DK176118B1 - Method of operating a diesel-fueled dual-fuel internal combustion engine and such engine - Google Patents

Description

DK 176118 B1 i

BACKGROUND OF THE INVENTION 1. Field of the Invention This invention relates to a method of operating a diesel-fueled dual-fuel internal combustion engine which is supplied to the cylinders in air and gas in a mixing ratio, after which the air and gas are compressed during the compression stroke and ignited by foreign ignition.

US-A-4,527,516 discloses such a method used in connection with a multi-cylinder diesel-fueled dual-fuel internal combustion engine, each with a series of purge air ports located in a lower section of the cylinder and at least one exhaust valve in the cylinder cover as well as with a valve controlled gas supply with a single gas port designed as a pipe extending from a distributor line to one of the purge air ports in the cylinder port line. The engine of the US patent is a high-speed, two-stroke car engine with an engine speed of about 1800 rpm, and it is stated as a disadvantage of this fast-running engine that the gas can be difficult to burn because the flame front propagates only one-third as fast as by combustion of diesel fuel.

US-A 5,035,206 discloses a similar two-stroke diesel engine with gas supply and pilot oil ignition. This engine is also supplied with gas in the area around the purge air ports, the gas being supplied to the cylinder through a single tube which opens into the cylinder wall for a relatively short distance above the range of purge air ports. It is described that the gas injection continues after the piston has shut off the flushing air ports. The amount of gas is controlled by changing the start time 30 but not the end time of the injection.

From US-A-4 924 822 it is known to supply a diesel engine with gas from an LNG tank, where a portion of the evaporated gas is compressed to relatively low pressure and pre-mixed with the engine's intake air, while the remaining portion of the evaporated gas is high-pressure compressed air. - DK 17611® B1 2 is added to approximately 250 bar and injected through high pressure nozzles. Although compression work is saved by injecting some of the gas into the engine's intake air, the patent states that only a small part of the evaporated gas can be supplied to the intake air, otherwise abnormal combustion can occur as a result of a temperature rise during the adiabatic compression stroke.

WO 96/08641 describes an Otto-type engine which is supplied with premixed air / fuel and ignited with spark plugs. To increase engine efficiency and reduce ΝΟχ discharge, water is injected into the cylinder in the compression stroke. It is preferred that a particularly oxygen-rich stream of suction air is also supplied.

15 The above-mentioned diesel engines, which are supplied with gas at low pressure in the cylinders before the compression is started, are all subject to the problem that the gas / air mixture must not autoignite or detonate during the compression stroke. If self-ignition occurs, the pressure in the cylinder can rise substantially above the maximum combustion pressure in normal operation, and in frequent self-ignitions, the cylinder elements can be overloaded. Self-ignition usually also results in poorer utilization of the fuel value in the fuel.

Therefore, it is crucial to the diesel engine's potential for using low pressure supplied gas as fuel that there is control over how close the gas / air mixture introduced into the cylinder lies to the self-ignition limit. For a given fuel, the self-ignition limit 30 becomes lower, ie. self-ignition occurs more easily when the engine rpm is smaller, the cylinder bore is larger, when the compression ratio is greater, when the charge air pressure is greater, and when the excess air number is closer to 1. Furthermore, the self-ignition limit depends significantly on the actual fuel, which ____ DK 176118 B1 3 for gases is expressed by the metal number, with the lower metal number indicating a fuel which is easier to ignite.

These conditions have been decisive for the above-mentioned diesel engines being either laid out as fast-running diesel engines with relatively small cylinder bore or as an engine where a substantial part of the gas is high-pressure compressed and injected after compression to keep the cylinder contents as lean during compression. that there is no possibility of 10 self ignition.

The invention has for its object to provide a method for operating a dual fuel diesel engine with a better option than in the known diesel engines for using low pressure supplied gas to meet the fuel demand.

To this end, the method according to the invention is characterized in that the air to gas mixing ratio gives an air excess number of at least λ = 1.3, that with a number of injection nozzles per cylinder at least in a part of the crank angle 20 of the compression stroke 20 to 150 ° after the bottom dead point, water is injected into downward clouds directly into the mixture of air and gas in the cylinder space between the upward piston and the cylinder cover, and the water is supplied to the injection nozzles at a pressure of at least 50-25 bar.

