CN100554652C - Crosshead type large two-stroke diesel engine and control valve and application thereof - Google Patents

Abstract

The invention relates to a large two-stroke engine (1) of the crosshead type, in which the exhaust valves (11) are hydraulically actuated with high-pressure hydraulic fluid by providing a hydraulic actuator (19) associated with each exhaust valve (11). Fuel or heavy fuel oil can be used as the hydraulic fluid. The invention also relates to a control valve (25) for use in such an engine (1), a hydraulic system for use in such an engine (1) and a hydraulic actuator (19) for use in such an engine (1).

Description

Crosshead type large two-stroke diesel engine and its control valve and application

technical field

The present invention relates to large two-stroke diesel engines of the crosshead type in which each exhaust valve is hydraulically actuated by supplying high pressure hydraulic fluid to a hydraulic actuator associated with the exhaust valve.

Background technique

Large two-stroke diesel engines of the crosshead type are commonly used in marine propulsion and as prime movers in power plants. These internal combustion engines are constructed differently from any other internal combustion engine not only because of their overall size. The two-stroke principle and the use of heavy fuel oil with a viscosity below 700cSt at 50°C (which does not flow at room temperature) put them in a class of their own in the engine world.

In many conventional engines of this type, the exhaust valves and fuel injection system are actuated by rotating cams coupled directly to the engine's crankshaft. Two-stroke engines use gas exchange ports to control the intake to the cylinders, so intake timing is rigidly related to crank angle. This just makes the control of the exhaust valve and fuel injection more flexible.

The demands on fuel consumption, reliability and power output of such engines are extremely high. Recently, environmental requirements have led to a demand to reduce exhaust emissions. Meeting these sometimes conflicting requirements requires comprehensive and flexible control over fuel injection timing and quantity, as well as control over opening and closing timing, relative to conventional rotary cam-actuated exhaust valves and fuel injectors. And the opening of the exhaust valve is fully and flexibly controlled.

Due to the size of such engines, electrical actuators cannot be used to operate the exhaust valves. In the largest such engines, such exhaust valves can weigh up to 450Kg.

The ME engine series from MAN B&W Diesel is a large two-stroke diesel engine of the crosshead type with electrohydraulic control of exhaust valves and electrohydraulic actuation of fuel injection systems. The hydraulic system works with oil from the engine lubrication system. The lube oil system is operated by a low pressure pump at 3-4 bar. Another high-pressure type pump delivers lubricating oil to the common rail at a pressure of about 200 bar. Lubricant oil from the common rail is directed through a hydraulic valve to the fuel booster, which increases the common rail's 200 bar pressure to the required 1000 bar pressure in the fuel line. The fuel lines are heated to 90-150°C to ensure the fuel flows and has the proper viscosity. Lubricant oil from the common rail is directed through timing valves to the hydraulic exhaust valve actuators to operate the exhaust valves.

However, the lubricating oil from the lubricating systems of these engines is not clean enough for use in common rail hydraulic systems. The oil thus needs to be filtered to remove any particles above 5-10μ before it can be pumped to the common rail.

/Sulzer RT-flex系列发动机是具有电液控制式排气门和电液启动式燃料喷射系统的十字头型大型两冲程柴油发动机。用于气门致动的液压系统以专用液压油操作。润滑系统完全与液压系统分开。

/Sulzer RT-flex series engines are large crosshead two-stroke diesel engines with electro-hydraulic controlled exhaust valves and electro-hydraulic activated fuel injection systems. The hydraulic system for valve actuation operates with special hydraulic oil. The lubrication system is completely separate from the hydraulic system.

EP 1 130 251 discloses a pump arrangement for providing an accumulator of a common rail system used in a large two-stroke diesel engine, the pump arrangement having: at least two pumps for delivering fluid into the accumulator and an intermediate accumulator for buffering the dynamic pressure component, wherein each pump is connected to the intermediate accumulator via a separate pump line. Three pumps and lines with check valves connected at the ends to the intermediate vessel can be used. The valves that control the flow of fuel from the common rail to the individual cylinders are simple on/off valves.

DE 103 11 493 discloses a large two-stroke diesel engine having at least one cylinder provided with a discharge port in which a piston controlling the discharge port is slidably accommodated. An outlet valve is actuated by a hydraulic actuator to close the exhaust outlet, and a proportional valve is shared by the hydraulic actuator and the fuel injection pump. The proportional valve has an inlet connected to a high pressure fluid source and two outlets connected to a hydraulic actuator and fuel injection pump.

EP 1 471 236 discloses a fuel supply system and a fuel supply method for an automobile engine. In order to carry out fuel injection through sufficient expansion at a fuel pressure as high as required in the operating range from startup to self-sustaining operation, the fuel supply fuel system for direct fuel injection internal combustion engines is provided with a high-pressure fuel pump to directly Fuel, which has been pressurized by a high-pressure fuel pump, is injected from the injector into the combustion chamber of the engine. An electric motor is provided to assist in driving the high pressure fuel pump. Driving of the high-pressure fuel pump is performed or facilitated by an auxiliary power unit, such as an electric motor, when the engine is started.

GB 2 102 065 discloses a pneumatic biasing device comprising a volume of gas which is compressed due to valve opening. The compressed gas volume acts on a cap secured to the valve stem to exert a force in the longitudinal direction of the valve stem, thereby biasing the valve to its closed position in which the valve head is squeezed against the valve seat.

WO0012895 discloses a system for decelerating a linearly moving valve undergoing a closing movement. The system includes a housing, a first hydraulic fluid chamber disposed in the housing, and a slave piston for moving the valve in accordance with hydraulic fluid supplied into the first hydraulic fluid chamber. The deceleration of the valve may be accomplished by selectively throttling hydraulic fluid released from the first chamber to the second chamber. The hydraulic fluid in the second chamber resists the closing movement of the valve, gradually slowing it down to achieve valve seating. Gradual throttling is used to maintain a nearly constant hydraulic pressure in the second chamber during seating. Gradual throttling can be achieved by selecting a suitable size and shape of the throttling holes and a proper throttling distribution for the holes.

