Combustion Chamber Design for Lean Burn LPG Engines

Combustion Chamber Design for Lean Burn LPG Engines

Suresh Devunuri Sampath Vitta

Department of Mechanical Engineering School of Mechanical Engineering and Building Sciences MVSR Engineering College VIT University

Hyderabad, India Vellore,India

AbstractLean operation is an attractive operational method to increase thermal efficiency and to decrease exhaust emissions and fuel consumption. Gaseous fuels as clean, economical and abundant fuels can improve the lean operating limits. Liquefied petroleum gas (LPG) is one of the members of natural gases and declared as the cleaner fuel.

Lean operation of homogeneous-charge spark-ignited engines reduces peak combustion temperatures, thereby reducing NOx emissions. Lean operation is normally restricted by the air-fuel ratio above which ignition is impossible, or combustion is incomplete. Operation under lean conditions also reduces the mixture burning rate, which can lead to increased spark advance and lower thermal efficiency.

In this paper, in order to increase the burning rate under lean air-fuel ratios combustion chamber shapes have been developed. The combustion chamber designs in such that they develop squish velocity inside combustion chamber, which generates increased levels of turbulence just before ignition and during the early phase of combustion. This increased burning rate gives the engine to operate with a smaller spark advance under lean conditions, thereby increasing the lean-limit of operation and increasing the thermal efficiency. The additional turbulence levels generated with the squish-jet type of combustion chamber is improves the completeness of combustion, thereby reducing unburned hydrocarbon emissions.

This paper presents three combustion chamber designs for lean operating LPG-SI engine, aimed at optimizing the squish velocity generation at optimized compression ratio as 11, in the mixture just before ignition and uses to increase the burning rate during lean operation, thereby increases the thermal efficiency and to decreases the exhaust emissions and fuel consumption.

Key Words Lean burn, Spark Ignition, LPG, Thermal Efficiency, Economic Fuel Control, Emission Control.

1. INTRODUCTION

   To meet future regulations for stringent emissions, LPG (Liquefied Petroleum Gas) fueled spark ignition engines are being used (1).A stoichiometric LPG fueled engine has limited applications due to high exhaust gas temperatures and a lower thermal efficiency. However, the lean burn technology implemented to overcome these difficulties.

   Before 1970s a very few experimental works were made to study on the lean combustion technology in SI engines. This technology was studied first during 1908 to demonstrate the advantages of higher thermal efficiency (2).Later on the need for emission control and fuel

   economy improvement became evident and hence the lean combustion technology shows to offer the lower emissions, higher thermal efficiency and also improves the fuel economy (3).

   The principal benefits of this operating technique are a reduction in greenhouse gas emissions and NOx emissions. Lean operation is normally restricted by the air- fuel ratio above which combustion is incomplete. A disadvantage of lean operation is that the burning rate can be significantly lower than with stoichiometric combustion. The reduction in burning rate increases the overall combustion duration and also leads to low flame velocities(4). For a successful implementation of the lean burn technology to decrease the exhaust emissions and increase the thermal efficiency, burn rate enhancement is necessary. A number of design variables such as spark plug location, intake port configuration and combustion chamber shape have been shown to influence the burn rate. Among these, squish-jet motion by using combustion chamber shape is having greater influence on the burn rate (5).

2. LITERATURE REVIEW

   M.A.CEVIZ (6) investigated on the cyclic variations on liquefied petroleum gas (LPG) and gasoline lean burn spark ignition (SI) engine in order to reduce the cyclic variation in the SI engine; they use LPG as a fuel for the SI engine in terms of lean operation. Finally they concluded that higher laminar flame speed of LPG and good mixing of gaseous fuels in air causes the decreases in cyclic variation and LPG more suitable for lean operations in SI engines.