The air excess number of at least λ = 1.3 means that the process relates to engines powered by lean-burn principle, where for proper combustion it is essential that the mixture between gas and air is very uniform. If the mixture is dissimilar, in an area of the combustion chamber there can be such a large excess air that the fuel cannot be combusted, resulting in the utilization of the fuel's combustion value and the emission of unburned gas, which causes the risk of explosions in the exhaust system 4 DK 17611118 B1 and involves an adverse environmental impact in the form of emissions of the unburned hydrocarbons. In another area of the combustion chamber, the mixture may be too greasy to cause self-ignition. The problem of obtaining a sufficiently uniform mixture of gas and air is particularly pronounced when the gas is supplied to the suction air at a relatively short distance from the cylinder or even first inside the cylinder.

The lean-mix principle is not in itself sufficient to achieve the desired enhancement of the diesel engine's ability to use low-pressure supplied gas, and, as mentioned, the principle is difficult to realize in a diesel engine due to the actual variations in the mix.

15 The direct injection of water into the gas / air mixture during the compression stroke has a number of effects that separately move the gas / air mixture away from the self-ignition limit and at the same time mitigate or solve the problem created by the lean-mixing principle. It is a primary effect that the water injection affects the gas / air mixture in a mechanical respect, the injected water clouds drawing the gases into the cylinder and producing flows that differ from the flows produced by the inflow of gas and air at the filling of the cylinder, whereby the mixture becomes more uniform in the cylinder, and local self-ignition in areas of oily mixture is counteracted. The injection into downward clouds of high-pressure atomized water supplied at a pressure of at least 50 30 bar promotes the mixing of the cylinder contents all the way down through the cylinder, right from the injection starting, even if the piston is at a great distance from the cylinder cover.

Furthermore, the water injection made in the crank angle range from 20 ° to 150 ° after the bottom dead-end engages in the chemical reaction that determines how fast a lean given gas / air mixture self ignites. The evaporation of the injected water during the compression stroke cools the gas / air mixture and both the lower temperatures during the compression and the resulting lower final temperature at the end of the compression stroke inhibit the chemical reaction. Another independent factor for this reaction is the instantaneous pressure in the cylinder.

10 As the pressure drop due to cooling is approximately five times greater than the rise in pressure due to the formation of water vapor, the water injection causes lower pressure levels which also inhibit the chemical reaction. The combined effect of these conditions is a significant reduction in the risk of self-ignition. The water injection in said interval occurs before the mixture of air and gas is ignited by means of the foreign ignition.

This opens up a variety of possibilities, such as using engines of unchanged basic construction to use gases with lower metal numbers, or to use a smaller excess air number, ie. a fatter gas / air mixture so that the engine's fuel needs are primarily met by low pressure gas supplied to the engine's charge air and the use of costly high pressure compression of the gas is avoided.

A major advantage of the method of the invention is that it enables the use of low pressure injected gas in diesel engines having high bore, low rpm and high compression ratios such as 30 bore greater than 400 mm, rpm of, for example, 200 rpm or less at full load and a compression ratio of 1:14 or more.

It is preferred that the water be supplied to the injection nozzles at a pressure ranging from 35 150 to 500 bar. In order to obtain an effective DK 176118 B1 β "stirring" in the cylinder contents, there should be two or more injection nozzles for water and the injection pressure should be high as higher pressures are very effective in obtaining the even mixture of gas and air. The water pressure may well be greater than 500 bar, but this will involve considerable compression work. In a highly effective form, the water injection is capable of setting up at least one vortex in the cylinder contents, with the gases flowing downward near the cylinder wall and upward near the axis of the cylinder. This vortex interacts with the circular cylindrical vortex produced by the air, which is coaxial with the axis of the cylinder, and provides a highly efficient and uniform mixing of the gases in the cylinder.

15 It is advantageous to inject very large quantities of water during the compression stroke, because the effect of the water injection on the self-ignition limit is significant.

