CN106536889A - Combustion method - Google Patents

Abstract

The present invention relates to a self-igniting internal combustion engine comprising at least one cylinder, a piston reciprocating within the cylinder, and an injector, the piston defining with the cylinder a combustion chamber and comprising a combustion well facing the combustion chamber Bowl, the injector is located above the combustion bowl for injecting fuel into the combustion bowl.

Description

burning method

The present invention relates to a compression-ignition internal combustion engine having a piston with a piston recess and, in combination with selected injection means in particular at the nozzle injection location, to a method for forming a fuel/air mixture with improved homogeneity method.

DE 197 07 873 (A1) discloses an injection device for an internal combustion engine, in particular a diesel engine, which has in each case a combustion chamber formed by a piston, a cylinder, and a cylinder head, as well as injection nozzles and package The cladding extends from the cylinder head into the combustion chamber and surrounds the nozzle in such a way that at top dead center (TDC) a pre-chamber is disposed between the cladding and the piston, where the pre-combustion Pre-combustion occurs in the chamber.

DE 27 29 050 A1 already discloses a cylinder head for a reciprocating piston internal combustion engine, in particular a diesel engine, having inlet and outlet ducts provided thereon for a gaseous medium to be introduced into the interior of the internal combustion engine In the combustion chamber and to be discharged therefrom again, these inlet and outlet ducts are equipped with valves opening into openings in the corresponding combustion chamber and closing said openings, and the stems of said valves are in the region of said ducts guided axially in the guide, and each combustion chamber is assigned at least one injection nozzle for the fuel that is also to be introduced into the combustion chamber arranged on the cylinder head, and the cylinder head itself serves to close the opening towards it and forming the combustion chamber, characterized in that the cylinder head has neither cooling fins nor water core pipes and is equipped with at least one line for coolant and that the line is guided on the cylinder head at At least one component that is highly loaded by the heat of the combustion gases.

EP 1 045 136 A1 discloses a method for operating a reciprocating piston internal combustion engine in which fuel is injected directly into the operating space through injection nozzles open towards the operating space formed in the cylinder between the cylinder head and the piston between and including the piston recess into which the fuel is injected centrally at a flat angle to the piston crown just before top dead center in the lower part load range for the formation of a non-homogeneous mixture, in the In the subsequent partial load range, the fuel is at least partially injected into one area at a sharper angle relative to the piston crown so as to create a homogeneous mixture, and in the full load range, a portion of the fuel is injected at from 180° to 20° The range of crankshaft angles before top dead center is injected at a sharper angle relative to the piston crown to first serve to create a homogeneous mixture, and the rest of the fuel is injected relative to the piston crown in the range around top dead center A flatter angle injects into the plunger recess for a formed, non-homogeneous mixture.

EP 2 003 303 B1 discloses a piston having a piston crown and a piston recess provided in the piston crown, in particular for an internal combustion engine operating in two possible different operating modes, the piston being located in the Translatable during operation along the longitudinal axis of the piston, and wherein the piston recess has a concave region for configuring an omega-shaped base shape, the recessed surface of the omega-shaped piston recess facing the longitudinal axis of the piston has a substantially cylindrical A partial region, which is oriented along the piston longitudinal axis and merges into the piston crown to form a convex recessed edge. Furthermore, the invention relates to a method of forming a fuel/air mixture using a piston of this type. A piston of the type mentioned above is provided which enables improved control of controlled compression ignition and enables the use of the LTC method for higher speeds and higher loads. This is achieved by a piston of the above-mentioned type having a height T2 in the substantially cylindrical partial area, where T2 is equal to 0.025d, and the convex concave edge has a radius of curvature R2, where R2 is equal to 0.006d, d describes is the piston diameter.

The disadvantages of the conventional diesel engine approach are firstly the soot emissions due to inhomogeneity of the fuel/air mixture of the direct injection diesel engine. Due to the inhomogeneity of the fuel/air mixture, mixture formation takes place in locally narrow, restricted ranges to such an extent that compression ignition conditions locally prevail, while the mixture is still very inhomogeneous in other ranges. Accordingly, the compression ignition is started earlier than would be advantageous for further homogenization of the mixture. In this case, soot is formed in particular in the region of mixtures with substoichiometric local air ratios (lambda<1).

A further disadvantage in the conventional direct injection diesel engine approach is the high nitrogen oxide emissions caused by the high temperatures resulting from the process.