US 2002/0184996 discloses an actuator comprising a cylinder, first, second and third openings, an actuating piston, a control piston and a control spring. The cylinder defines a longitudinal axis and includes first and second ends. A first opening communicates with the first end of the cylinder, a second opening communicates with the second end of the cylinder, and a third opening communicates with the cylinder between the first and second ends. An actuating piston is disposed in the cylinder and is movable in first and second directions along the longitudinal axis. The actuation piston includes first and second sides. A control piston is also arranged in the cylinder and is movable in first and second directions along the longitudinal axis. The control piston includes first and second sides, the first side of the control piston facing the second side of the actuation piston. A control spring biases the control position in at least one of first and second directions. A method of controlling the actuator is also provided.

Contents of the invention

Based on the above background, it is an object of the present invention to provide a large two-stroke diesel engine of the crosshead type with improved control over fuel injection.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; an associated hydraulic actuator for each of said exhaust valves; a common fuel rail having one or more accumulators connected thereto; a high-pressure fuel pump supplying fuel at high pressure to said common fuel rail, each of said fuel injectors being fed from said common fuel rail fuel operation; and a proportional valve associated therewith with each cylinder whereby the proportional valve controls the flow of fuel from the common fuel rail to the individual fuel injectors.

The use of proportional control valves allows more precise and flexible control of the fuel injection timing and dosage curves. Furthermore, the use of proportional control valves allows rate shaping and pre-injection to be achieved without additional equipment, eg rate shaping is achieved substantially entirely through the control signal sent to the proportional valve.

Another object of the present invention is to provide a large two-stroke diesel engine of the crosshead type with improved control over exhaust valve actuation.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; a hydraulic actuator associated with each of said exhaust valves; a common fuel rail, having one or more accumulators connected thereto; a high pressure fuel pump supplying fuel at high pressure to said common fuel rail, each of said fuel injectors operating on fuel from said common fuel rail and a proportional control valve associated with each cylinder, whereby the proportional control valve controls the flow of fuel from the common fuel rail to the respective hydraulic actuators.

The use of proportional control valves provides adequate and flexible control in opening and closing timing and the degree of opening of the exhaust valves. Also, the position of each cylinder's exhaust valve is controlled in a flexible manner, allowing eg a particular cylinder's exhaust valve to open slightly during its compression stroke to facilitate engine starting.

Another object of the present invention is to provide a large two-stroke diesel engine of the crosshead type with an overall simpler and more flexible hydraulic system.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; a hydraulic actuator associated with each of said exhaust valves; a common fuel rail, having one or more accumulators connected thereto; a high pressure fuel pump supplying fuel at high pressure to said common fuel rail, each of said fuel injectors operating on fuel from said common fuel rail , the hydraulic actuators are connected to the common fuel rail via respective supply conduits, wherein the supply conduits are provided with heating means.

Thus, when heavy fuel oil, also called HFO, etc. is used as a hydraulic medium, the HFO is maintained at an appropriate viscosity.

Another object of the present invention is to provide a large two-stroke diesel engine of crosshead type with a hydraulic exhaust valve actuation system capable of operating over a wide temperature range.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; a hydraulic actuator associated with each of said exhaust valves; a common fuel rail, having one or more accumulators connected thereto; a high pressure fuel pump supplying fuel at high pressure to said common fuel rail, each of said fuel injectors operating on fuel from said common fuel rail , each of said hydraulic actuators operates on fuel from said common fuel rail, wherein said hydraulic actuators are provided with means for compensating for dimensional changes caused by:

operating at different temperatures,

Trimming, such as lapping of valve seats, and

manufacturing tolerances.

Thus, the hydraulic actuator will assume the proper position over a wide temperature range and ensure that the valve head reaches the valve seat correctly at all times.

Another object of the present invention is to provide a method of controlling the temperature of fuel in the supply line of a crosshead type large two-stroke diesel engine.

This object is achieved by providing a method of controlling the temperature of fuel in a pressure line of a large two-stroke diesel engine of the crosshead type, said pressure line connecting a common fuel rail to a hydraulic actuator of an exhaust valve, and the method Including the step of controlling the temperature of the fuel entering the pressure conduit to maintain a temperature gradient of said fuel below a predetermined threshold during a change in the operating temperature of the fuel.

Thus, the fuel can be used as a hydraulic medium for operating the hydraulic actuators of the exhaust valves which are sensitive to changes in operating temperature.

Another object of the present invention is to provide a crosshead type large two-stroke diesel engine having a hydraulic exhaust valve actuation system capable of operating with a variety of hydraulic fluids.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; a hydraulic actuator associated with each of said exhaust valves; a common fuel rail, having one or more accumulators connected thereto; a high pressure fuel pump supplying fuel at high pressure to said common fuel rail, each of said fuel injectors operating on fuel from said common fuel rail , the hydraulic actuators are connected to the common fuel rail and ultimately to other hydraulic components such as valves through corresponding hydraulic lines, characterized in that the connections between the pipelines of the engine and other hydraulic components are sealed The static gasket in the valve actuator and the dynamic gasket in the valve actuator are made of the following materials: cast iron, steel, polytetrafluoroethylene (PTFE), fluoroelastomer, (FPM), copolymer (NBR), nitrile rubber, poly( Dimethicone) (SI) or combinations and/or mixtures thereof.

Selection of the gasket from these materials allows non-proprietary hydraulic fluids, such as fuel, to be used in the hydraulic system without causing damage to the gasket.