For example, the amount of water injected may amount to at least 0.75 kg per 42.7 MJ calorific value of the added fuel quantity, preferably from 0.85 kg to 2.0 or 4.0 kg thereof. It should be noted here that the calorific value of a kilogram of oil is approximately 42.7 MJ, from which it can be seen that the amount of water supplied by weight may be approximately equal to or may be greater than the amount of oil which should have been supplied if the engine had been powered. on fuel oil at that engine load. The calorific value of gas can vary somewhat depending on the gas type and composition. If the quantity is less than 0.75 kg, for large, slow-moving engines it can be difficult to avoid self-ignition. Water volumes of between 1.0 and 2.0 kg per 42.7 MJ calorific value allow for high compression ratios and / or operation with noticeable fluctuations in the gas number of gas. Very uniform and lean gas / air mixtures and a highly humidified combustion air are obtained with increased heat capacity. It is also an environmental part - that the energy consumption for high-pressure compression of gas is avoided, and the invention allows to make a gas-powered diesel engine with minimal environmental impact.

Advantageously, the water injection may be made at least in the portion of the compression stroke comprising the crank angle range from about 20 ° to 90 ° after the bottom dead-end, ie. from the time the flushing air ports are closed and in the first part of the compression stroke. The temperature drop caused by the water evaporation 10 at the beginning of the compression stroke, in itself, results in a corresponding decrease in the final temperature at the end of the compression stroke, but in addition to this purely cooling effect, a corresponding decrease in the final temperature is achieved as a result of a lower temperature rise 15 in the compression. Thus, water injection at the beginning of the compression stroke has a dual effect.

As a supplement or alternative to water injection at the beginning of the compression stroke, water may be injected at least in the portion thereof comprising the 20 crank angle range from 120 ° to 150 °, and suitably from 90 ° to 150 °, after the bottom dead point. During this angular interval, the compression has an elevated temperature level which gives a rapid evaporation of the injected water so that the water is not deposited in large quantities on the surface of the cylinder elements, and at the same time the water injection is not so late during the compression stroke that the temporal effect of the the water injection fails. It is clear that the water injection can start earlier 30 and / or end later and that it can be intermittent.

The method according to the invention is particularly advantageous if the dual fuel engine is a two stroke long flush engine with the exhaust valve located in the cylinder cover, where the gas is supplied in the region of the cylinder flush ports located in a lower cylinder section, preferably such that the gas ports are in the flush air ports and where the gas supply is opened significantly later than the purge air ports. With such a longitudinally flush engine, it is known that the gas supply is made directly into the cylinder, because gas ports are located in the area of the flushing ports and that the gas is not intentionally discharged into the exhaust system during the cylinder rinse. The preferred location of the 10 gas ports in the purge air ports provides a great advantage in terms of mixing, with the inflowing gas being delivered directly into the inflowing air and being mixed therewith immediately after delivery.

When the overall components of the engine are determined, many of the factors for the course of the chemical reaction leading to self-ignition are also determined. Nevertheless, the operation of the engine can be continuously adjusted to variations of the other factors for the course of the chemical reaction, such as variations in fuel quality and / or external conditions, by reducing the effective compression ratio in operation by adjusting the exhaust valve closing time. when the gas / air mixture, prior to ignition with foreign ignition, gets too close to its self-ignition limit, and vice versa. This continuous and instantaneous adjustment of the effective compression ratio can keep the gas / air mixture near the ignition limit at varying operating conditions without significantly affecting engine efficiency because the expansion ratio during operation is unaffected by the exhaust valve closing time during the compression stroke.

For the safety of the engine system, such a design of the gas supply to the cylinders of the engine is preferred that a gas supply line to the cylinder concerned contains a flow meter which is measured if there is an absolute gas flow during the period of the cycle in which the piston blocks the flow connection between the purging air ports and the cylinder space above the piston. With such general flow monitoring in the time window when there should not be an absolute gas flow, the monitoring can detect even small gas leaks due to a small leakage in the seat of an otherwise normally functioning control valve in the gas supply line, while providing a sudden any failure in the normal shut-off function of the control valve is immediately detected.