One possibility for reducing nitrogen oxide emissions is the use of exhaust aftertreatment systems, such as SCR catalytic converters or exhaust gas recirculation (EGR), by means of which the combustion temperature, in particular the peak process temperature, can be reduced. The formation of nitrogen oxides also decreases as the combustion temperature decreases.

Disadvantages of this possibility are cost and packaging, or in the case of internal EGR, poor transient behavior of the engine in unstable state operation.

Against this background, the object of the present invention is to provide a method for forming a fuel/air mixture which does not have the disadvantages described in conventional methods. .

Said object is achieved by a piston having a piston recess and by a corresponding method for operating an internal combustion engine having at least one cylinder and a piston movable to and fro in the cylinder, the piston being connected to the cylinder A combustion chamber is defined and has a piston recess facing the combustion chamber, with an injector arranged above the piston recess for injecting fuel into the piston recess. Further advantageous developments are indicated in the subclaims. This engine has pistons with omega-shaped dimples. The depressions are flat and of broad design with a pronounced central cone. The piston recess is optimized for very low pollutant emissions in part-load operation and for optimum air utilization at full load.

The diesel engine operating method is based on compression ignition of the fuel introduced into the combustion chamber. The fuel injection here takes place directly into the main combustion chamber as direct injection or into the pre-chamber or swirl chamber as indirect injection. The fuel evaporates, mixes with the compressed hot cylinder contents, and is autoignited. This internal mixture formation thus has a non-homogeneous fuel/air mixture. The main characteristic characteristic of diesel engines is mass regulation, that is to say, load setting is carried out at a predefined air mass by changing the air ratio by the injected fuel mass.

Due to the fuel injection in the region of top dead center and the short time period of only a few milliseconds available for introducing fuel, the injection system and injection parameters are given decisive significance.

The operation of mixture formation and the partly simultaneous conventional diesel combustion process after fuel injection can be broken down into these partial processes indicated in the following text: atomization, evaporation+mixing, pre-reaction, ignition, combustion+pollutant formation.

Ignition delay is defined as the time between start of injection and ignition of compression. Here, jet propagation, atomization, and evaporation are referred to as physical ignition delay, while the chemical reaction that starts at the onset of evaporation is referred to as chemical ignition delay.

Due to the high injection pressure, the injection jet leaves the nozzle at high velocity and breaks up into a large number of droplets of different sizes due to aerodynamic forces generated by the relative velocity of the fuel to the cylinder contents. The quality of fuel atomization depends on the velocity, density and surface tension of the fuel, on the level of injection pressure, on the nozzle geometry, and on the density and movement conditions of the combustion chamber air. Fuel atomization is promoted by the turbulence present in the injection jet. The interaction between the droplets results in a collision process, whereby fragmentation or combination of the droplets can occur. The most pronounced braking of fuel droplets occurs in the region of the jet tip and jet edge. Subsequent injected fuel penetrates the jet tip, which leads to its mushroom-like form and to a rather pronounced structure in the jet. The turbulent interaction of the fuel droplets during this operation favors mixture formation.

The geometry of the combustion chamber likewise contributes a substantial part to the optimization of the combustion method. Adapting the injection system and the combustion chamber geometry to one another is of particular importance for utilizing the maximum potential. The parameters of the combustion chamber to be optimized are above all the combination of the shape of the piston recess and the air swirl that is generated by the inlet duct, as well as the minimization of the closed volume in the flat gap, in the ignition boss and in the cylinder head of.

The incoming flow into the cylinder, which rotates around the cylinder axis, is called a swirl flow, the formation of which depends on the geometry of the inlet section. The predominant content movement in the top dead center region of the combustion chamber is caused by the swirling flow generated during the intake phase, which has a decisive influence on the subsequent mixture formation and combustion. The measurement of the cylinder head flow of the test vehicle is carried out on a stationary flow test rig.

A substantial influence on the engine characteristics at part load and full load is made by means of a fixed compression ratio. In addition to the actual combustion chamber, this cylinder head also has a substantial influence on combustion. Four-valve technology with two inlet and two outlet valves per cylinder has established itself in most diesel engines for passenger vehicles, with these valves usually actuated via two camshafts. Compared to the two-valve concept, advantages arise in terms of a larger opening cross section and thus improved cylinder filling and a variably configurable swirl. Here, one of the two inlet ducts is designed as a filling duct with high throughput and low swirl formation, and the second inlet duct is designed as a more pronounced swirl generating air swirling motion and reduced throughput pipeline. This design makes it possible to achieve swirl control by cutting off the filling duct at low load points (since the mass-regulated diesel engine is in any case running with a large air excess). The result is improved air-side mixture formation with a reduced effect on soot formation and thus improved compatibility of exhaust gas recirculation.