   A.V. SITA RAMA RAJU (7) presented a paper which describe the effect of intensified swirl and squish on the performance of lean burn engine operated on liquefied petroleum gas (LPG).In order to produce the swirl and squish motion they masked the intake valve and provided swirl grooves on the piston crown. They found that combined swirl and squish configuration resulted in a small extension of the lean misfire limit and no significant change in the performance.

   JOSEPH SHANKAL (8) observed the flame propagation in an liquefied petroleum gas (LPG) lean burn spark ignition engine, to investigate the combustion characteristics of a heavy duty liquefied petroleum gas lean burn engine. They varied swirl ratio and piston cavity configuration to investigate their effects on combustion and engine performance. Finally they concluded that with decreases in

   mixture strength flame speed and exhaust temperature also decreases and by increasing the squish area burn duration of flame also decreases.

   LIGUANG LI (9) proposed a study on liquefied petroleum gas (LPG) lean burn in a motor cycle spark ignition engine. They compared the lean burn limits with the parameters such as engine speed; compression and advanced spark ignition etcare tested. They concluded that the emission of liquefied petroleum gas engine is significantly reduced and lean burn limit can be improved by using the higher compression ratio and spark ignition.

   SULAIMAN (10) presented a paper to analyse the performance of single cylinder spark ignition engine running with liquefied petroleum gas (LPG) as a fuel. They found that decreased on power output up to 4 % compared to petrol due to volumetric efficiency and specific fuel consumption is reduced.

   O.BADR (11) reported a paper on parametric study on the lean misfiring and knocking limits of gas fueled spark ignition engines. They used three different criteria for defining the engine lean limit. Finally they concluded that as the compression ratio increased, the misfiring limit of liquefied petroleum gas and air mixture slightly decreased. ERIC KASTANIS (12) presented a paper on The Squish- Jet Combustion Chamber for Ultra-Lean Burn Natural Gas Engines. In this paper they use the blow in piston concept to generate the squish jet motion. Finally they concluded that squish jet design operated with more advanced MBT ignition timing than the blow in piston chamber.

   MIKIO FURUYAMA AND XU BO YAN (13) reported a

   paper on Mixing Flow Phenomena of Natural Gas and Air in the Mixer of a CNG Vehicle. In this paper they accomplished visualization by means of the Schlieren method in a two dimensional flow channel model of a CNG engine mixer.

3. PRESENT WORK

   The basic piston which is used in the desiel engine with compression ratio(C.R) 17.5 is shown in figure 1, and for this compression ratio swept volume and clearence volume produced by piston are following,

   In this proposed work a four stoke, single cylinder, air-cooled, stationary diesel engine is modified to run as a spark ignited gas engine by replacing the injector and fuel pump by a spark plug, a gas carburettor and an ignition system. Specifications of the test engine are shown in Table 1.In this present work squish jet piston concpt used. This concept uses bowl in piston to generate a squish motion and series of jets of it directed towards the centre of combustion chamber. The basic piston which is used in the desiel engine with compression ratio(C.R) 17.5. The volume of the bowl in the piston was simultaneously changed without altering its clearance volume in order to maintain the compression ratio 11. The volume of the bowl in piston based on the theoretical calculations (14)

   parameters	values
   Cylinder bore (cm)	87
   Stroke length (cm)	110
   Compression ratio	11
   Volumetric efficiency (%)	85
   Speed (rpm)	1500
   Power (kw)	4.4

   Table1. Specifications of engine

   =

   Fig.1 Basic pistion

   = +

   cm3

   swept volume,

   Vs = (/4)×87.52×110 = 661452.5 mm3= 661.5

   clearance volume,

   Vc= 66.15 cm3

   In this present work the deisel engine is modified in to spark ignition (S.I) engine and simultaniously redusing the compression ratio to 11. For this compresion ratio the required clearence volume is 66.15 cm3. To achieve this clearence volume and different squish velocities the piston bowl is modified as shown below by varying the squish land, piston bowl depth and clearence hight as shown in Model 1, model 2, model 3 and the corresponding values are tabulated in the table 2.