As an alternative or supplement to the flow monitoring, while the engine is running, 15 air samples may be taken from each engine cylinder flush box and hydrocarbon measurement may be taken on the samples and / or measured at least one cylinder parameter representative of the cylinder performance such as exhaust temperature or cylinder pressure, as well as a prior comparison. -20 specific load-dependent limit values for at least one cylinder parameter. The hydrocarbon measurement will show if there are gas leaks to the charging air outside the cylinder, ie. to the flushing box located around the bottom section of the cylinder. Since a leak causes the gas content of the air to be higher than the predetermined intended gas dosage, ie. if the gas / air mixture gets fatter, the leak can lead to elevated exhaust temperature and / or to disruptions of combustion, for example in the form of self-ignition, which can be detected as sudden, very strong increases in cylinder pressure. Thus, gas leaks can be detected by pressure and temperature measurement sensors, and these sensors are well known and well-tested in conjunction with conventionally powered diesel engines. The use of the 35 known sensors for detecting gas leaks makes it possible to achieve safe detection of a gas leak and rapid remediation, for example in the form of interrupting the gas supply to the faulty cylinder or to the entire engine.

The gas quantity dosage dependent on engine load can conveniently be regulated by the gas delivery pressure. Unlike the known dosing methods in which the dosing occurs by adjusting the injection time, the pressure controlled dosing 10 allows dosing at an approximately constant crank angle interval, which may correspond, for example, to about 2/3 of the angular range in which the rinsing air ports are open. Thus, a significantly longer gas injection can be achieved when only a relatively small amount of gas is to be injected, which gives more uniform mixing of gas and air in the cylinder. This effect is particularly important in the very lean mixtures, where there may be a risk of incineration in areas of the cylinder contents.

The invention further relates to a multi-cylinder diesel-fueled dual-fuel internal combustion engine, each having a series of purge air ports located in a lower portion of the cylinder and at least one exhaust valve in the cylinder cover as well as a valve controlled gas supply 25 having a plurality of gas ports. In order to be able to achieve a more extensive coverage of the fuel demand by means of low pressure supplied gas, also for engines of larger dimensions and at lower rpm than automotive engines, the engine according to the invention is peculiar in that in each of the cylinders with gas ports there is at least one a water injector for high pressure injection of water into the cylinder compartment between the piston and the cylinder cover, and that the engine has a water feeding system for supplying the water injectors with water at a pressure substantially greater than the instantaneous cylinder pressure. The water injection provides the advantages described above for DK 176118 B1 11.

In one embodiment, the nozzle holes of the water injectors are oriented in directions that inject the water down towards the plunger, and conveniently away from the inside of the cylinder. As mentioned, the downwardly directed clouds of atomized water provide a strong downward and upward circulation throughout the cylinder contents, with consequent improved mixing. When the nozzle holes are located at a suitable radial distance within the cylinder wall, the injection direction may be substantially parallel to the longitudinal axis of the cylinder. If the nozzle holes are close to the inside, the injection may conveniently be: oriented in a direction away from the inside of the cylinder, ie. is located downwards growing 15 distances from the inside. This gives the advantage that the water cannot strike directly into the barrel run surface of the piston rings, where a sprayed cloud of water would destroy the lubricating oil film and lead to unacceptable cylinder wear.

20 The water injection can suppress the risk of self-ignition to such an extent that the engine can be a two-stroke cross-head motor with a speed of less than 250 rpm and a cylinder bore of at least 230 mm.

Apart from the calorific value 25 in the pilot fuel, the engine's fuel needs can be covered by low-pressure injected gas at least at full load.

In a preferred embodiment, the gas ports in the purge air ports open and are supplied with gas from a ring chamber extending around the cylinder and communicating with a gas supply line containing a shut-off valve. This gives the advantage over placement of the gas ports at a higher level than the purge air ports that the gas is delivered directly into the air flowing into the cylinder, which ensures a very uniform mixing. The use of a gas chamber with gas distribution for several gas ports per cylinder also promotes the distribution of gas in the air.

The aforementioned hydrocarbon meter can advantageously be common to several cylinders which alternately deliver sampled gas quantities to the meter.

