; COMPARATIVE STUDY OF PERFORMANCE OF DUAL PLUG AND SINGLE PLUG SI ENGINE AT DIFF
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COMPARATIVE STUDY OF PERFORMANCE OF DUAL PLUG AND SINGLE PLUG SI ENGINE AT DIFF

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									    INTERNATIONAL JOURNAL OF ADVANCED RESEARCH ISSN
International Journal of Advanced Research in Engineering and Technology (IJARET),IN
0976 – 6480(Print), ISSN 0976 – 6499(Online) Volume 4, Issue 5, July – August (2013), © IAEME
               ENGINEERING AND TECHNOLOGY (IJARET)
ISSN 0976 - 6480 (Print)
ISSN 0976 - 6499 (Online)                                                    IJARET
Volume 4, Issue 5, July – August 2013, pp. 188-197
© IAEME: www.iaeme.com/ijaret.asp
Journal Impact Factor (2013): 5.8376 (Calculated by GISI)                    ©IAEME
www.jifactor.com




     COMPARATIVE STUDY OF PERFORMANCE OF DUAL PLUG AND
    SINGLE PLUG S.I ENGINE AT DIFFERENT COMPRESSION RATIOS

                  Narasimha Bailkeri1, Krishna Prasad2, Shrinivasa Rao B.R3
        1
          Dept. of Mechanical Engg, NMAM Institute of Technology, Nitte, Karnataka, India
        2
          Dept. of Mechanical Engg, NMAM Institute of Technology, Nitte, Karnataka, India
        3
          Dept. of Mechanical Engg, NMAM Institute of Technology, Nitte, Karnataka, India


ABSTRACT

        The present work involves some experimental investigation on multiple spark plug engines.
A new dual spark ignition engine has been developed by introducing two spark plugs at different
locations and the experiments are conducted at different load conditions and at three different
compression ratios. The results are compared with that of a single plug operation. The results have
shown that performance of dual plug engine is comparatively better than the conventional single plug
ignition engine under all three compression ratios. The results have shown considerable
improvement in thermal efficiency, and reduction in HC and CO emissions in dual plug mode of
operation. However, there is a small increase in NOX emission.
        Effect of compression ratio in dual plug engine system has not been investigated in detail so
far with respect to engine performance and exhaust emissions. In this paper it is observed that there
is an optimum compression ratio which gives the best performance with respect to the above
parameters due to ill effects of combustion knock at higher compression ratios.

Key Words- Engine performance, Compression ratio, Exhaust emission, Dual plug SI engine.

1. INTRODUCTION

         In spark ignition engines as the load decreases engine power reduces by throttling. Due to
throttling, the initial and maximum compression pressures decrease which will lead to charge
dilution with the residual gases, which in turn affects the formation of self propagating flame nucleus
and prolongs the ignition delay. Though this difficulty can be overcome by using slightly rich
mixture at part loads (10-20% richer than stoichiometric mixture) but still it is difficult to avoid after
burning. Hence poor part load performance and necessity of mixture enrichment are among the main
disadvantages of spark ignition engines, which cause wastage of fuel and increased pollutants
concentration in the exhaust [1]. Several techniques are employed to extend the lean limit and to
improve the part load performance of SI engines. These techniques include spark plugs of different

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0976 – 6480(Print), ISSN 0976 – 6499(Online) Volume 4, Issue 5, July – August (2013), © IAEME