   The squish velocity is determined by using the following formula,

   = [()

   ] +

   Vsq = squish velocity (m/s) DB = Bowl diameter (m) B = Cylinder Bore (m)

   VB = Volume of the piston bowl (m3)

   Ac = Cross sectional area of the cylinder= (B2/4) in m2 z = c+Z

   Where, c=clearance height Z=l+a-s

   l=connecting rod length=220mm a=crank radius=55mm

   s=distance between the crank axis and the piston pin axis (m)

   = + ( )/

   =crank angle

   Sp = Instantaneous piston speed

   Fig.2 Mode 1(P1)

   = [ +

   ]

   ( )

   Sp=mean piston speed=2LN/60 (m/s) L =stroke length(110mm)

   N=crank speed (1500rpm) R= l/a =4

   Fig.3 Model 2 (P2)

   Fig4. Model 3 (P3)

   Table 2:

   Model	Clearance height (mm)	Squish land (mm)	Depth of bowl (mm)	Volume of bowl (cc)	Squish velocity (m/sec)
   basic	0.846	17.75	25	35	17.79
   P1	1.846	8.05	24	55.28	4.5
   P2	2.52	12.75	25	45.91	5.5
   P3	0.846	7.05	25	61.07	6.5

4. EXPEREMENTAL SETUP AND EXEPEREMENTATION

   The schematic diagram of the experimental setup is shown in figure. The modified engine was coupled to an eddy current dynamometer and LPG was supplied from a cylinder into the venturi of the gas carburetor through a pressure regulator, orifice meter, surge tank and a needle valve. The pressure drop across the orifice meter was measured with a micro-manometer to calculate the gas flow rate.

   Arrangements were made to measure the temperature and pressure of LPG before it enters into the orifice meter.

   Fig.5 Experimental Setup

   1. Engine 2 Dynamometer 3.Dynamometer controller 4.LPG cylinder 5.Pressure regulator 6. Vaporizer 7.Gas flow meter 8.Air flow meter 9.air drum 10.venturi 11.TDC encoder 12.Analog to Digital converter 13.Computer 14.HC/CO analyzer 15.Nox analyzer

   Air flow rate was measured by using a flow meter. Pressure-crank angle data was acquired on a personal computer using a flush mounted piezoelectric pressure transducer. This data was processed by software to find combustion parameters. Arrangement was made for measuring the spark timing with the help of a stroboscope. A NDIR (non-dispersive infrared) gas analyzer was used for the measurement of HC & CO in the exhaust. A CLD (chemiluminiscence device) analyzer which works on the chemiluminiscence principle was used for measuring the NO concentration in the engine exhaust. Some relevant properties of LPG are given in Table 3.

   Tests will conduct with LPG, using three different types of conventional bowl-in piston combustion chambers as shown in models (1, 2, 3). Initially, test conducted for finding out the MBT spark timing at different air fuel ratios and then performance test was conducted for three pistons at 25 % throttle and 100% throttle by using MBT spark timing.

   Table 3.Properties of LPG

   Parameter	Quantity	Units
   Composition: Butane Propane	70 30	% by volume % by volume
   Density	2.26	kg/m3
   Calorific Value	47731	kJ/kg
   Minimum Ignition Temperature	410	°C
   Octane Number: Research Motor		99 110
   Flame Speed	0.37	m/s
   Flammability Limits: Rich Lean	0.4019 1.9140	Excess air ratio

5. RESULTS AND DISCUSSIONS

   Performance Parameters:

   Brake Power (KW)

   1. Brake Power:

      4

      3

      2

      1

      P1

      P2 P3

      0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3

      equivalence ratio

      0

      Brake Power (KW)

      Fig.6 At 25% throttle

      5

      4

      3

      2

      1

      0

      P1

      P2 P3

      0.5 0.6 0.7 0.8 0.9

      equivalence ratio

      Fig.7 At 100% throttle

      Figures 6,7 shows that variation of brake power with equivalence ratio at 25% and 100% throttle openings respectively. The model P3 gives maximum brake power