Examples of the invention are explained in greater detail with reference to the schematic drawing, in which fig. 1 is a side view of a diesel engine 10 according to the invention; FIG. 2 is a longitudinal section through a cylinder of the motor of FIG. 1; FIG. 3 shows a section of a lower region of the cylinder; FIG. 4 is an enlarged sectional view of a preferred embodiment of a gas port; FIG. 5 and 6 are diagrams of the ignition limit for a gas in engines as a function of the excess air number and shown for various operating parameters; and FIG. 7 is a diagram of pressure and temperature gradients in a cylinder, respectively, without and with water injection.

In FIG. Figure 1 shows a two-stroke diesel main engine 1 of the diesel type for propulsion of a ship or for power generation in a stationary power plant. When the motor draws a generator or is connected to a propeller with a controllable pitch, it will normally run at a predetermined fixed speed independent of the motor load. This can be considered an advantage in the present context, since the speed will not have a varying influence on the self-ignition limit. The engine can also be directly connected to a propeller not shown in a ship, in which case the speed varies with the engine load.

The engine 1 has several cylinders 2 which protrude from an engine rack 3. A common feed line 4 for pilot fuel, such as oil, supplies fuel pumps 5, which may be piston pumps, with pilot fuel, so that the pumps on the desired adjustable timing of the motor cycle can inject pilot fuel into cylinder 2 via at least one fuel injector 5 'to initiate combustion 5 in accordance with the well known dual fuel principles. Instead of foreign ignition by pilot fuel, the engine according to the invention can be designed with foreign ignition by another heat source, such as an electrically activated ignition body, which may be located in a known chamber which is fed to a grease gas. air mixture than the combustion chamber of the cylinder 6.

During the compression stroke prior to the injection of pilot fuel, the air in the cylinder is supplied with 15 main fuel in the form of gas which is supplied at an advantageously low pressure which must be higher than the cylinder pressure at the time of blow-in. For example, the gas pressure can be less than 15 bar, and often less than 10 bar, which requires only a little compression work and uncomplicated compression machines. As discussed further below, gas pressure can be variable.

The engine 1 has a water feeding system comprising a fresh water source or other suitable water type, which is supplied via a conduit 7 to a high pressure generator 8, 25 which may, for example, be a piston pump or at least one tank which is periodically filled with water and pressurized by compressed air. at a pressure corresponding to the desired water feed pressure for water injectors 9 · If water and air in the tank are desired to be kept separate from each other by means other than a membrane, the tank may be provided with some floating bodies such as air-filled balls floating on top of the water and transmits the air pressure thereto. In this way, it is possible to produce very high water pressures and to deliver large quantities of water to the cylinder without having to take into account the lubrication conditions in a pump cycle in which the piston runs partially in water. The water can be supplied to the water injectors 9 on a cylinder by opening and closing a control valve, time controlled in dependence on the motor cycle. The control valve may be electronically controlled by a control unit which receives signals for the instantaneous crankshaft position. The water pressure may be variable and increasing during a compression stroke to account for the increasing compression pressure.

Each water injector 9 comprises in a well-known manner a valve housing with an inner seat and a movable slider pressed against the seat by a spring to block the connection between a water connection and a nebulizer nozzle. When the valve is actuated to initiate the water injection, the slider is displaced away from the seat, which can be done, for example, by a control valve opening the effect of the water pressure on a downward sliding surface, whereby the water pressure displaces the slider away from the seat. An alternative option is to use control oil to displace the slider away from the seat, in which case the valve's open-closing function may be independent of the water pressure. The nebulizer holes near the lower end of the injector are directed downwardly so that the injected water does not strike directly into the cylinder wall. For example, the jet (cloud) of water emitted by a nebulizer hole may be directed downwards at an angle to the longitudinal direction of the cylinder located in the range 0-20 °. The different nebulizer holes may have different orientations to partly distribute the water 30 relatively evenly over the cylinder contents, and partly to create at least one downward swirling system near the cylinder wall and upward flow near the center axis of the cylinder. It is preferred that the nebulizer holes be oriented so that their jet directions strike 35 within the circumference of the piston when this is in the position where the flushing air ports are completely shut off.

