USE IN CI AND SI ENGINES

Technical Field of the Present Invention

The present invention relates to a combustion chamber for an internal combustion engine, more particularly the present invention relates to a combustion chamber of an internal combustion engine for improving engine economy, performance and combustion efficiency and reducing hazardous emissions by optimizing the shape of the combustion chamber for achieving turbulent flow inside the combustion chamber. The invention is suitable for use in compression ignition systems with diesel as fuel and spark ignition systems with gasoline or natural gas as fuel without need for any structural modifications. The invention is suitable for use in automobiles, tractors, ships, locomotives and other machinery.

Background of the Present Invention

Conventionally, single- and multi-cylinder direct injection diesel engines are of open-chamber type. Open-chamber type engines rely on spray characteristics and air motion for mixing of fuel and air. To achieve homogeneous mixing in the combustion chamber before ignition, common rail system, consisting of a fuel injection system where fuel is supplied at high pressure (>1400 bar) by solenoid valves, is widely used. Use of injectors consisting of multiple (8-10) nozzles having small diameters in the common rail system allows for fuel to be sprayed into the combustion chamber at a pressure of 1400-2500 bar, improving the performance and economy of the engine as well as reducing soot emission.

However, use of high pressures in the common rail system makes this technology costly to manufacture and repair. On the other hand, high pressure spraying of fuel into the combustion chamber can lead to sudden ignition and combustion of fuel which causes excessive increase of pressure and temperature. Excessive pressures can damage components of the engine and excessive temperatures can increase engine noise and formation of environmentally hazardous NO _x emissions. To meet the NO _x emission and engine noise control standards, technologies such as multiple injection strategies (consisting of a pilot injection, a main injection and post injections), exhaust gas recirculation (EGR), particulate filters and selective catalytic reduction (SCR) are developed and used in engines. These technological advances have partially ensured that the aforementioned emission standards are met; however, they have also increased fuel consumption and made engine manufacture overly complex and more costly.

For this reason, conversion of diesel engines used in public transportation vehicles, trains, ships and generators to natural gas (LNG or CNG) is being evaluated, as natural gas is cheaper and more environmentally friendly. This is considered to be the best way to ensure compliance with emission standards for diesel engines without using additional overly complex and costly systems in light of recent technological advances.

In the known art, diesel engines are converted from compression ignition (CI) systems to spark ignition (SI) systems. However, as CI systems have a high compression ratio (ε> 14: 1), when the engine is converted to an SI system for use of 100% natural gas as fuel, it is necessary to reduce the compression ratio down to 10.5: 1 to 11 : 1 to prevent detonation (engine knocking). To achieve this, the pistons used in diesel engines need to be replaced by new ones, which makes the conversion of diesel engines difficult and costly. On the other hand, reduction of the compression ratio causes the engines to have lower performance and efficiency. To keep engine performance at a desirable level and to ensure spark ignition works efficiently, air-fuel equivalence ratio, λ, needs to be kept between 0.90-1.05, which, while reducing CO, hydrocarbon and particulate matter (PM) emissions, actually increases NO _x emissions. In this case, use of emission reduction systems such as EGR and SCR is required to comply with the standards. In addition, use of 100% natural gas as fuel requires the use of richer air fuel mixtures (λ=0.90- 1.05) compared to diesel (λ> 1.5), which causes 8%-10% increase in CO2 emissions.

The attempts made in the state of the art to alleviate these problems associated with combustion chambers are exemplified by EP 0 412 552 Bl, EP 2 799 686 Al, JP 2013-053529 A, US 2015/354439 Al, US 4,779,587 A, US 6,640,772 B2, US 8,156,927 B2, US 9,284,877 B2, WO 2011/055055 Al and WO 2011/133664 A2.

The present invention aims to improve on the problems described in the prior art. The invention makes use of a combustion chamber promoting two-axis turbulent flow to provide an internal combustion engine with good efficiency and economy and reduced emissions, suitable for use with a variety of fuels including diesel (CI), gasoline and natural gas (SI). Two-axis turbulent flow comprises turbulent flow about horizontal axis Y caused by gas inflow (air for CI engine, fuel-air mixture for SI engine) along axis X through gas inlet towards the central projection of the combustion chamber and turbulent flow about vertical axis X caused by tangential flow of air (or air-fuel mixture where suitable) into the combustion chamber through sickle-shaped inlet channels.