designs like use of higher power, higher energy or longer duration discharges and ignition systems
that initiate the main combustion process with a high temperature reacting jet namely plasma jet and
flame jet ignition systems and multiple gap spark plugs or use of multiple ignition systems [2].
Among these techniques, multiple ignition system finds its way with respect to rapidity and
completeness of combustion, as rapidity and completeness of combustion are the two basic
requirements for healthy operation of an SI engine.
         Multiple spark plug engines initiate flame propagation at two or more number of points in the
combustion chamber depending on the number of spark plugs employed. If two plugs are employed
the flame front travels from two ignition centers in the cylinder and the effective flame travel
distance for each flame is reduced which improves the knock resistance. The concept of dual plug
spark ignition is under consideration for more than last three decades. Several experimental studies
were made in the area of dual ignition engines regarding optimization of spark plug location and to
prove their efficient operation at part loads, extended exhaust gas recirculation (EGR) tolerance and
relatively clean burning compared with single spark ignition systems [3- 11]. H Kuroda et al., [3] for
example conducted experimental study on Nissan NAPS-Z engine to optimize combustion chamber
shape and spark plug location to equalize flame propagation from two plugs. It was found that
combustion with 20% EGR was almost same as that of a conventional engine without EGR. They
observed marked improvement in fuel economy, reduction in HC and NOX emissions due to fast
burning under heavy EGR. Masonari Harada et al., [5] also conducted experiments on Nissan NAPS-
Z engine and obtained similar results. Peter O Witze [9] conducted experimental studies to
understand the trade off that exists between spark plug locations and swirl rate. It was concluded that
use of two spark plugs along a common diameter near the mid radius position gives significantly
faster rate of combustion, than single plug at the same radial location. Increased c-b-c fluctuations
were observed with increased burning duration and decreased swirl level. A Ramtilak et al., [12]
conducted experimental investigations on 150 DTS-i (digital twin spark ignition) engine and noticed
the benefits like higher compression ratio, improved fuel economy, increased specific output per
litre, torque, better driveability and reduced emission levels due to rapid combustion brought by twin
spark plugs. It was shown experimentally that dual ignition system is advantageous in engines
operating under the “conditions unfavorable to ignite” like poor fuel-air mixture quality or with
significant misfiring. [13, 14]. Amer Ahmad et al., [15] investigated the effect of charge motion,
namely tumble on the burn characteristics of the Chrysler Hemi S.I engine. They used CFD
simulations (AVL-FIRE CFD code) to evaluate the effect of piston top and number of spark plugs. It
was observed that dual plug operation offers considerable improvement on burn characteristics and
engine performance. F Bozza et al., [16] developed a twin spark S.I engine with variable valve
timing (VVT) device. Both experimental and theoretical analyses were made and a quasi
dimensional model was used to find the proper combination of VVT device position (EGR level) and
spark advance. Ismail Altin et al., [17] developed a thermodynamic based cycle simulation of twin
spark engine to investigate the effect of spark plug locations.
          This brief review indicates that use of dual ignition sources increases the rate of combustion
which will result in rapid completion of combustion process. Thus the dual spark plug operation
tends to improve the engine efficiency and results in smoother engine operation, and reduced
pollutants concentration in the exhaust. In the present work effect of compression ratios is analyzed
in detail with respect to engine performance and emission parameters.

2. ENGINE MODIFICATION, EXPERIMENTAL SET UP AND TEST PROCEDURE

       Experiments were conducted on a four-stroke air cooled petrol engine with necessary
modifications to accommodate dual plugs. The engine specifications are given in Table 1. Apart
from the original spark plug ‘A’, whose diameter is 14mm, one more 14mm hole is threaded in the

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International Journal of Advanced Research in Engineering and Technology (IJARET), ISSN
0976 – 6480(Print), ISSN 0976 – 6499(Online) Volume 4, Issue 5, July – August (2013), © IAEME

engine cylinder head to fit the second spark plug ‘B’. The second spark plug B is located
diametrically opposite to spark plug ‘A’ as shown in Fig1. The spark plug ‘A’ is made to fire as per
the original ignition timing set by the manufacturer. Spark timing of plug ‘B’ can be varied by using
a spark timing variation unit fitted on to the engine shaft. The position of TDC and graduations of 10
are marked on either side up to 400 before and after TDC.
        Since similar ignition timings of both the plugs provide better results as compared with
advanced or retarded spark timing of the plug B with respect to the plug A [18], in this study similar
ignition timing for both the plugs is employed.




                    Fig 1. View of cylinder head with provision for Dual plug


                           Table 1 Specifications of the Engine
          Number of cylinders                             1
          Number of strokes                                   4
          Cooling                                             Air cooled
          Rated power                                         6kW @7500rpm
          Cylinder diameter                                   53 mm
          Stroke length                                       45 mm
          Compression ratio                                   9.5
          Orifice diameter                                    13 mm
          Dynamometer Type                                    Eddy current
          Dynamometer arm length                              185 mm
          Coefficient of discharge for air flow orifice       0.64

        To measure the air flow rate, an orifice tank is used. The pressure difference across the orifice
is indicated by a manometer fitted to the tank. By noting the difference in water level in the two
limbs of manometer the air flow rate can be calculated. The rate of fuel consumption is computed by
recording the time taken for consumption of 10 cc of fuel.