Another object of the present invention is to provide a crosshead type large two-stroke diesel engine having a hydraulic exhaust valve actuation system capable of operating with various hydraulic fluids.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; a hydraulic actuator associated with each of said exhaust valves; a common fuel rail, having one or more accumulators connected thereto; a high pressure fuel pump supplying fuel at high pressure to said common fuel rail; for delivering fuel from said common fuel rail to respective fuel injectors, a supply conduit and valve means associated with each cylinder; a supply conduit and valve means associated with each cylinder for delivering fuel from said common fuel rail to a respective hydraulic actuator; and a heated return conduit, A conduit for delivering fuel from said hydraulic actuator to a fuel tank or to an inlet leading to said high pressure fuel pump.

Thus, HFO having a low viscosity can be used as a hydraulic medium.

Another object of the present invention is to provide a crosshead type large two-stroke diesel engine with a hydraulic exhaust valve actuation system capable of operating in a wide temperature range.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder being provided with at least one fuel injector and at least one exhaust valve; a hydraulic actuator associated with each of said exhaust valves; a common fuel rail, having one or more accumulators connected thereto; a high pressure fuel pump supplying fuel at high pressure to said common fuel rail; associated with each cylinder for feeding fuel from said common fuel rail to the respective fuel injector pressure conduits and valve arrangements; supply conduits and valve arrangements associated with each cylinder for delivering fuel from said common fuel rail to the respective hydraulic actuators; and return conduits for a conduit conveying fuel from said hydraulic actuator to a fuel tank or to an inlet to said high-pressure fuel pump, wherein at least one of said conduits includes a tube for smoothing out variations in conduit size due to changes in operating temperature Affected devices.

Thus, the hydraulic system will be able to operate over a large temperature range and ensure no mechanical stress due to temperature-induced dimensional changes.

Another object of the present invention is to provide a new application of the proportional valve.

This object is achieved by providing a use of a proportional valve for controlling the flow of fuel from a common fuel rail of a large two-stroke diesel engine of the crosshead type to fuel injectors and/or fuel running components.

Another object of the present invention is to provide an electrically controlled valve for controlling the flow of fuel from a common fuel rail of a crosshead type large two-stroke diesel engine to one or more fuel-running or fuel-consuming engine components.

This object is achieved by providing an electrically controlled valve for controlling the flow of fuel from a common fuel rail of a large two-stroke diesel engine of the crosshead type to one or more fuel-running or fuel-consuming engine components, The electrically controlled valve includes a valve body and a solenoid, whereby the solenoid is thermally insulated from the valve body.

Another object of the present invention is to provide a crosshead type large two-stroke diesel engine with improved hydraulic system circulation when the engine is stopped.

This object is achieved by providing a crosshead type large two-stroke diesel engine comprising: a crankcase bracket supporting a crankshaft and a cylinder bracket mounted on the crankcase bracket; a plurality of cylinders supported by said cylinder bracket, each cylinder provided with at least one fuel injector and at least one exhaust valve; a common fuel rail; and a high pressure fuel pump which supplies fuel under high pressure to the said common fuel rail; supply conduits and valve arrangements associated with each cylinder for delivering fuel from said common fuel rail to respective fuel injectors; said high pressure fuel pump is driven by said crankshaft during engine operation Driven mechanically and electrically by an electric motor during engine stop to circulate fuel at low pressure through said supply conduit and/or said common fuel rail and/or through other engine components running on fuel.

By using a high pressure pump as a high pressure source as well as a low pressure source to provide circulation during engine stop, parts count is reduced making the overall construction and maintenance cost more competitive.

Another object of the present invention is to provide a hydraulically actuated gas exchange valve for use in an internal combustion engine with an improved air spring.

This object is achieved by providing a hydraulically actuated gas exchange valve for an internal combustion engine, said hydraulically actuated gas exchange valve comprising: a stationary valve body; and comprising an elongated valve stem having a valve head at one end and a free end at the opposite end; a hydraulic actuator comprising a valve stem acting on the free end of said valve stem a piston for urging the valve to a non-seated position when the hydraulic actuator is supplied with pressurized hydraulic fluid; an air spring for urging the valve to the seated position, the The air spring comprises: a cylinder fixed to the valve stem, the cylinder being closed towards the free end of the valve stem and open towards the valve head; and a matching stationary piston which Housed in the cylinder, the piston is fixed to the valve body and together with the cylinder forms a spring chamber for the air spring.

The configuration of the air spring reduces the chance of hydraulic medium from the actuator entering the spring chamber.

Another object of the present invention is to provide an improved hydraulically actuated gas exchange valve for an internal combustion engine.

This object is achieved by providing a hydraulically actuated gas exchange valve for an internal combustion engine comprising: a stationary valve body; The valve moves between an open non-seated position and includes an elongated valve stem having a valve head at one end and a free end at an opposite end; a hydraulic actuator comprising a piston acting on the free end of the valve stem for urging the valve to a non-seated position when the hydraulic actuator is supplied with pressurized hydraulic fluid; an air spring which urges the valve Urging to the seated position, wherein the stroke length of the valve in the opening direction is determined by the balance of opposing forces of the hydraulic actuator and the air spring.

Thus, the actuator does not need to be provided with an end-of-stroke stop at the end of the opening stroke, and does not need to abruptly cut off the supply of high-pressure hydraulic fluid at the end of the opening stroke. The absence of an end-of-stroke stop reduces mechanical loads and vibrations, while the lack of abrupt shutoff of high-pressure hydraulic fluid avoids potentially damaging hydraulic shock waves.

Additional objects, features, advantages and characteristics of the large two-stroke diesel engine and its method of operation will become apparent from the following detailed description.