The present invention provides a combustion chamber for an internal combustion engine as provided by the characterizing features defined in Claim 1. Objects of the Present Invention

The object of the invention is to provide a combustion chamber for an internal combustion engine that can be used in both CI system with diesel as fuel and SI systems with gasoline and natural gas as fuel.

A further object of the invention is to provide a combustion chamber for an internal combustion engine that has an optimized shape to promote two-axis turbulent flow for improving performance and combustion efficiency and reducing hazardous emissions.

Brief Description of the Technical Drawings

Accompanying drawings are given solely for the purpose of exemplifying a combustion chamber, whose advantages over prior art were outlined above and will be explained in brief hereinafter.

The drawings are not meant to delimit the scope of protection as identified in the Claims, nor should they be referred to alone in an effort to interpret the scope identified in said Claims without recourse to the technical disclosure in the description of the present invention.

Figure 1 demonstrates a vertical sectional view of an embodiment of the present invention.

Figure 2 demonstrates a top view of an embodiment of the present invention as depicted in Figure 1. Figure 3 demonstrates a vertical sectional view of an alternative embodiment of the present invention.

Figure 4 demonstrates a top view of an alternative embodiment of the present invention as depicted in Figure 3.

Figure 5 demonstrates a perspective view of the invention as depicted in Figure 3.

Figure 6A demonstrates indicator diagram for a known combustion chamber.

Figure 6B demonstrates indicator diagram for a combustion chamber according to the present invention in CI engine.

Figure 6C demonstrates indicator diagram for a combustion chamber according to the present invention in SI engine. Detailed Description of the Present Invention

The following numerals are referred to in the detailed description present invention:

10 Combustion chamber

11 Piston

12 Piston head

13 Intake port

14 Exhaust port

15 Sickle-shaped channel

16 Fuel injector

17 Central projection

18 Lateral face

19 Fuel spray

20 Gas inlet

21 Turbulent flow about axis Y

22 Turbulent flow about axis X Figures 1 and 2 illustrate an embodiment of the present invention, referred to as combustion chamber (10). Combustion chamber (10) is formed by a bowllike depression in piston (11). In this embodiment, combustion chamber (10) and piston (11) are concentric, so that the vertical symmetry axis of combustion chamber (10) (axis X) and the vertical symmetry axis of piston (11) (axis Z) overlap. Combustion chamber (10) is of spherical shape and comprises a central projection (17) corresponding to axis X. Fuel injector (16) is placed along axis X. Tube diameter of combustion chamber (10) is denoted d and preferably d is also equal to the distance between the centers of the two circles of the cross- section of combustion chamber (10) along axis Y (indicating that combustion chamber (10) has a horn torus structure) and the diameter of gas inlet (20). Diameter of piston (11) is denoted D and preferably d≤0.4D. Distance between top of piston (11) and axis Y is denoted h, and preferably h= d/2 + 3 to 10 mm.

Gas inflow is achieved through intake port (13) during the intake stroke. When gas inflow reaches piston head (12), it is partially guided by sickle- shaped channels (15) to provide tangential entry of gas to combustion chamber (10) through gas inlet (20). Preferably at least two sickle-shaped channels (15) are present. More preferably at least three sickle-shaped channels (15) are present. Gas tangentially entering combustion chamber (10) grazes lateral face (18) and swirls around combustion chamber (10) creating turbulent flow about axis X (22). Gas inflow is also achieved along axis X through gas inlet (20). Gas inflow is guided by central projection (17) and swirls around combustion chamber (10) creating turbulent flow about axis Y (21). Therefore, two-axis turbulent flow is established in combustion chamber (10) before combustion occurs.

Fuel spray (19) is injected (<350 bar) from fuel injector (16) along axis X onto central projection (17). Due to angle of injection and the circular shape of lateral face (18) of combustion chamber (10), fuel spray (19) is guided by central projection (17) and swirls around combustion chamber (10). A higher amount of the fuel is smeared on lateral face (18) as a thin film and thereby more easily evaporates at lower temperatures (<400°C) due to turbulent flow about axis X (22) and turbulent flow about axis Y (21). Evaporation takes place at lower temperatures so pyrolysis is prevented and hydrocarbon structure of fuel is conserved prior to ignition. As a result, harmful emissions resulting from incomplete combustion of fuel, such as carbon and particulate matter are reduced.