      4.25 KW at equivalence ratio of 0.695 for 100% throttle opening.

      bth (%)

   2. Brake Thermal Efficiency:

      30

      25

      20

      15

      10

      5

      0

      P1

      P2 P3

      0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3

      equivalence ratio

      bth (%)

      Fig.8 At 25% throttle

      35

      30

      25

      20

      15

      10

      5

      0

      P1

      P2 P3

      0.5 0.6 0.7 0.8 0.9

      equivalence ratio

      Fig.9 At 100% throttle

      Brake thermal efficiency variation with equivalence ratio at 25% and 100% throttle are as shown in figures 8 and 9. The model P3 gives maximum brake thermal efficiency 29.56% at equivalence ratio of 0.695 for 100% throttle opening compared to other models.

      BSFC (Kg/KWh)

   3. Brake Specific Fuel Consumption:

      1

      0.8

      0.6

      0.4

      0.2

      0

      P1

      P2 P3

      0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3

      equivalence ratio

      Fig.10 At 25% throttle

      1.4

      1.2

      1

      0.8

      0.6

      0.4

      0.2

      0

      P1

      P2 P3

      0.5 0.6 0.7 0.8 0.9

      equivalence ratio

      BSFC (Kg/KWh)

      Fig.11 At 100% throttle

      Figures 10,11 shows that variation of brake specific fuel consumption with equivalence ratio at 25% and 100% throttle openings respectively.

      Emission Parameters:

      HC Emission (PPM)

      1. HC Emissions:

         1000

         800

         600

         400

         200

         0

         P1

         P2 P3

         0.5 .6 0.7 0.8 0.9 1 1.1 1.2 1.3

         equivalence ratio

         HC Emission (PPM)

         Fig.12 At 25% throttle

         700

         600

         500

         400

         300

         200

         100

         0

         P1

         P2 P3

         0.5 0.6 0.7 0.8 0.9

         equivalence ratio

         Fig.13 At 100% throttle

         Hydrocarbon emission variation with equivalence ratio at 25% and 100% throttle are as shown in figures 12 and 13. The model P3 gives less HC emissions at 25% and 100% throttle openings compared to other models. We can see that as air fuel mixture becomes rich HC emissions increases.

         CO Emission (%Vol)

      2. CO Emissions:

         2.5

         2

         1.5

         1

         0.5

         P1

         P2 P3

         0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3

         equivalence ratio

         0

         CO Emission (% Vol)

         Fig14. At 25% throttle

         0.14

         0.12

         0.1

         0.08

         0.06

         0.04

         0.02

         0

         P1

         P2 P3

         0.5 0.6 0.7 0.8 0.9

         equivalence ratio

         Fig15. At 100% throttle

         Corbonmonoxide emissions variation with equivalence ratio at 25% and 100% throttle are as shown in figures 14 and 15. The model P3 gives less CO emissions at 25% and 100% throttle openings compared to other models. We can see that as air fuel mixture becomes rich CO emissions increases.

         NOx Emission (PPM)

      3. NOx Emissions:

   1000

   900

   800

   700

   600

   500

   400

   300

   200

   100

   0

   P1

   P2 P3

   0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3

   equivalence ratio

   Fig.16 At 25% throttle

   0.5 0.6 0.7 0.8 0.9

   equivalence ratio

   0

   P1

   P2 P3

   2000

   1500

   1000

   500

   NOx Emission (PPM)

   Fig.17 At 100% throttle

   Nitric oxide emissions variation with equivalence ratio at 25% and 100% throttle are as shown in figures 16 and 17. The model P1 gives less NOx emissions at 25% and 100% throttle openings compared to other models. We can see that as air fuel mixture becomes rich NOx emissions decreases.