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         The engine is directly coupled to an eddy current dynamometer. The dynamometer has the
capacity to absorb the maximum power that can be produced by the engine. The brake power
produced by the engine is measured by the dynamometer and is displayed on a digital load indicator.
         A computerized engine test rig is used. It is fitted with sensors to measure mean effective
pressure, the exhaust gas temperature, the rate of fuel consumption, the air flow rate, engine speed
and a load sensor on the dynamometer unit. A PC loaded with necessary engine software is
connected to the control and measuring unit of the engine. The data from the sensors is directly fed
to the computer and the engine software processes all the information required like speed, load,
torque, brake power, indicated mean effective pressure, brake mean effective pressure, indicated
power, air consumption, fuel consumption, air-fuel ratio, specific fuel consumption, mechanical
efficiency, brake thermal efficiency, indicated thermal efficiency, volumetric efficiency and exhaust
gas temperature. The results are stored for one minute of test period, and the recorded values are
averaged for this 1 minute of test period under steady state conditions.
An AVL Digas 444 exhaust gas analyzer is used to indicate the value of CO in %, NOx and UBHC
in ppm present in the exhaust gas.
         The experiments were conducted at 3000 rpm. The original spark plug ‘A’ was made to ignite
at its standard ignition timing of 260 BTDC and the spark timing of other spark plug ‘B’ is also set
for 260 BTDC to ensure that both the plugs fire simultaneously. The test was conducted separately in
single plug and dual plug mode of operation with pure gasoline as fuel at different load conditions
and compression ratios. The different load conditions were 0%, 25%, 50%, 75% and 100% of the full
load capacity of the engine at 3000 rpm and different compression ratios were 7.5, 8.5 and 9.5. The
schematic diagram of the engine test set up is shown in Fig 2.




                            Fig.2. Schematic layout of engine test set up




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International Journal of Advanced Research in Engineering and Technology (IJARET), ISSN
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3. RESULTS AND DISCUSSIONS

        Results obtained from the experiments conducted with single and dual spark plugs at three
different compression ratios, using pure gasoline are presented in Figures 3 to 9

3.1 Brake Thermal Efficiency v/s Load
        Fig 3 shows variation of brake thermal efficiency with load for three compression ratios. The
following observations can be made:
     • Brake thermal efficiency is maximum for the compression ratio of 8.5
     • Variation becomes more pronounced as the load increases.
     • Efficiencies in all the cases are higher in the dual plug mode.
        Theoretically cycle efficiency should increase with increase in compression ratio. However
the efficiency drops due to knocking at higher compression ratios.
        Thus at a compression ratio of 7.5 efficiency is low. Again at a compression ratio of 9.5,
efficiency drops due to knocking. Hence there is an intermediate compression ratio (of 8.5) which
shows the best efficiency, which corresponds to the best knock free compression ratio for this
engine. In all the cases the dual plug mode gives relatively better efficiency, about 2% higher at the
best compression ratio of 8.5 at full load. The knocking is observed to be more severe with the dual
plug mode at compression ratio of 9.5, with a greater penalty on thermal efficiency. This may be
attributed to the more severe cylinder and exhaust valve temperature conditions with dual plug
system due to faster combustion, which could promote severe knock near the exhaust valve centre.
Occurrence of combustion knock was identified by the characteristic audible knocking sound
emanating from the engine head.




                    (a)                                                   (b)
   Fig.(3) Variation of Brake Thermal Efficiency with Load at different Compression ratios
                                (a) Single plug (b) Dual plug

3.2. Brake Specific Fuel Consumption v/s Load
       Fig 4 shows variation of BSFC with load. Here BSFC is compared for three different
compression ratios. BSFC curve is the mirror image of the efficiency curve. Thus it is lowest for the
compression ratio of 8.5 for the reasons mentioned earlier for thermal efficiency.
           • BSFC continues to decrease with increase in load in all three cases.
Again in all cases BSFC is lower in the dual plug mode relative to the single plug mode.