Description of drawings

In the following detailed description of the invention, the invention will be described in more detail with reference to exemplary embodiments shown in the accompanying drawings, in which

Figure 1 shows a front view of the outline of a cylinder in a two-stroke crosshead engine with a cylinder head,

Figure 2 shows a cross-sectional view of the cylinder profile in the engine shown in Figure 1,

Figure 3 is a schematic diagram of the hydraulic and lubrication system shown in Figure 1,

Figure 4 is another embodiment of the schematic diagram of the hydraulic and lubrication system shown in Figure 3,

Figure 5 shows a cross-sectional view of the pressure tube used here,

Figure 6 is another cross-sectional view showing an alternative pressure tube as used herein,

Figure 7 is a longitudinal sectional view of a first embodiment of the hydraulically actuated exhaust valve of the cylinder of Figure 2 with the valve seated and the piston in its retracted position,

Figure 8 shows a typical opening curve of a hydraulic actuator of an exhaust valve according to the invention,

Figure 9 shows on a larger scale a cross-sectional view of the actuator shown in Figure 7 with the piston in a partially extended position,

Figure 10 shows the same view as Figure 9 with the piston in the retracted position and the piston cap in the uppermost position in its axial extent,

Figure 11 is a detailed view of the upper part of the actuator with the piston approximately in the retracted position and the piston cap approximately at the uppermost position of its axial extent,

Figure 12 is a detailed view of the upper part of the actuator with the piston in the retracted position and the piston cap approximately at the lowest position of its axial extent, and

Figure 13 is a detailed view of the upper portion of the actuator with the piston in the retracted position and the piston cap approximately midway in its axial extent.

Detailed ways

Figure 1 shows an engine 1 according to the invention. The engine is a low speed two stroke crosshead diesel engine which may be a propulsion engine in a ship or a prime mover in a power plant. These engines typically have 6 to 16 in-line cylinders. The engine is built on a foundation 2 with main bearings for the crankshaft 3 . The base is divided into sections of suitable size according to the available manufacturing equipment. A crankcase bracket 4 of welded design type A is installed on the base. The cylinder bracket 5 is mounted on top of the crankcase bracket 5 . Tension screws (not shown) connect the base to the cylinder bracket and hold the structure together. The cylinder 6 is carried by the cylinder bracket 5 .

FIG. 2 shows the cylinder 6 of the internal combustion engine. The cylinder 6 is of the unidirectional flow type and has an exhaust port 7 in an air box 8 which is supplied with exhaust gas supercharged by a turbocharger 10 ( FIG. 1 ) from an exhaust gas receiver 9 ( FIG. 1 ). . A crosshead, not shown, connects the piston rod 14 to the crankshaft 3 ( FIG. 1 ).

The exhaust valve 11 is mounted centrally in the cylinder head 12 on top of the cylinder. At the end of the expansion stroke, the exhaust valve 11 opens before the engine piston 13 passes down the exhaust port 7, whereby the combustion gases in the combustion chamber above the piston 13 flow out through the exhaust passage 16 leading to the exhaust gas receiving device 17 , and the pressure in the combustion chamber 15 is released. The exhaust valve 11 closes again at an adjustable time during the upward movement of the piston 13 , which may depend, for example, on the effective compression ratio required for the subsequent combustion. During this closing movement, the exhaust valve is driven upwards by the air spring 18 .

The exhaust valve 11 can advantageously be controlled very precisely in view of the durability of the valve 11 and in view of an advantageous, precise control of the conditions in the combustion chamber and thus the efficiency of the engine.

The exhaust valve 11 is opened by means of a hydraulic actuator 19 . Hydraulic fluid (fuel) is supplied through a pressure conduit 20 connecting an inlet on the actuator 19 to a control port on the upper surface of a distribution block 21 supported by a console 22 . A return conduit 43 connects the outlet on the actuator 19 to the return opening on the upper surface of the distribution block 21 .

Each cylinder 6 is provided with two or three fuel injectors 23 (only one shown) connected by an annular duct (not shown). Fuel is supplied from distribution block 21 to fuel injector 23 via supply pipe 24 . The fuel injector 23 is connected to a return port on the distribution block 21 via a return line 49 .

The console 22 is connected to a return line to the supply line and to a common fuel rail (40 in Figure 3, not shown in Figure 2).

The distribution block 21 carries a proportional control valve 25 which controls the port and return line (43 in FIG. 3 ) on the top of the distribution block 21 and the common fuel rail 40 ( FIG. 3 ) in the console 22 (not shown) Connection.

In console 22 , passage 41 ( FIG. 3 ) branching from common fuel rail 40 conveys pressurized hydraulic fluid to an inlet on proportional control valve 25 .

Fuel in common fuel rail 40 ( FIG. 3 ) is used as hydraulic fluid to drive valve actuators 19 and fuel injectors 23 . The pressure in the common fuel rail 40 varies according to the operating state of the engine 1 such as operating speed and load conditions. Typically, the pressure in the common fuel rail 40 varies between 600 bar and 2000 bar.

Each cylinder 6 of the engine 1 is associated with an electronic control unit 26 which receives the general synchronization and control signals via a cable 27 and which, among other things, transmits electronic control signals via a cable 28 to a proportional control valve 25 . There may be one control unit 26 for each cylinder, or several cylinders may be associated with the same control unit (not shown). The control unit 26 may also receive signals from an overall control unit (not shown) common to all cylinders.

Referring to FIG. 3 , the hydraulic system and lubrication system of the engine 1 are shown in block diagram form. The hydraulic system serves as the fuel injection system and as the exhaust valve actuation system.

The lubrication system consists of a lubricating oil tank, filter and electrically driven low pressure pump. The lubrication system is completely separated from the hydraulic system.

The hydraulic system runs on fuel, typically HFO (water-emulsified and non-emulsified). Water is often emulsified into HFO to reduce NOx emissions. This emulsification takes place in a separate emulsification unit (not shown). Fuel for engine operation is stored in the heating tank 29 . The fuel used is commonly known as Heavy Fuel Oil (HFO), which has a viscosity of 500-700 cSt at 50°C and is non-flowable at room temperature. The HFO in the tank was kept at about 50°C substantially all the time, ie also during engine stop. Typically, ships with this type of engine are provided with a generator set (Genset), ie a small diesel engine that provides electricity and heat to the ship and the main engine during periods when the main engine is stopped.