Here, it is important to adapt the injection system to the swirl formation to avoid fuel jets drifting into each other. In addition, the air system of the diesel engine has an effect on combustion. Here, the design of the exhaust-gas turbocharger exhibits dynamic response characteristics during start-up or acceleration from low engine speeds and maximum boost (possible due to the turbo reaching the surge limit at the rated output point of the engine) A compromise between pressures. The introduction of variable turbine geometry brings about substantial improvements here. In the future, a two-stage charging concept or sequential charging will offer further potential, in which case a small high-pressure turbocharger ensures a satisfactory dynamic response by increasing the boost pressure at low engine speeds and loads characteristics, while the large low-pressure turbocharger enables much higher boost pressures in the higher speed range of the engine. With late opening of the inlet valve, the piston travels down at the beginning of the intake stroke when the inlet valve closes and expands the residual gas in the cylinder, thus creating a vacuum in the combustion chamber. As the inlet valve opens, intake air then flows into the cylinder at a higher flow rate due to the pressure differential between the intake manifold and the combustion chamber. With this strategy, it is possible to increase the overall content movement, but above all the swirl flow in the combustion chamber, which promotes air-side mixture formation. Depending on the operating point and air requirements of the engine, this strategy can considerably improve the air-side mixture formation and thus have a positive influence on the emission characteristics. On the inlet side, the duct for generating the movement of the contents can be configured as a swirl duct, and another duct for optimizing the cylinder filling and realized as a tangential duct.

Owing to the high temperatures in the combustion chamber, vaporization of firstly the easily volatile constituents and then of the semivolatile constituents of the fuel droplets takes place immediately after the atomization of the fuel. Diffusion and convection lead to mixing of fuel with air, in which case a very inhomogeneous mixture is formed, the region close to the nozzle and its jet core being quite substoichiometric and having a high proportion of still liquid fuel. As the distance from the nozzle and jet axis increases and as time elapses after the start of injection, the mixture becomes more dilute and gaseous. Due to the stratification, there is always a region in the outer region of the jet with a mixed state that favors compression ignition.

In addition to the described physical process of mixture preparation, chemical reactions, which depend greatly on pressure and temperature, also result in the generation of heat and the formation of radicals at positions with the most favorable prerequisites for this purpose. Compression ignition occurs. The inhomogeneous mixture in the combustion chamber of the diesel engine always has regions with a favorable ignition state. Here, the ignition of the first combustion preferably takes place in the region of the richer mixture with an air ratio between I=0.6 and I=0.8. The total area ratio predominating in the combustion chamber has little effect on ignition delay. In the case of increasingly richer mixtures, slower reactions occur, mainly due to more pronounced mixture cooling during the endothermic evaporation of larger fuel masses. In addition to pressure, temperature, and injection volume, the cetane number of fuel, injection pressure, nozzle hole diameter, boiling range of fuel, and combustion chamber flow rate are decisive for ignition delay.

Another important parameter affecting the chemical prereaction is exhaust gas recirculation. The time profile of conventional diesel engine combustion is broken down into stages.

First, a zone of locally homogeneous fuel/air mixture is ignited with premixed combustion with fast flame propagation and with a satisfactorily prepared and reactive mixture (first stage). The rapid flame propagation of this premixed combustion produces steep pressure gradients.

The combustion phase is characterized by a distinct peak in the combustion process. In the subsequent controlled diffusion combustion of the mixture, a further conversion of the fuel quantity takes place (second stage). These regions of incomplete mixing of the reaction participants burn in a controlled manner due to the mixing process by allowing fuel and air to flow into the flame in a diffusion-controlled manner. Here, the fuel is mixed with air and combustion gases with various components. The mixture formation operations that take place during the combustion have a decisive influence on the combustion process. The gradually decreasing gas temperature and the low concentration of the reaction participants cause combustion of the residual fuel mass, which takes place in a very sluggish manner in the post-combustion phase (third phase).

In the case of diesel engine combustion, hydrocarbon emissions (HC) and carbon monoxide emissions (CO) as a product of incomplete combustion are very low compared to gasoline engines. Reducing nitrogen oxides and particulates has proven to be much more difficult in the case of diesel engines. In contrast, in the case of homogeneous or partially homogeneous combustion of diesel fuel, much higher HC and CO emissions partially occur.