There may also be one or more than two water injectors 9, depending on the amount of water desired to be injected and the desired distribution.

5 The engine is supplied with rinsing and charging air at a pressure which may vary with the engine load. The flushing and charging air may be supplied by one or more turbochargers, possibly. supplemented by powered auxiliary fans when the engine is running at part load or low load.

The motor may be a medium or high speed motor, but it is preferred that it is a slow-moving motor with an exhaust valve 10 located at the top of the cylinder in the cylinder cover 11 and a series of flushing ports 12 located in a lower cylinder section which is 15 is surrounded by a flushing box 13 which communicates via an opening 14 with a supply of pressurized rinsing air, such as a flushing air receiver 30, which is an elongated container common to several cylinders. The rinsing boxes 13 may be separate for each cylinder, i. may be 20 spaced apart by transverse walls between the cylinders, and may be downwardly delimited by an intermediate bottom 14 of the engine rack so that there is an effective separation of the crankcase and air supply. A piston 15 in the cylinder is mounted on a piston rod 16 which passes through the intermediate bottom of a piston rod packing box 17. In another embodiment, the flushing box may be common to several or all cylinders.

A gas source 18 supplies gas to the engine's gas system.

A pressure control unit 19 controls the gas supply pressure in a distributor line 20 connected to each cylinder 2 via a branch line 21 with one of the shut-off valves. If the gas is evaporated gas from a tank or supplied from an external supply line at low pressure, the unit 19 can make an adjustable compression of the gas. It is also possible to supply the gas from a source with somewhat higher gas pressure, in which case the unit 19 can be an adjustable pressure limiter.

When the motor, as indicated in FIG. 2 is a two-stroke, long-flush diesel engine, it is not acceptable to premix the gas in the engine's intake air because the flushing air supplied in the first part of the flushing process flows directly into the exhaust system before the exhaust valve 10 is closed and the cylinder is charged with air to it. next combustion. Therefore, the gas is supplied directly to the cylinder. A preferred embodiment of the gas supply to a cylinder is shown in FIG. 3 and 4.

The branch line 21 immediately adjacent to the cylinder has a central shut-off or control valve 15 22 which at appropriate times of the motor cycle can open and close the gas supply to a ring chamber 23 formed in the cylinder wall as a groove closed externally with a cover ring. 24. From annulus 23, several nozzle tubes 25 forward the gas to gas ports 20 located in the purge air ports. The number of nozzle pipes can be from 4 to 6, evenly distributed around the perimeter. In order to obtain the smoothest gas / air mixture, there may be a gas port in each purge air port.

In a simple embodiment, the gas port is the outlet opening of the nozzle tube, but it is also possible to incorporate a non-return valve in the nozzle tube so that hot gas return to the annulus 23 is avoided at the end of the expansion stroke. As an alternative to placing a nozzle tube in a bore, a bore from the annular chamber 23 can be made simply and inclined downwards and inwards to open in the upper wall of the rinsing air port or at the corner where the rinsing air gate passes into the inside of the cylinder. The estuary in this case constitutes the gas port.

In another embodiment not shown, a plurality of gas supply valves, which in number and distribution may be configured as above, are provided over the purge air ports, which are provided with gas from a common pipe extending around about the cylinder.

5 The gas inlet to the pipe is controlled by an electronic control valve located close to the cylinder. The gas blow-in valves are located in bores in the cylinder wall recessed relative to the inside thereof, and they can be simple nozzles with one or two holes and a counter-valve which prevents recoil at the start of the exhaust period.

In the branch line 21, there may be on the feed side of the control valve 22 a flow meter which can measure small absolute gas flows for a specified time interval.

The flow meter may comprise an aperture with associated sensor for detecting pressure drop across the aperture. The flow meter can measure the mean value of the pressure changes over the specified time interval.

The gas dosing to the cylinder can be done for a specified 20 and approximately fixed interval of the motor cycle.

If the period of open rinsing air ports, which is normally expressed in a certain crank angle rotation, is subdivided into 100 sub-periods, the gas dosing may advantageously occur during the sub-periods from about 30 to about 95.

