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EXPERIMENTAL STUDY OF THE

EFFECT OF IN CYLINDER AIR SWIRL

ON DIESEL ENGINE PERFORMANCES.L.V.PRASAD

1

, Prof .V.PANDURANGADU2V.V.PRATIBHA BHARATHI

3

V.V.NAGA DEEPTHI4

1

Department of Mechanical Engineering, Sri Sai College of Engineering &Technology, Anantapur, A.P India

2 & 4

Department of Mechanical Engineering, JNT University, Anantapur, A.P., India

3

Department of Mechanical Engineering, SKD, Gooty, A.P., India

ABSTRACT

In this present work a study about influence of the air swirl in the cylinder upon the performance andemission of a single cylinder diesel direct injection engine is presented. In order to achieve the different swirl

intensities in the cylinder, three design parameters have been changed: the cylinder head, piston crown, and inlet

duct. In this way, the piston crown is modified i.e. alteration of combustion chamber to enhance the turbulencein the cylinder. This intensification of the swirl is done by cutting grooves on the crown of the piston. In this

work three different configurations of piston i.e. in the order of number of grooves 3,6,9 were used to intensify

the swirl for better mixing of fuel and air and their effects on the performance and emission are recorded

Key words: Diesel engine , air swirl ,cylinder, efficiency ,emissions

INTRODUCTION

Internal combustion engines have been a relatively inexpensive and reliable source of power for

applications ranging from domestic use to large scale industrial and transportation applications for most of the

twentieth century. DI Diesel engines, having the evident benefit of a higher thermal efficiency than all other

engines, have served for both light- duty and heavy-duty vehicles

The in-cylinder fluid motion in internal combustion engines is one of the most important factors

controlling the combustion process. It governs the fuel-air mixing and burning rates in diesel engines. The fluid

flow prior to combustion in internal combustion engines is generated during the induction process and

developed during the compression stroke [1,2]. Therefore, a better understanding of fluid motion during theinduction process is critical for developing engine designs with the most desirable operating and emission

characteristics [3]

To obtain a better combustion with lesser emissions in direct--injection diesel engines, it is necessary to

achieve a good spatial distribution of the injected fuel throughout the entire space [4]. This requires matching ofthe fuel sprays with combustion chamber geometry to effectively make use of the gas flows. In other words,

matching the combustion chamber geometry, fuel injection and gas flows is the most crucial factor for attaining

a better combustion [5]. In DI diesel engines, swirl can increase the rate of fuel-air mixing [6], reducing the

combustion duration for re-entrant chambers at retarded injection timings. Swirl interaction [7] with

compression induced squish flow increases turbulence levels in the combustion bowl, promoting mixing. Since

the flow in the combustion chamber develops from interaction of the intake flow with the in-cylinder geometry,

the goal of this work is to characterize the role of combustion chamber geometry on in-cylinder flow, thus the

fuel-air mixing, combustion and pollutant formation processes.

It is evident that the effect of geometry has a negligible effect on the airflow during the intake stroke

and early part of the compression stroke. But when the piston moves towards Top Dead Centre (TDC), the bowl

geometry has a significant effect on air flow thereby resulting in better atomization, better mixing and better

combustion. The re-entrant chamber without central projection and with sharp edges provides higher swirl

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number than all other chambers[8]. This higher swirl number reduces the soot emission at the cost of higher Nox

level.

EXPERIMENTAL SETUP AND PROCEDURE

In the present work the effects of air swirl in combustion chamber are experimentally studied on

performance of single cylinder light duty direct injection diesel engine. The experiments were conducted on asingle cylinder Kirloskar make direct injection four stroke cycle diesel engine. The general specifications of the

engine are given in Table-1. Water cooled eddy current dynamometer was used for the tests. The engine isequipped with electro-magnetic pick up, piezo-type cylinder pressure sensor, thermocouples to measure the

temperature of water, air and gas, rotameter to measure the water flow rate and manometer to measure air flow

and fuel flow rates, The smoke density is measured with a Bosch Smoke meter. To intensify the air swirl

numbers of elliptical grooves were made on the piston crown .The different types of piston configurations which

were tested in diesel engine are shown in fig. 1

GP 3 GP 6 GP 9

Fig.1 Different types of configurations of piston crowns

N P - Normal Piston ; GP 3 - Piston with 3 grooves

GP 6 - Piston with 6 grooves; GP 9 - Piston with 9 grooves

Specifications of Diesel Engine Used for Experimentation

Item Specification

Engine power 3.68 kW

Cylinder bore 80 mm

Stroke length 110 mm

Engine speed 1500 rpm

Compression ratio 16.5:1

Swept volume 553 cc

RESULTS AND DISCUSSIONS

BRAKE THERMAL EFFICIENCY

The brake thermal efficiency with brake power for different configurations are compared with the

normal engine configuration and is shown in fig. No.2. The brake thermal efficiency for normal piston at full

load is 29.8%. It can be observed that the engine with G P 9 and G P 6 give thermal efficiencies of 31.8 % and

30.8%, respectively, at full load. It is observed that there is a gain of 6 % with G P 9 compared to normal

engine. The thermal efficiencies of G P 6 and G P 3 are lower compared to that G P 9. From Fig.1, it wasinferred that the brake thermal efficiencies were increasing with an increase in brake power for configurations

that were under consideration. These configurations were found to offer better thermal efficiencies than the

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normal engine. This might be due to the enhanced mixing rate in the case of G P 9 carried by turbulence in the

combustion chamber.