From the heating box the HFO is passed to a filter or centrifuge 30 and to a pre-heater 31 . The temperature of the HFO leaving the pre-heater 31 is controlled according to the operating state and level of the HFO. During engine stop, when HFO is circulated at low pressure in the hydraulic system, the temperature of HFO is kept in the range of 45-60°C. During engine operation, the temperature of the HFO leaving the pre-heater 31 is maintained between 90-150° C., depending on the viscosity of the HFO. A sensor (not shown) measures the viscosity of the HFO immediately downstream of preheater 31 (or other suitable location). The temperature of the HFO leaving the pre-heater 31 is typically controlled such that the viscosity at the point of measurement is in the range of 10-20 cSt.

A forked intermediate conduit 32 connects the pre-heater to a high-pressure fuel pump 33 and an auxiliary low-pressure flow-through pump 34 . A check valve 35 is provided downstream of each pump to prevent suckback.

During engine operation, high pressure fuel pump 33 is driven by gear 33 on crankshaft 3 via gear 37 . Thus, the high pressure fuel pump 33 produces a nominal pressure of 1000-1500 bar, but this pressure can fluctuate between 600-2000 bar depending on operating conditions.

During engine stop, the auxiliary low pressure flow pump 34 is driven by the electric motor 38 . Thereby, a pressure of about 3-10 bar is delivered to circulate HFO in the hydraulic system during engine stop.

A common fuel rail 40 extends along all cylinders and the connection to cylinders 6 not shown in Figure 3 is indicated by short vertical lines protruding from this common fuel rail.

The cylinders 6 shown in FIG. 3 are supplied with HFO through a supply duct 41 branching from the common fuel rail 40 and leading to the inlet of the proportional control valve 25 . The supply conduit 41 is provided with a number of fluid accumulators 42 which deliver most of the fluid when the proportional control valve 25 is open and are post-feed from the common fuel rail 40 when the proportional control valve 25 is closed.

A pressure line 20 connects one of the two outlets of the proportional control valve 25 to the inlet of the hydraulic actuator 19 . A supply line 24 connects the other of these two outlets to a fuel injector 23 . The two control ports on the proportional control valve 25 are connected by channels in the distribution block to corresponding discharge ports on the upper surface of the distribution block. The proportional control valve 25 also has two outlets connected to a return line 43 for used hydraulic fluid (HFO).

The proportional control valve 25 is an electromagnetically driven servo valve with three positions. Electromagnet 44 receives control signals from control unit 26 ( FIG. 2 ) via cable 28 . The electromagnet 44 is mounted to the housing of the proportional control valve 25 with a ceramic plate 45 interposed therebetween to insulate the electromagnet 44 from the proportional control valve 25 which may reach a temperature exceeding 150° C. during engine operation. This configuration prevents the sensitive electromagnet 44 from overheating. According to another embodiment (not shown), the electromagnet 44 is connected to the valve body by means of an insulating gasket.

In the central position, in which the electromagnet 44 is deactivated, the two inlets of the proportional control valve 25 are closed and the two outlets of the proportional control valve 25 are connected to the return line 43 . When the solenoid is actuated to push the valve spool to the left (the left side in Fig. 3), the inlet of the proportional control valve is connected to the pressure line 20, to which high-pressure HFO is delivered, so that the actuation The device 19 opens the exhaust valve 11. In this position, the supply conduit 24 is connected to the return conduit 43 . When the electromagnet 44 is activated to push the spool valve to the right (the right side in FIG. 3 ), the inlet of the proportional control valve 25 is connected to the supply pipe 24 to which high-pressure HFO is delivered so that the fuel injector 23 Fuel is injected into the combustion chamber 15 . In this position, the pressure line 20 is connected to the return line 43 . The timing of fuel injection, the volume of fuel injected and the shape of the injection pattern are controlled by proportional valves. According to another preferred embodiment, the flow of fuel (not shown) from the common fuel rail to the fuel injectors is controlled by on/off valves. The on/off valve may be a separate valve from the valve controlling flow to and from the hydraulic actuator.

The separate valves that control flow to and from the actuator may also be on/off valves.

A conventional fuel limiter 46 is placed in the supply line 24 to prevent excess HFO from entering the cylinders if the proportional control valve is erroneously opened for too long.

The pressure in the return line 43 is maintained at an overpressure of several bars to avoid the penetration of air into the hydraulic system and to prevent the formation of steam bubbles from the water in the water emulsified HFO. A pressure control valve 47 at the downstream end of the return line 43 ensures that a predetermined minimum overpressure is maintained in the return line 43 . The overpressure in the return line 43 is preferably 3-10 bar. An accumulator or expansion vessel 48 is connected to the return line 43 to absorb pressure fluctuations that may occur when the proportional control valve 25 changes position.

A second return line 49 connects the outlet of the fuel injector 23 to the return line 43 . Downstream of the pressure control valve 47, a return line 43 supplies spent HFO to the preheater 31 to complete the cycle.

The pipes that carry HFO from the outlet of the pre-heater 31 to the common fuel rail 40 and from the common fuel rail 40 to the hydraulic actuators 19 and fuel injectors 23 via the proportional control valve 25 are provided with the heating shown in FIG. A coil represents a heating device. The pipe can be heated over its entire length, for example by a steam insulated heater or by electric heating elements. The heating of these tubes serves to minimize heat loss as the hot HFO travels downstream from the pre-heater. During engine operation, the temperature of the HFO in the conduits towards the fuel injectors and hydraulic actuators remains close to 150°C, however this depends on the viscosity of the HFO used. Adjacent conduits running in parallel over part of their length, such as the pressure conduit 20 and the supply conduit 24, can be provided with common heating means (not shown).