By terminating the gas dosing, just before the purging air ports are closed by the upward piston, by means of the ejector action, exhaust gas is extracted from the annulus, so that the risk of gas leaks into the space under the piston is minimal.

The quantity control can conveniently be done by adjusting the gas supply pressure by means of the pressure control unit 19 to be more or less above the purge air pressure at the instantaneous engine load. If an attempt is made to keep the excess air rate constant, which is usually about λ = 18 OK 170118 B1 2, the fuel quantity in the cylinder depends on the amount of air which in turn depends on the purge air pressure. As the purge air pressure on engines with exhaust-driven turbochargers varies with the engine load, it may be necessary to blow off purge air after the compressor at certain 5 engine loads or rapid load changes to control the engine power, including making quick power changes.

FIG. 5 illustrates self-ignition limits for different engine sizes as a function of the air excess number λ. The fully drawn curve a shows the conditions of a mid-speed motor with a speed of 600 rpm. It is seen that at an air excess number of λ = 2, self-ignition occurs when the piston reaches the top dead center, 15 and it is also seen that the self-ignition limit is theoretically raised if the gas / air mixture is made leaner (larger λ). However, the very lean mixtures cannot be used in practice because the combustion becomes progressively unstable and incomplete by increased thinning. Since the foreign ignition 20 usually occurs in an interval of 5 ° before and after the peak dead-end, the self-ignition limit should preferably not be lower than 5 ° before the peak dead-end.

The curve shown a applies to a specific gas. If a gas of the same metal number is used in a larger engine 25 with a bore of 350 mm and an associated lower speed of 209 rpm, the self-ignition limit as shown by curve b goes up. It is clearly seen that the effect of increasing λ is considerably less at the lower rpm, and the self-ignition limit is too low to allow proper engine operation. Similarly, curve c for an engine with a speed of 123 rpm and a bore of 600 mm, and curve d for an engine with a speed of 94 rpm and a bore of 900 mm show that the conditions are further deteriorated on even larger 35 motors.

DK 176118 B1 19 In fig. 6, the significance of changing the operating conditions of the engine is shown on the basis of the motor with the bore 350 mm and the speed of 209 rpm.

The curve b is here fully drawn. Curve e shows the effect of reducing the cylinder pressure by 10 bar, which can be done, for example, by reducing the effective compression ratio of the engine by keeping the exhaust valve open a little longer, so that the compression starts later in the motor cycle. The self-ignition limit is seen to be 10 improved, but far from sufficient. Therefore, changes in the effective compression ratio are only suitable for making continuous fine-tuning of the self-ignition limit to compensate, for example, minor variations in the fuel count.

The effect of injecting water according to the invention is far more marked, as exemplified by the curves f and g, where a quantity of water corresponding to 100% and 125% of the equivalent fuel oil quantity is injected, ie.

20 respectively 1.0 kg and 1.25 kg of water per 42.7 MJ of fuel added. Water injection is seen to have a significant growing effect at larger volumes of water.

In FIG. 7 shows how cylinder pressure p1 and cylinder temperature TCy .-6 extend during the compression stroke of the engine with the bore of 350 mm and the speed of 209 rpm, shown respectively without water injection at the curves h, with a water injection corresponding to 100%. of the equivalent fuel oil quantity at the pressure curve j and the temperature curve k. It is seen that the significant effect of the water injection occurs during the last 90 ° rotation before the peak dead-end.

If the injected water rate (amount per time) is varied during the compression stroke, for example by intermittent injection or by varying the injection pressure of the water, it is preferable that the greatest water rate occurs in the angular range from 90 ° to 150 ° after the bottom dead point.

The present invention, by means of powerful water injection directly into the entire cylinder content during the compression stroke in quantities comparable to or greater than the fuel volumes, allows even large two-stroke diesel engines to operate on low-pressure injected gas as main fuel. so high pressure injected fuel should only be used 10 as ignition aid. The invention can also be used to use lower-metal gas in smaller gas / diesel engines running at higher rpm.