As evaporation occurs, fuel is transferred by turbulent flow about axis Y (21) from the bottom of combustion chamber (10) towards the center where it is hottest. Combustion initially takes place in the combustion chamber (10) proper, so that rapid combustion phase takes place away from the comparatively cooler walls of the cylinder. In this way, flame is not cooled and so combustion efficiency is increased and emission of CO and hydrocarbons is decreased. In addition, as combustion follows fuel vaporization, the rate of pressure increase within combustion chamber (10) is decreased, leading to prevention of overheating of combustion products due to excessive pressure increase, which cause NO _x and noise emissions, and an optimum combustion process is obtained.

In the embodiments, a compression ignition diesel engine is described. However, the invention is also suitable for use in SI engines by replacing fuel injector (16) with a spark plug. The invention is also suitable for use with fuels other than diesel, such as gasoline and natural gas with no further modifications. Gas inflow may refer to air inflow and air-fuel mixture inflow, for example from a carburetor, where suitable.

When combustion chamber (10) is used in an SI engine, air-fuel mixture is taken into combustion chamber (10) by gas intake (20) via intake port (13) and sickle-shaped channels (15). As described above, combination of tangential and vertical intake of air-fuel mixture and circular shape of lateral face (18) allow formation of two-axis turbulent flow (21 and 22). Fuel is concentrated in the central region (close to axis X) of combustion chamber (10) due to an apparent centrifugal force about axis X (22) caused by turbulent flow acting on air-fuel mixture, since natural gas and volatile components of gasoline are lighter than air.

This creates a lean mixture (λ>1) in the outer region (away from axis X) and a rich mixture (λ<1) in the center of combustion chamber (10). As a result, a stratified concentration gradient is obtained rather than a homogeneous air- fuel mixture. Therefore, combustion process occurs with a mixture which has a lean general composition (λ>1.35), which increases combustion efficiency and reduces exhaust gas emissions, in particular NO _x and CO2 emissions. On the other hand, two-axis turbulent flow (21 and 22) system of combustion chamber (10) allows high compression ratios (ε>14) generally used with diesel to be used with any fuel regardless of its octane rating. Also, detonation (engine knocking) in SI systems is prevented. As a result, this engine has better economy compared to diesel engines and harmful emissions, such as CO, hydrocarbons, NOx and CO2 are reduced to comply with emission standards without requiring use of additional emission reduction systems.

Figures 3, 4 and 5 illustrate an alternative embodiment of the present invention. In this alternative embodiment, combustion chamber (10) and piston (11) are eccentric, so that axis X and axis Z do not overlap. Fuel injector (16) is placed along axis X. Gas inflow enters combustion chamber (10) tangentially via sickle-shaped channels (15). Preferably at least one sickle-shaped channel (15) is present. Examples The performance and emissions of combustion chamber (10) was tested with a single-cylinder engine (stroke ratio 1.117) and indicator diagrams (pressure versus crank angle) were obtained. Said indicator diagrams are illustrated in Figure 6 for the following systems: A, known combustion chamber in CI engine (Figure 6A); B, combustion chamber (10) according to present invention in CI engine (Figure 6B); and C, combustion chamber (10) according to present invention in SI engine (Figure 6C), and the results are given in Table 1. The following operation conditions were held constant: 3000 rpm, ε=17.5, λ=1.42 and volumetric efficiency, η _ν =0.9.

Table 1

In system A, using a known combustion chamber in CI engine, injection commenced at crank angle, θ=-15° (15° before TDC) and injection pressure, approximately ρ,=500 bar and ignition occurred at θ=-5° (Figure 6A). At time of ignition, pressure inside combustion chamber, p _c =4.5 MPa. Maximum pressure, p _Z max=7.7 MPa was reached in 10° at θ=5° (5° after TDC). Therefore, rate of increase of pressure (Δρ/ΔΘ) in combustion chamber was determined to be 0.32 MPa/l°. In addition, it was found that, mean effective pressure, p _e =0.653 MPa; effective efficiency, η _ε =0.306; smoke color, k=2.52 1/m; and NO emission was 1100 ppm (Table 1). In system B, using combustion chamber (10) according to present invention in CI engine, injection commenced at crank angle, θ=-7.3°, as at least 95% of the fuel is smeared on lateral face (18) as a thin film which reduces the length of the ignition delay period, and injection pressure, p,<350 bar and ignition occurred at θ=2° (Figure 6A). At time of ignition, p _c =4.5 MPa. Maximum pressure, p _Z max=5.7 MPa was reached in 9° at θ=11°. Therefore, rate of increase of pressure (Δρ/ΔΘ) in combustion chamber was determined to be 0.12 MPa/l°, showing 62.5% decrease compared to a known combustion chamber. This has prevented overheating of combustion products due to excessive pressure increase, which has reduced NO emission to 302, 72.5% decrease compared to a known combustion chamber (Table 1).