The hydrocarbons in diesel engine exhaust are partly products of incomplete combustion. The cause of incomplete combustion can be very lean areas in the combustion chamber that are not reached by the flame, are at low temperature, do not react in time, or only partially react. If the combustion is shifted to the expansion phase, the flame may extinguish due to the gradually decreasing chamber temperature and an increase in HC emissions also occurs.

Furthermore, hydrocarbon emissions increase, for example due to uncontrolled late penetration of fuel due to fuel nozzle leaks, or due to relatively large amounts of fuel impinging on cylinder or recess walls.

Carbon monoxide will be primarily attributed to incomplete oxidation of unburned hydrocarbons. Since the diesel engine operates with a lean mixture overall, although it locally has a rich mixture region during the combustion process, sufficient oxygen is present for CO oxidation. The rise occurs at an air ratio close to the stoichiometric ratio.

Regarding the formation of nitrogen oxides, a distinction is made between essentially three mechanisms (thermal NO according to Zeldovich, rapid NO according to Fenimore, and combustion NO).

According to Zeldovich's thermodynamic NO; during combustion at high temperatures (greater than 2200 K), the formation of nitrogen that is present in the intake air and does not exhibit strictly inert properties occurs in localized regions with excess oxygen Nitrogen oxidation substances (NOx) reaction. The nitrogen oxides are called thermal NO and are produced according to the Zeldovich mechanism. Overall, at least 16 known reactions are involved in the formation of nitrogen oxides. All equations each have two reaction participants on each side (bimolecular two-way reaction). Eight oxidation forms are known, of which only nitrogen monoxide and nitrogen dioxide are relevant to diesel engine combustion. Nitric oxide is formed during the combustion phase. In contrast, nitrogen dioxide is a secondary product formed by post-oxidation of nitrogen monoxide at low temperatures. The main parameters affecting the formation of thermal NO are thus the temperature, the oxygen concentration at the combustion site, and thus the local air ratio, and the high temperature dwell time.

The fast type of NO according to Fenimore is formed at the flame front, specifically in the fuel-rich state. The reaction of hydrocarbon groups with nitrogen molecules forms cyanides, from which NO is formed in a secondary reaction with oxygen carriers. The formation of the fast type of NO occurs primarily in the absence of air, since acetylene (carbide gas) is formed as a precursor of CH groups only under fuel-rich conditions. Due to the low concentration of CH groups, however, this type of nitrogen oxide formation plays a minor role.

During combustion, simple amines and cyanides are formed by decomposition of nitrogen bound in the fuel. The secondary nitrogen compound further reacts with oxygen to form NO. Since the proportion of nitrogen contained in the fuel is very low and only part of it is converted into NO, the proportion of fuel NO in the nitrogen oxide emissions is likewise negligible.

These Zeldovich equations are sufficient for basic considerations. Around the limit temperature of 2200K, there is a continuous slowdown of the _NOx formation mechanisms, which eventually stall below this limit temperature.

Law-limited particulate emissions consist of soot and accumulated hydrocarbons, sulphides, ash, and metal wear. This formation occurs according to a very complex mechanism and has not been explained in detail. The combustion of hydrocarbons leads to soot formation if the fuel burns in localized air shortages and at temperatures above 1400K or if pyrolysis processes occur in the fuel. The soot formation process can be divided into the partially simultaneous stages of new particle formation, surface growth, aggregation, agglomeration, and post-oxidation.

In regions of high oxygen concentration, the oxidation of soot particles takes place in parallel with the soot formation process. Due to the described process, the soot emissions from diesel engine combustion are only approximately 5% of the maximum soot concentration occurring during this period. Soot oxidation occurs at temperatures greater than 1300K. In addition to molecular oxygen, other reaction participants such as OH groups are also of significance during the oxidation of soot. The oxidation rate increases with increasing temperature and/or oxygen partial pressure.