Claims (15)

A method of operating a pressurized dual fuel combustion engine (1) of the diesel type, which is supplied in the cylinders (2) with air and gas in a mixing ratio, after which the air and gas are compressed during the compression stroke and ignited by foreign ignition, characterized by that the air to gas mixing ratio gives an air excess number of at least λ = 1.3 that with a number of injection nozzles (9) per cylinder at least in a portion of the crank angle of the compression stroke from 20 ° to 150 ° 15 after the bottom dead point in downward clouds directly into the mixture of air and gas in the cylinder space between the upward piston (15) and the cylinder cover (11), and that the water is supplied to the injection nozzles at a pressure of at least 50 bar. 20 Process according to claim 1, characterized in that the water is supplied to the injection nozzles at a pressure in the range of 150 to 500 bar. 25 30 35 DK 176118 B1 21 Method according to claim 1 or 2, characterized in that the injected amount of water is at least 0.75 kg per 42.7 MJ calorific value of the fuel quantity supplied, preferably from 0.85 kg to 2.0 or 4.0 kg thereof. . Method according to one of claims 1-3, characterized in that at least water is injected into the part of the compression stroke comprising the crank angle range from 20 ° to 90 ° after the bottom dead point. Method according to one of claims 1-4, characterized in that at least water is injected into the part of the compression stroke comprising the crank angle range from 120 ° to 150 °, and suitably from 90 ° to 150 °, after bottom dead center. Method according to one of claims 1 to 5, characterized in that the dual fuel motor (1) is a two stroke long flush motor with the exhaust valve (10) located in the cylinder cover (11), the gas being supplied in the area around the rinsing air ports of the cylinder (2). (12) located in a lower cylinder section, preferably such that the gas ports are in the flush air ports and that the gas supply is opened substantially later than the flush air ports. Method according to claim 6, characterized in that the effective compression ratio is reduced in operation by adjusting the closing time of the exhaust valve (10) when the gas / air mixture, prior to ignition with foreign ignition, gets too close to its self-ignition limit. vice versa. Method according to one of claims 1-7, characterized in that a gas supply line for the cylinder in question contains a flow meter, which measures whether there is an absolute gas flow during the 35 period of the cycle in which the piston (15) shuts off DK 176118 B1. 22 the flow connection between the purging air ports (12) and the cylinder space above the piston. Method according to one of claims 1-8, characterized in that, while the engine is running, 5 samples are taken from each engine cylinder flushing box (13) and hydrocarbon measurement is performed on the samples and / or at least one cylinder parameter representative of the cylinder performance, such as exhaust temperature or cylinder pressure, and a comparison is made with 10 predetermined load-dependent limit values for the at least one cylinder parameter. Process according to one of claims 1-9, characterized in that the gas quantity dosage dependent on the engine load is controlled by the gas delivery pressure. 11. A multi-cylinder diesel fuel pressurized dual fuel combustion engine (1), each having a series of purge air ports (12) located in a lower portion of the cylinder and at least one exhaust valve (10) in the cylinder cover (11) and with a valve controlled gas supply having a plurality of gas ports, characterized in that in each of the gas port cylinders there is at least one high pressure water injector (9) for injecting water into the cylinder space between the piston and the cylinder cover, and that the engine has a water supply system for supplying the water injectors (9) with water at a pressure substantially greater than the instantaneous cylinder pressure. Pressure-loaded dual fuel internal combustion engine according to claim 11, characterized in that the nozzle holes of the water injectors (9) are oriented in directions that inject the water downwards towards the piston (15) and suitably away from the inside of the cylinder. Pressure-loaded dual fuel internal combustion engine according to claim 11 or 12, characterized in that the 35 engine is a two-stroke cross-head motor with a speed of less than 250 rpm and a cylinder bore of at least 230 mm. Pressurized dual fuel internal combustion engine according to one of claims 11-13, characterized in that the gas ports open into the purge air ports and are provided with gas from a ring chamber extending around the cylinder and communicates with a gas supply line containing a shut-off valve. Pressurized dual fuel internal combustion engine according to one of claims 11 to 14, characterized in that the flushing air ports of each cylinder are surrounded by a flushing box separate from the cylinder, which is provided with flushing and charging air from a flushing air receiver, and that the flushing boxes each have their own air sampler. 15 delivering the air samples to a hydrocarbon meter common to several cylinders.