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                       (a)                                                   (b)
  Fig. (4)Variation of Brake Specific Fuel Consumption with Load at different Compression
                           ratios (a) Single plug   (b) Dual plug

 3.3. Volumetric Efficiency v/s Load
        Fig 5 shows the variation of volumetric efficiency with load for both single and dual plug
mode of operations.
Following observations are made:
    • Volumetric efficiency is maximum at compression ratio of 8.5
    • It increases with load in all cases due to quantity type governing in carburetted engines.
    Volumetric efficiency in general, is slightly less in the case of twin plug mode, relative to single
plug mode since cylinder wall temperature increases due to faster combustion, resulting in higher
combustion temperatures.
    At optimum compression ratio (8.5) owing to higher thermal efficiency, residual gas temperature
is less, hence volumetric efficiency is high. At lower compression ratio (7.5) due to increased
clearance, volumetric efficiency is low. At compression ratio of 9.5, due to higher cylinder wall and
residual gas temperature, volumetric efficiency is less.




                       (a)                                                 (b)
     Fig.(5) Variation of Volumetric Efficiency with Load at different Compression ratios
                                (a) Single plug (b) Dual plug

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3.4. Fuel-air equivalence ratio v/s Load
        Variation of fuel-air equivalence ratio (Ø) with load is shown in Fig.6, both for single and
dual plug mode of operation. The air flow rate is controlled by throttle position and the fuel flow is
controlled by carburetor characteristics. The fuel-air equivalence ratio calculated from the air and
fuel flow rate shows that mixture required is fairly rich at lower load, gradually tending towards
stoichiometric strength as load increases to full load.




                       (a)                                               (b)
   Fig.(6) Variation of Fuel-air equivalence ratio with Load at different Compression ratios
                                 (a) Single plug (b) Dual plug

3.5. UBHC Emission v/s. Load
       Fig 7 shows the variation of UBHC emission with load for both single and dual plug mode of
operations.




                       (a)                                               (b)

Fig.(7) Variation of UBHC emission with Load at different Compression ratios (a) Single plug
                                      (b) Dual plug



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International Journal of Advanced Research in Engineering and Technology (IJARET), ISSN
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3.6. CO emission v/s Load
        In Fig 8 carbon monoxide emission is compared for the three compression ratios for different
loads and the following observations are made.
    • Minimum CO emission is observed for the compression ratio of 8.5 due to more complete
        combustion
    • CO emission decreases with load and reaches minimum at 75% load and then again increases
        in all three cases. This increase is attributed to richer mixture at full load.
    In the dual plug mode, CO emission is found to be less in all the cases, due to higher combustion
temperature which promotes oxidation of CO.




                   (a)                                                     (b)

          Fig.(8). Variation of CO emission with Load at different Compression ratios
                                  (a) Single plug (b) Dual plug


3.7 NOx Emission v/s Load
    The Nitrogen Oxide concentration v/s load is plotted for all three compression ratios in Fig 9.
From the graphs following observations are made.
    • Higher NOx emission is observed for the compression ratio of 9.5 due to higher combustion
       temperature.
    • The NOx emission increases from no load and reaches a maximum at 75% load and then
       decreases in all three cases.
       Rapid combustion of the fuel increases temperature inside the engine cylinder. At high
temperature nitrogen reacts with oxygen to form its oxides. At full load, the increase in fuel-air ratio
decreases NOx emission.
       In the dual plug mode, due to rapid combustion, gas temperature increases which increase the
NOX formation, relative to single plug mode.




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                         (a)                                               (b)

          Fig.9 Variation of NOX emission with Load at different Compression ratios
                                (a) Single plug (b) Dual plug

4. CONCLUSION

    Following conclusions are made based on the experimental results
       i) Brake Thermal Efficiency under dual plug operation is around 2% more than that at single
plug mode, at the optimum compression ratio of 8.5. This corresponds to a minimum BSFC of 0.35
kg/kWh.
       ii) Volumetric efficiency under dual plug operation is around 2% less than that in the single
plug operation, at the compression ratio of 8.5.
       iii) At this optimum compression ratio, exhaust emissions of CO and UBHC are also reduced
       iv) UBHC emissions reduced by 15% and CO emissions by 17% under dual plug mode at full
load.
       v) NOX emissions increased by about 40% under dual plug mode at compression ratio of 8.5, at
full load.
          It is experimentally observed that for any given engine configuration there exists an optimum
compression ratio, which gives best performance with respect to efficiency and exhaust emissions, in
dual plug mode. Hence it is suggested that care must be taken to select a compression ratio in the
design stage, so as to obtain maximum benefits from the dual plug combustion engines.

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0976 – 6480(Print), ISSN 0976 – 6499(Online) Volume 4, Issue 5, July – August (2013), © IAEME

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