The return lines 43 and 49 are also provided with heating means of the same type as described above. The temperature of the HFO in the return line is less critical and the heating means are adjusted to ensure that the temperature of the HFO does not drop below 50°C.

During engine stop, the HFO is circulated through the hydraulic system by the circulation pump 34 (at a lower pressure of 3-10 bar) to avoid air entering the hydraulic system and to avoid localized cooling and hardening of the HFO. During engine stop, the temperature of the oil leaving the pre-heater 31 is set at about 50°C to avoid solidification of the HFO.

During flow, the proportional control valve changes position periodically in order to reach both the fuel injector 23 and the hydraulic actuator 19 . According to another embodiment, the proportional control valve is provided with a fourth bypass position (not shown). In this position, the proportional control valve is open to both the fuel injector and the hydraulic actuator. According to another embodiment (not shown), separate bypass valves are provided which allow simultaneous flow of HFO from the common fuel rail to the fuel injectors and hydraulic actuators.

The heated pressure line 20 is provided with means so that the line can operate in a temperature range from 50°C during circulation to about 150°C during engine operation. After the engine is stopped and the temperature of the HFO rises from about 50°C to about 150°C, thermal expansion causes the length of the pressure conduit 20 to increase, and vice versa.

As shown in FIG. 5 , the pressure pipe 20 is provided with one or more U-shaped sections 50 , which can absorb length differences at different operating temperatures through the flexibility of the U-shaped sections. Alternatively, or in combination, the parts of the pressure pipe 20 and other pipes required to operate at low and high temperatures can be axially freely suspended between the two brackets 51 and 52 , as shown in FIG. 6 . Each bracket comprises a bush 53 in which one end of the pressure pipe 20 is housed such that the pressure pipe 20 is radially fixed but axially movable. O-rings made of cast iron, steel, polytetrafluoroethylene (PTFE), Viton, (FPM), copolymer (NBR), nitrile rubber, poly(dimethylsiloxane) (SI), or similar 54 or similar gasket to ensure a substantially hermetic seal between the end of the pipe and the liner. The pressures exerted on the two opposite free ends of the duct 20 balance each other. Axial variations in the length of the duct 20 are absorbed by its axially freely overhanging duct ends.

Gaskets in hydraulic systems are made from cast iron, steel, polytetrafluoroethylene (PTFE), Viton, (FPM), copolymer (NBR), nitrile rubber, poly(dimethylsiloxane) (SI), mixtures thereof or Similar materials are selected to ensure a substantially hermetic seal between components of the hydraulic system. A specific gasket is described below with reference to FIG. 9 .

Figure 4 shows another preferred embodiment of the hydraulic system. This embodiment is basically the same as that shown in FIG. 3 , however, the high-pressure fuel pump 33 also acts as a low-pressure pump for circulating HFO during engine stop. Here, a clutch 56 controlled by the central control unit is provided between the gear 37 and the high-pressure fuel pump 33 . During engine operation, clutch 56 is engaged and high pressure fuel pump 33 is driven by crankshaft 3 . During engine stop, clutch 56 is disengaged. Another clutch 55 controlled by the central control unit is provided between the high pressure fuel pump 33 and the electric motor 38'. Clutch 55 is disengaged during engine operation and engaged during engine stop. The electric motor 38' drives the high pressure fuel pump 33 during engine stop, but at a much lower speed than it does during engine running, to provide sufficient hydraulic pressure for HFO to circulate at a pressure of 3-10 bar.

A preferred embodiment of the actuator 19 and the air spring 18 will be described in detail below with reference to FIGS. 7-11 .

The exhaust valve 11 has a valve stem 57 extending vertically upward from a valve head 58 and the upper end of the valve stem 57 supports a cylinder 59 mounted firmly on the valve stem 57 to form a pressure seal and longitudinally on a stationary piston 60 can be moved. The stationary piston 60 is part of a spring housing 61 . On the stationary piston 60 there is a spring chamber 62 connected to a pressurized air supply (not shown) which keeps the spring chamber 62 filled with pressurized air at a predetermined minimum pressure of eg 4.5 bar overpressure. Other gas pressures may also be used, for example 3-10 bar. The minimum pressure is selected based on the desired elastic characteristics of the air spring. It is possible to interconnect the spring chambers on a number of different cylinders, but preferably each spring chamber is individually shut off by a check valve 63 at the pressurized air supply. The pressurized air in spring chamber 62 exerts a continuous upward force on cylinder 59 . This upward force increases as the cylinder 59 moves downward and squeezes the air in the spring chamber 62, the check valve 63 preventing the air in the spring chamber 62 from flowing out.

Spring housing 61 defines a cavity 64 surrounding and above air spring 18 . The chamber 64 is connected to a discharge 65 such that the chamber has atmospheric pressure. Any oil leaking from the actuator 19 will enter the cavity 64 and be expelled via the drain 65 . The spring is configured to make it difficult for leaking oil to enter the spring chamber 62 because the cylinder 59 is formed as an umbrella which forces the leaking oil to flow over it and down to the bottom of the cavity 64 without risking entering the spring chamber 62 . This is very important because when the leaking oil tries to pass through or even further into the pneumatic system, the leaking oil (HFO) can collect and harden in the spring chamber or clog the pneumatic lines.

Referring to FIGS. 7 and 9 , the hydraulic actuator 19 is constructed from a cylinder 66 supported by the top of the spring housing 61 . Piston 67 is housed in the central bore of cylinder 66 . The central bore is closed at the top of the cylinder 66 and is open at the bottom of the cylinder 66 . The central bore is arranged coaxially with the bore 68 in the spring housing 61 . The upper (proximal) end of the piston 67 is housed in this central bore, while the distal end of the piston 67 acts on the top of the valve stem 57 .