The geometry of the combustion chamber likewise contributes a substantial part to the optimization of the combustion method. Adapting the injection system and the combustion chamber geometry to one another is of particular importance for utilizing the maximum potential. The parameters of the combustion chamber to be optimized are above all the combination of the shape of the piston recess and the air swirl that is created by the inlet duct, as well as the minimization of the closed space in the flat gap, in the ignition boss and in the cylinder head of. A substantial influence on the engine characteristics at part load and full load is made by means of a fixed compression ratio. For modern passenger vehicle diesel engines, the current development trend exhibits a reduction in the compression ratio. Future concepts feature compression ratios of 17:1 and below. Nitrogen oxide formation is reduced by reducing end-of-compression pressure and end-of-compression temperature. The longer free injection jet length due to the larger dimple diameter improves mixture preparation and thus reduces soot formation. Air utilization in full load operation is also improved, which together with an earlier start of injection without increasing the maximum peak pressure results in an increase in the power output of the diesel engine. A reduction in the compression ratio has a negative effect on the formation of hydrocarbons and carbon monoxide, since these reactions stall due to the lower compression end pressure and lower compression end temperature, and thus the combustion does not proceed completely. Further disadvantages are the reduced efficiency of combustion (which can be partly compensated by adaptation of the injection start) and the impairment of the cold-start behavior, especially at very low external temperatures. Disadvantages in cold start behavior can be countered by improved glow plugs and heating strategies, optimized injection strategies, and adapted positions of the injection jets relative to the glow plugs.

Conventional diesel combustion proceeds as a mixture-controlled diffusion combustion with a very inhomogeneous fuel/air ratio after the premixed combustion of the introduced fuel quantity in the ignition delay. This firstly contributes to the formation of soot in the zone of local oxygen deficiency, and secondly the diffusion combustion takes place at the edge of the jet in the near-stoichiometric range, which leads to local high temperatures in the combustion zone and contributes to The thermodynamic NO is formed by the Aldovic mechanism. In contrast, a homogeneous fuel/air mixture exists in a conventionally operated gasoline engine. Combustion takes place by spark ignition through an advancing flame front with locally very high temperatures and thus high NO formation and virtually no particle formation.

In the case of diesel engine combustion, this injection system is given significance. The direct introduction of fuel into the combustion chamber and the preparation of the fuel/air mixture substantially affects the combustion process, how efficient it is, and the formation of pollutant emissions. In addition to the injection pressure, the injection nozzle represents one of the most important criteria in the case of an injection system. For the combustion process, it is above all an optimal adaptation of the hole diameter, pitch angle, flow coefficient, rounding of the nozzle hole inlet and outlet, and the number of nozzle holes to the combustion chamber geometry and the air movement in the combustion chamber .

In the case of diesel engines, exhaust gas aftertreatment by means of a three-phase catalytic converter and regulation of I=1, as in the case of gasoline engines, are not possible due to operation with excess air. For this reason, currently only oxidation catalytic converters are used for minimizing carbon monoxide and hydrocarbon emissions and particle filters are increasingly used.

It is possible to maintain current and future exhaust emission standards in terms of HC and CO by conventional diesel combustion methods with a combination of direct injection and oxidation catalytic converters.

In contrast, reducing nitrogen oxide emissions and particulate emissions represents a much greater challenge and for this reason development activities in the case of diesel engines are focused on this.

In general, a distinction is made between internal engine measures to reduce raw emissions and the use of exhaust aftertreatment systems. In the area of exhaust aftertreatment, diesel particulate filters are being further developed in order to minimize particulate emissions. However, using the concept in the current state of the art is associated with considerable additional costs. These include increased fuel consumption during regeneration operation of the particulate filter and _NOx storage catalytic converter, additional additives in the case of certain particulate filter technologies, reducing agents to be carried in the vehicle (e.g. In this case urea), and the production costs of these components mainly for the precious metals used. Furthermore, these regenerative operations cannot be readily provided at low engine speeds and loads by necessarily raising the exhaust gas temperature, with the result that further challenges exist here for vehicles operating exclusively in urban traffic. For these reasons, the development of the combustion process in diesel engines must continue to focus on the reduction of raw emissions inside the engine in order to make it possible to avoid exhaust aftertreatment concepts as much as possible or to design them more simply. An advantageous development of the invention provides that the injection nozzle is inclined from approximately 10° to 25° relative to the piston longitudinal axis. The cylinder bore is in the range from 90mm to 115mm.

The recessed diameter of the piston (designated D1) is in the range from 70 to 95 mm. The average diameter D2 of the piston well land in the uppermost region of the piston well bottom comprises a range from 0 to 10 mm. The piston depression radius R1 in the deepest region of the piston depression bottom is in the range from 5 to 15 mm. The piston recess radius R2 is in the range from 0 to 2 mm. The piston depression angle alpha 1 is in the range from 0° to 50°. The piston depression angle alpha 2 is in the range from 131° to 160°. The injection nozzle discharge angle gamma is in the range from 110° to 130°. The eddies according to the _Tippelmann method are in the range from DTI =0.3 to _DTI =0.4.