In addition, while p _zm ax showed a 26% decrease compared to a known combustion chamber, this has not caused an adverse effect on engine performance. p _e , He and k values remained approximately constant (Table 1). This is due to fact that maximum pressure was reached at θ=11°, which corresponds to the most opportune moment for piston-connecting rod mechanism to have maximum momentum. Also, as piston-connecting rod mechanism is exposed to lower pressure, mechanical losses and noise emissions were reduced.

For system C, using combustion chamber (10) according to present invention in SI engine with ε=17.5, the results are given in Figure 6C and Table 1. %100 CNG with octane rating of 120 was used as fuel. No detonation (engine knocking) was observed during operations or from the pressure versus crank angle diagram, which is a novel feat at such a high compression ratio. Pzmax=6.5 MPa was reached at θ=11°, indicating that flame propagation was slower compared to diesel engine in spite of the high compression ratio. Despite p _zm ax showing a 16% decrease, no adverse effect on engine performance was observed. On the contrary, as maximum pressure was reached at θ=11°, a time most suitable for the connecting rod to exert tangential force on the crankshaft, engine performance was improved. Compared to a known combustion chamber, p _e of engine showed 9.3% increase and η _ε showed 8.4% increase. Additionally k was reduced to zero and NO emission showed 63.4% decrease (Table 1).

Combustion chamber (10) according to present invention can be used in both CI and SI systems with a variety of fuels such as diesel, gasoline and natural gas. Use of combustion engine (10) increases engine efficiency by 5%-10% while reducing hazardous emissions (CO, hydrocarbons, CO2, NO _x , etc.) in compliance with current EU emission standards (Euro 6, Stage III-IV, Tier 3- 4) without need for additional emission control systems. The invention therefore proposes a combustion chamber (10) for a compression ignition or spark ignition internal combustion engine characterized by said combustion chamber (10) having a central projection (17) extending from the center of the bottom of a circular concave depression formed in a piston head (12) along an axis X; said combustion chamber (10) consisting of two circular sections defined by lateral face (18) divided by central projection (17); said circular sections defined by lateral face (18) of combustion chamber (10) having a diameter d; said circular sections defined by lateral face (18) of combustion chamber (10) having a distance between their centers along axis Y of d; said piston head (12) comprising at least one sickle-shaped channel (15) such that tangential intake of gas into combustion chamber (10) is facilitated and a cylindrical gas inlet (20) having a diameter d such that intake of gas into combustion chamber (10) along axis X is facilitated; and said lateral face (18) providing circulatory gas flow around axis X and axis Y so as to effectuate two-axis turbulent flow (21 and 22) mixing of air and fuel.

In a further variation of the invention, said combustion chamber having a distance between piston head (12) and axis Y h equal to d+3 to 10 mm.

In a further variation of the invention, said piston (11) having a diameter D where d≤0.4D. In a further variation of the invention, said combustion chamber (10) and said piston (11) are concentric

In a further variation of the invention, said combustion chamber (10) and said piston (11) are eccentric.

In a further variation of the invention, said combustion chamber (10) comprises a fuel injector (16) along axis X of central projection (17).

In a further variation of the invention, said fuel injector (16) comprises one nozzle such that direction of fuel spray (19) along axis X towards central projection (17) and smearing of fuel along lateral face (18) in a thin film are facilitated. In a further variation of the invention, said combustion chamber (10) comprises a spark plug along axis X of central projection (17).

In a further variation of the invention, diesel fuel is usable.

In a further variation of the invention, gas intake comprises air.

In a further variation of the invention, gasoline or natural gas fuel is usable. In a further variation of the invention, gas intake comprises air and fuel.

In a further variation of the invention, compression ratios above 13:1 are usable for all fuel types.