A main pressure chamber 69 is defined between the cylinder 66 and the top of the piston 67 . Hydraulic fluid (HFO) is supplied to and discharged from the hydraulic actuators through openings 70 . The opening 70 leads into an intermediate pressure chamber 71 arranged below the main pressure chamber 69 and defined between the cylinder 66 and the middle part of the piston 67 . The opening 70 is alternately connected to the pressure conduit 20 and the return conduit 43 controlled by a proportional control valve, shown exemplarily in this view as an on/off valve 25', although a proportional valve could alternatively be used. An auxiliary pressure chamber 73 is defined between the increased diameter portion 74 of the piston 67 and the corresponding increased diameter portion of the central bore. Optionally, a gasket 68' can be disposed between the enlarged diameter portion 74 and the cylinder 66 to reduce the amount of leakage oil entering the cavity 64. During the first part of the opening stroke of the hydraulic actuator 19 , the auxiliary pressure chamber 73 is supplied with high pressure HFO from the intermediate chamber 71 via the axial channel 75 formed by the notch 75 of the piston 67 . During the opening stroke, the axial channel 75 is closed by a control flange 76 on the cylinder 67 at a predetermined intermediate position. At the same time, the opening 77 connects the auxiliary pressure chamber 73 to the return duct 43 , since the upper edge of the enlarged diameter portion 74 is now below the upper edge of the opening 77 . Thus, the increased diameter portion 74 helps overcome the greater force exerted on the valve head 58 by the pressure in the combustion chamber 15 during the first part of the opening stroke of the hydraulic actuator 19 . In a predetermined intermediate position of the piston 67 , the supply of high-pressure fluid to the auxiliary pressure chamber 73 is interrupted and it is emptied via the opening 77 . The pressure in the combustion chamber 15 is now reduced and the action of the enlarged diameter portion 74 is no longer necessary.

Figure 8 shows a typical opening profile of an exhaust valve. In phase I, the initial segment of the opening movement requires a greater force from the hydraulic actuator 19 to overcome the pressure in the combustion chamber 15 and to accelerate the heavier exhaust valve 11 . During this phase the hydraulic actuator 19 has to provide maximum force. However, hydraulic shock waves due to rapid opening of the control valve 25 or 25' should be avoided. In phase II, the exhaust valve 11 reaches a fully open position, and during this phase, the exhaust valve 11 should gradually slow down to a stop, preferably without abutting other objects against each other. In phase III, the return movement of the exhaust valve 11 should start gently, hydraulic pressure waves due to sharp opening and closing of the control valve 25 or 25' should be avoided. Gradual and precise landing of the valve head 58 on the valve seat is most critical at the end of stage IV, as the metal objects will be against each other. It is therefore crucial that the exhaust valve 11 and piston 67 are gradually slowed down so as to minimize the enormous acceleration forces and avoid the valve head impinging on the valve seat. A suitable opening curve of the exhaust valve 11 can be obtained in various ways according to the invention. One way is through the use of simple hydraulic actuators for the exhaust valves, such as hydraulic cylinders (not shown) in combination with suitable control of the proportional control valve such that the substantially exclusive degree of opening of the proportional control valve ensures A suitable force and resistance exerted by the actuator on the exhaust valve is used to obtain a suitable opening curve. Another approach is by using the hydraulic actuators and valve springs described herein, which have inherent characteristics such that a suitable opening curve for the exhaust valve can be obtained by an on/off control valve. Actuators with inherent characteristics can also be combined with proportional valves.

When the exhaust valve is about to open and the proportional control valve 25 supplies high pressure fluid to the opening 70 and pressurizes the main pressurization chamber, the intermediate pressurization chamber and the auxiliary pressurization chamber. High pressure hydraulic fluid in the primary and secondary pressurization chambers causes the piston to be pressed down.

The piston 67 (first piston part) is provided with a piston cap 78 (second piston part). The upper (proximal end) of the piston 67 slidably engages a piston cap 78 forming a compensation chamber 79 between the piston 67 and the piston cap 78 . According to a preferred embodiment, a piston cap 78 fits over the top of the piston 67 . However, a piston cap 78 may also be provided to fit within the top of the piston 67 (not shown). Spring 80 urges piston 67 and piston cap 78 away from each other, thereby enlarging compensation chamber 79 . The first flow path is arranged between the compensation chamber 79 and the main pressure chamber 69 . The first flow path includes a valve member 81 that fits within a receiving hole in the top of the piston cap 78 . The spring 80 urges the valve member 81 upwardly towards the piston cap 78 . According to another embodiment (not shown), a separate spring may be provided for urging the piston cap 78 and the valve member 81 upwards. This allows the force applied to either element to be adjusted independently of the other.

The valve member 81 is provided with an axial bore 82 and two radial bores 83 and 84 which connect the compensation chamber 79 and the main pressure chamber 69 unless the valve member 81 is in its upper position within the housing bore. In this upper position ( FIGS. 9 and 12 ), the opening of the hole 84 is prevented by the wall of the receiving hole, so that the first flow path is closed. The first flow path is used to allow excess hydraulic fluid to escape from the compensation chamber 79 when the piston 67 is in its upper position, and the piston cap 78 is set further than required due to the excess amount of hydraulic fluid in the compensation chamber 79. Near the top of the main pressure chamber 69. In this situation ( FIGS. 10 and 11 ), the valve member 81 abuts against the end surface of the cylinder 66 and the valve member 81 moves downward relative to the piston cap 78 thereby opening the first flow path so that the compensation chamber 79 can be emptied. , until the valve head 58 rests on the valve seat. The first flow path is thus only open when the piston member is within a small predetermined axial extent at the upper (proximal) end of the cylinder 66 .

The second flow path is located between the compensation chamber 79 and the intermediate pressure chamber 71 . According to a preferred embodiment, the second flow path is formed by the annular gap 85 between the piston 67 and the piston cap 78 . Due to the narrowness of the annular gap 85, the second flow path has a relatively high flow resistance. The second flow path allows the compensation chamber 79 to be filled under the action of the spring 80 . A suitable filling law of the compensation chamber is obtained by choosing suitable characteristics for the force of the spring 80 and the resistance of the flow path 85 .