The depth of the depression, denoted T1, is in the range from 5 to 20 mm. The distance between the pistons 1 , denoted by T2 , is in the range from 2 to 10 mm.

In the following, the invention will be explained in more detail using an exemplary embodiment in the attached drawing, in which:

Figure 1 shows a diagrammatic representation of a piston and injection nozzle with an omega-shaped piston recess.

FIG. 1 shows a piston 1 with an ω-shaped piston recess 2 . The injection nozzle 3 is inclined approximately 20° relative to the longitudinal axis 4 of the piston or approximately 70° relative to the piston crown 5 . The cylinder bore is 98mm in one exemplary embodiment and 108mm in another advantageous refinement.

The diameter of the depression of the piston 1 (designated D1 ) is 81.2 mm in the case of the one exemplary embodiment and 87 mm in the case of the other modification. The mean diameter D2 of the piston well land in the uppermost region of the piston well bottom 6 comprises 4.7 mm in the case of the one exemplary embodiment and 5.7 mm in the case of the other refinement. The piston depression radius R1 in the deepest region of the piston depression bottom 6 is 8 mm in both exemplary embodiments. The piston recess radius R2 is 0.5 mm in both exemplary embodiments. The piston depression angle alpha 1 is 30° in both exemplary embodiments. The piston depression angle alpha 2 is 148° in both exemplary embodiments. The injection nozzle discharge angle gamma was 123° in both cases. The vortex according to the _Tippelmann method is DTI = 0.35 in both cases.

The depression depth 1 denoted T1 is 11.9 mm in the case of the exemplary embodiment and 12.75 mm in the case of the further development. The spacing of the pistons 1 designated T2 is 4.7 mm in the case of the exemplary embodiment and 5 mm in the case of the other development.

Claims (10)

1. a kind of compression ignition type explosive motor, the explosive motor have at least one cylinder and can come in the cylinder The dynamic piston of return, the piston define combustor with the cylinder and are recessed with the piston towards the combustor；The internal combustion is sent out Motivation has infusion appliance, and the infusion appliance is arranged at the piston depression top for injecting fuel in the piston depression.

2. compression ignition type explosive motor as claimed in claim 1, it is characterised in that the overall diameter D of the piston and cylinder Hole is in the range of from substantially 90 to 115mm.

3. compression ignition type explosive motor as claimed in claim 1 or 2, it is characterised in that the concave shaped portion of piston depression Divide the radius of curvature R 1 being at least partially substantially from 5 to 15mm.

4. the compression ignition type explosive motor as described in one or more in above claim, it is characterised in that from this Size T2 of the peak that piston crown is recessed to the piston is in the range of from substantially 2 to 10mm.

5. the compression ignition type explosive motor as described in one or more in above claim, it is characterised in that the work Peak equating and with the diameter D2 in the range of from substantially 0 to 10mm in the way of platform-like of plug depression.

6. the compression ignition type explosive motor as described in one or more in above claim, it is characterised in that from this Piston crown to the deepest point of piston depression size T1 is in the range of from substantially 5 to 20mm and is generally in R1's In the range of.

7. the compression ignition type explosive motor as described in one or more in above claim, it is characterised in that the work The sunk surface of plug depression has truncated-pyramidal portion subregion, the butt between the region of the deepest point and the piston crown Conical section region is oriented along the longitudinal axis of the piston, and by the depression of the frustoconical portion area configuration Surface forms angle alpha 1 with the longitudinal axis of the piston, and the angle alpha 1 includes the scope substantially from 10 ° to 50 °.

8. the compression ignition type explosive motor as described in one or more in above claim, it is characterised in that the work Plug depression formed from the deepest point of piston depression to peak, i.e. the branch of the platform-like equating part, the branch relative to The longitudinal axis of the piston are substantially 74 °.

9. the compression ignition type explosive motor as described in one or more in above claim, it is characterised in that pacified The infusion appliance for coming piston depression top is arranged to the longitudinal axis relative to the piston substantially into 20 ° of angle.

10. a kind of method for operating explosive motor, it is characterised in that using in such as above claim or many Explosive motor described in.