   Combustion Parameters:

   spark timing (degrees)

   1. Spark Timing:

   50

   40

   30

   20

   10

   0

   P1

   P2 P3

   0.5 0.6 0.7 0.8 0.9 1 1.1 1.2 1.3

   equivalence ratio

   Spark timing (Degrees)

   Fig.18 At 25% throttle

   50

   40

   30

   20

   10

   0

   P1

   P2 P3

   0.5 0.6 0.7 0.8 0.9

   equivalence ratio

   Fig.19 At 100% throttle

   Spark timing (MBT) variation with equivalence ratio at 25% and 100% throttle are as shown in figures 18 and 19. With increase in squish velocity the combustion duration and ignition delay decreases.

6. CONCLUSIONS

   This experimental work was aimed at investigating the optimum squish velocity for an optimum compression ratio 11 based on performance and emissions of a lean burn operated on liquefied petroleum gas (LPG). Based on results obtained the following conclusions are drawn.

   1. No significant improvement in power output and thermal efficiency, the model P3 gave best thermal efficiency and emission characteristics at 25%, 100% throttle compared to the model P2 and model P1.So the maximum possible squish velocity for compression ratio 11 is (6.56 m/sec) the optimum squish velocity corresponding to model P3.

   2. With increase in squish velocity the lean burn limit increases at 25%, and 100% throttle openings. For model P3 lean burn limits are 0.53 and 0.52 at 25% and 100% throttle openings respectively.

   3. With increase in squish velocity the combustion duration and ignition delay decreases.

7. REFERENCES

1. Goto, S., Shakal, J. and Daigo, H., LPG Engine Research and Development in Japan, International Congress on Transportation Electronics (Convergence 96), 1996.

2. Hopkinson B, the effect of mixture strength and scavenging on thermal efficiency,Proc.Inst.Mech.Engr.,417,1908

3. Jakson MW The influence of air-fuel ratio, spark timing, and combustion chamber deposits on exhaust hydrocarbon emissions, SAE technical progress series,

   Vehicle emission Vol 6 PP 175-194, 1964

4. Evans, R. L., 2007, Lean-Burn Spark-Ignited Internal Combustion Engines, Lean Combustion Technology and Control, D. Dunn- Rankin, ed., Academic, New York, pp. 95120.

5. Belaire, R.C., Davis, G.C., Kent, J.C. and Tabaczynski,R.J., Combustion Chamber Effects on Burn Rates in a High Swirl Spark Ignition Engine, SAE paper 830335(1983).

6. M. A. Ceviz and F. Yuksel. Cyclic variation on LPG and gasoline fuelled lean burn SI engine. Elsevier, Renewable Energy vol 3, 2005, pp 1950-1960.

7. A. V. Sita Rama Raju, A. Ramesh and B. Nagalingam, 2000,Effect of Intensified Swirl and Squish on the Performance of a Lean Burn Engine Operated on LPG.

   SAE paper no 2000-01-1951.

8. Daeyup lee, Joseph shakal, shinichigoto and hitoshiishikawaObservation of Flame Propagation in LPG Lean Burn SI Engine, SAE paper no 1999-01-0570.

9. LiguangLiA Study of LPG Lean Burn for a Small SI Engine SAE paper no

   2002- 01 -2844.

10. SulaimanPerformance of Single Cylinder Spark Ignition Engine Fueled by LPG. . Elsevier, Renewable Energy vol 53,579-585, 2013.

11. O.badr N. alsayed and M. manaf a parametric study on the lean misfiring and knocking lean limits of gas-fueled spark ignition engines.

12. Eric kastanisand Robert L. Evans the Squish-Jet Combustion Chamber for Ultra-Lean Burn Natural Gas Engines SAE paper no 2011-24-0112.

13. Mikiofuruyama and xuboyan Mixing Flow Phenomena of Natural Gas and Air in the Mixer of a CNG Vehicle SAE paper no 981391.

14. John.B.Heywood Internal combustion engine fundamentals McGraw Hill Book Co., 1988.