shows the fuel conversion efficiencies at various injection timings. As a result of IMEP behavior, conversion efficiency showed similar trend because injected fuel mass are uniform for all conditions. The peak conversion efficiency of 21.8% was realized at 170o CA BTDC start of injection.

Combustion characteristics were analyzed to further understand the performance behavior. Heat release rate (HRR) were derived from in-cylinder pressure data by the following equation,

��

��=

�

� − 1��

��+

1

� − 1��

��

where P is the cylinder pressure in kPa, V is the cylinder volume (m3) and θ is the crank angle. Peak heat release rate as shown in Figure 9 influenced engine output directly. Maximum value of 72 kJ/m3.oCA was achieved with 170oCA BTDC SOI. Highest HRR is a combined byproduct of highest charge efficiency, optimal mixing and combustion.

Figure 10: SOI influence on combustion parameters

Normalized mass burn fraction approach was used by differentiating pressure curve of reacting and non-reacting mixtures. The combustion is described by the total duration, ignition delay (10% of mass burnt), rapid burning (10% - 90% of mass burnt) and phasing angle (the crank angle where 50% of mass is burnt – CA50). In Figure 10, these parameters are presented with respect to SOI. Ignition delays display small differences across injection timings ranging between 15 and 18 oCA with the highest value found with SOI 170oCA BTDC. Rapid combustion show reverse trend compared to ignition delay. As SOI delayed from 150 to 170 oCA BTDC, rapid burning duration decreases to minimum duration but increases afterwards at much higher values. Total combustion duration shows a decreasing trend generally with delayed injection except with SOI 170oCA BTDC. The phasing angle (the location where 50% mass are burnt) fluctuated in a less predictive way. It goes up and down with the variation of 3oCA, the optimal point was found at SOI 170oCA BTDC, corresponding to 1.5oCA ATDC, where highest power was achieved. This optimal injection timing differs only slightly with the result from previous study where the optimal injection CA was 180 oCA BTDC [30]. This shows that the optimal balance of the four combustion parameters discussed above will give the best performance; long ignition delay, fast rapid burning, decreased total combustion duration and the optimal phasing angle, a few oCA after TDC. Thus it also show that injection timing has bigger influence on combustion and performance than the location of fuel injector with respect to spark plug.


4. Conclusions A spark ignition engine conversion from port injection to direct injection of methane, a natural gas substitute, was achieved with direct injection through spark plug. The effect of injection timing on the combustion and performance parameters was investigated. The engine was run at 1100 rpm, 25oCA BTDC start of ignition and stoichiometric mixture. At these test conditions the results can be summarized as the followings.

1. The optimal SOI was 170o CA BTDC at the test conditions. 2. At the optimal SOI, long ignition delay, short rapid combustion duration and phasing angle of

1.5oCA ATDC produced best performance. 3. Total combustion and rapid combustion durations were longer with early injection but

ignition delay affects the resulting phasing angle, which is a more dominant factor influencing engine output.

4. DI combustion yields faster burning rate but longer ignition delay compared to that of PI. This is due to richer mixture at spark plug vicinity during ignition phase.

5. DI operation improves the volumetric efficiency by 8-13% compared to PI operation. The increase is more dominant at low engine speeds.

6. The influence of injection timing are significant towards volumetric efficiency, MEP and fuel conversion efficiency. The maximum values are 84.3%, 625 kPa and 21.8% respectively.

7. Compared to some experimental studies on DI natural gas/methane engine, Natural gas engine; engine fuel conversion; direct fuel injection; burning rate; combustion characteristics; indicated performance this study indicates that injection timing has more influence on the combustion and performance than the location of fuel injector with respect to spark plug.

Acknowledgement The author thank Yanbu Industrial College for providing technical support for this work. Also to the Ministry of Science, Technology and Innovation, Malaysia under Science-Fund Project 03-01-02-SF0558.

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