Fig: 2 Comparison of Brake Thermal Efficiency with Different configurations of Piston

EXHAUST GAS TEMPERATURE

Figure 3 shows the comparison of exhaust gas temperature with brake power. The exhaust gas

temperatures are lower for GP 9 and GP 6 compared to that of normal engine. The exhaust gas temperature for

GP 9 varies from 120C at no load to 380C at full load. For G P 6, the exhaust gas temperature varies from

131C at no load to 400C at full load whereas for G P 3 it varies from 135C at no load to 410C at full load.

Lower exhaust gas temperature for G P 9 due to higher range turbulence created in the combustion chamber by

the modified piston. It was observed that there is a decrease of 11% for G P 9 compared with normal piston

configuration

Fig: 3 Comparison of Exhaust Gas Temperatures with Different configurations of Piston

IGNITION DELAY

The variation of ignition delay with brake power for different modes of combustion was shown in Fig.

4. It was inferred that ignition delay is decreased with an increase in brake power for almost all modes of

combustion. With an increase in brake power, the amount of fuel being burnt inside the cylinder gets increased

and subsequently the temperature of in- cylinder gases gets increased. This may lead to reduced ignition delay in

all modes of combustion. However, the ignition delay for diesel fuel was lower under G P 9 than the normal

piston configuration. It was observed that the ignition delay of G P 9 with brake power varies from 12.5oC A at

no load to about 8.8o

CA at full load. The reduction in the ignition delay of G P 9 is about 2.2% at max powercompared to normal engine.

0

5

10

15

20

25

30

35

0 1 2 3 4

Brakethermalefficien

cy

Power (kW)

N P

G P 3

G P 6

G P 9

100

200

300

400

500

600

0 1 2 3 4

Exhaustgastemperature(oC)

Power (kW)

N P

G P 3

G P 6

G P 9

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Fig: 4 Comparison of Ignition Delay with Different configurations of Piston

PEAK PRESSURE

The variations of peak cylinder gas pressure with brake power for different modes of combustion were

shown in Fig. 5. It was observed that the peak pressure gets decresed with an increase in brake power. During

measurements, the maximum peak pressure under NP configuration and lowest peak pressure under GP 9

configuration were observed. Because of angular momentum created by the groove in piston in GP 9

configuration, the entire fuel-air mixture got ignited and also burnt simultaneously (volumetric combustion).

This might be leading to lowestt peak pressure in GP 9 configuration compared to the other configurations.

There was a decrease of 6% in pressure when compared with normal engine configuration

Fig: 5 Comparison of Peak Pressure with Different configurations of Piston

SMOKE DENSITY

Smoke is nothing but solid soot particles suspended in exhaust gas. Fig.6 shows the comparison ofsmoke level with brake power. It can be observed that smoke increases with increase in brake power. The higher

prevailing operating temperatures due to higher turbulence in the combustion chamber result better combustion

and oxidation of the soot particles which further reduce the smoke emissions. Due to the complete combustion

of diesel with excess air the smoke emissions are marginal. At the rated load, the smoke emissions of GP 9 are

reduced by about 17% compared to normal engine..

0

5

10

15

20

0 1 2 3 4

Ignitiondelay(oC

A)

Power (kW)

N P

G P 3

G P 6

G P 9

40

50

60

70

0 1 2 3 4

Peak

Pressure

(Bar)

Power (kW)

N P

G P 3

G P 6

G P 9

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Fig: 6 Comparison of Smoke Densities with Different configurations of Piston

CONCLUSIONS:

The Configuration GP 9 enhances the turbulence and hence results in better air-fuel mixingprocess among all three configurations of diesel engine. As a result, the thermal efficiency is increased and SFC

and soot emission are reduced, the Nox emission is decreased owing to better mixing and a faster combustion

process. It can be concluded that GP 9 is the best trade-off between performance and emissions.

Based on this investigation, the following conclusions are drawn:

More power output is derived from the same given charge operating on the same compressionratio.

Lesser smoke due to far more complete combustion is provided. Lesser carbon deposits in the combustion chamber, piston crown and exhaust system occur due

to controlled complete combustion.

Exhaust gas temperatures is lower due to quicker and even flame propagation. There is better fuel economy due to improved and complete combustion. Decrease in the cylinder pressure due to the effective combustion

REFERENCES

[1] Xueliang H; ShuSong L ,(1990) ; Combustion in an internal combustion engine.Mechanical Industry Press, Peking[2] Shaoxi S; Wanhua S (1990) Some new advances in engine combustion research. Trans CSICE 8: 95-104[3] Wu Zhijun, Huang Zhen In-cylinder swirl formation process in a four-valve diesel engine Experiments in Fluids 31 (2001) 467 - 473

Springer-Verlag 2001

[4] Arturo de Risi, Teresa Donateo, Domenico Laforgia," Optimization of the Combustion Chamber of Direct Injection Diesel Engines"SAE2003-01-1064.

[5] Herbert Schapertons, Fred Thiele,"Three Dimensional Computations for Flow Fields in D I Piston Bowls". SAE60463.[6] Corcione. F. E, Annunziata Fusca, and Gerardo Valentino, "Numerical and Experimental Analysis of Diesel Air Fuel Mixing" SAE

931948.

[7] Ogawa. H, Matsui. Y, Kimura. S, Kawashima. J, "Three Dimensional Computations of the Effects of the Swirl Ratio in Direct-Injection Diesel Engines on [NO.sub.X] and Soot emissions" SAE: 961125.[8] Gunabalan, A.; Ramaprabhu, Nov 2009; Effect of piston bowl geometry on flow, combustion and emission in DI diesel engine-aCFD approach. , International Journal of Applied Engineering Research

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SMOKENUMBER(Bosch)

Power (kW)

N P

G P 3

G P 6

G P 9