A gas exchange duct 86 with a high flow restriction is arranged at the top of the cylinder 66 and connects the top of the main pressure chamber 69 formed by the damping chamber 87 to the return duct 43 .

Piston cap 78 has an axially tapered outer periphery with a diameter that gradually increases towards the top of the piston. The tapered portion cooperates with an inwardly projecting annular flange 88 extending from the central bore just above where the opening 70 opens into the central bore. Together with the annular flange 88, the tapered portion forms a narrow annular gap 89 whose size varies with the position of the piston. The hydraulic fluid has to be compressed through this annular gap 89 in order to flow from the intermediate pressure chamber 71 to the main pressure chamber 69 . This creates a pressure drop between the intermediate pressure chamber 71 and the main pressure chamber 69 . This pressure drop rises as the size of the annular gap 89 becomes smaller and gradually increases as the flow rate increases, effectively preventing the piston 67 from reaching high speeds. The tapered portion is dimensioned such that the annular gap 89 is smaller towards the end of the opening stroke. Towards the end of said stroke, the velocity of the piston 67 is thus effectively limited, even though the supply pressure of the hydraulic fluid is relatively high. The tapered portion is shown in Figures 9-11 as having a slightly outwardly curved profile, but other profiles such as a frusto-conical, slightly inwardly curved profile, a combination of both, or any desired predetermined profile are also possible of. This profile can be determined by test methods, computer simulation methods or analytical methods that indicate how much flow restriction should be at each position of travel in order to have optimal dynamic characteristics of the valve actuator. The tapered portion can then be configured accordingly.

The downward force of the actuator 19 and the upward force of the air spring 18 are balanced at the end of the outward stroke, ie the piston 67 and the exhaust valve 11 will stop by themselves, as shown in stage II of FIG. 8 . Neither shutting off the supply of high-pressure HFO nor stroke limiters to stop the piston and exhaust valves are required. Since the supply of HFO does not need to be cut off abruptly, there is no hydraulic shock wave which would otherwise stress the entire hydraulic system. There are no travel limiters, resulting in lower mechanical loads and shocks.

The pressure of the HFO supplied to the hydraulic actuator 19 and the pressure of the air supplied to the air spring 18 are controlled to ensure that the exhaust valve 11 reaches the proper open position. The actuator 19 and air spring 18 are dimensioned such that they readily balance opposing forces in the open position.

As the piston 67 approaches the fully open position, the flow path between the flange 88 and the tapered portion of the piston cap 78 narrows. The narrow gap has a damping effect on the movement of the piston 67 so that the piston stops in the open position with little or no shock and subsequent vibration.

The piston 67 returns to the retracted position under the action of the air spring 18 . The hydraulic actuator 19 is provided with end-of-stroke damping in the form of a damping chamber 87 at the top of the cylinder 66 (at the proximal end). The top of the piston cap 78 is sized to fit with a slight clearance in the damping chamber 87 and when the top of the second piston part 78 is inserted into the damping chamber, on the return stroke, hydraulic fluid is forced through the gap formed by the annular gap 90 A small gap flows out of the damping chamber 87, absorbing most of the kinetic energy of the piston 67 and exhaust valve 11, and the valve head 58 gently seats on the valve seat.

The flow resistance of the flow path between the opening 70 and the main pressure chamber 69 is adjusted by varying the design of the tapered portion according to the pressure required by the main pressure chamber 69 at various positions of the piston 67 . The hydraulic actuator 19 is thus able to cooperate well with high pressure sources of varying pressure. Relatively low supply pressure will result in low valve acceleration. Accordingly, the electronic control unit 26 continuously varies the timing and length of valve opening to compensate for pressure variations in the high pressure hydraulic fluid supply. When the supply pressure is low, the electronic control unit 26 will instruct the proportional control valve 25 to open earlier and remain open for a longer period of time to ensure that the exhaust valve is open long enough to properly vent the combustion chamber, and when And vice versa when the supply pressure is high.

The cylinder 66 includes a vent and recirculation conduit 86 whereby warm hydraulic fluid can circulate through the actuator and back into the return conduit 43 . This helps to keep the valve at operating temperature when the engine is not running, and it also efficiently removes air.

Operation of hydraulic valves

In the closed position of the exhaust valve 11 , the position of the piston cap 78 is such that its top is located inside the damping chamber 87 and within the allowable position range of the valve member 81 . Figure 10 shows the highest possible position of the piston cap 87, with the first flow path open, and Figure 12 shows the lowest possible position of the piston cap 78, with the valve member 81 closed. In the range of positions, there is always a narrow annular gap 90 between the top of the piston cap 78 and the wall of the damping chamber 87 .

The exhaust valve 11 supplies the opening 70 ( FIG. 10 ) with high pressure medium ( HFO or fuel oil) to open. The hydraulic medium thus flows through the annular gap 89 and the annular gap 90 into the main pressure chamber 69 and the damping chamber 87 and builds up a pressure which presses the piston 67 downwards. The hydraulic fluid flowing out of the opening 70 also flows into the intermediate chamber 71 and into the auxiliary pressure chamber 73 via the axial channel 75 . Thus, the pressure acting on the enlarged diameter portion 74 increases the force that urges the piston 67 downward.

When the resultant force on the piston 67 exceeds the reaction force of the air spring 18 and the pressure in the combustion chamber, the exhaust valve 11 begins to open. At the beginning of the opening movement, the restricted flow through the annular gap 90 into the damping chamber 87 creates a low pressure in the damping chamber 87, thereby ensuring a smooth beginning of the opening movement, without sharp accelerations and without There is a hydraulic shock wave, see phase I in Figure 8.

When the exhaust valve 11 is partially opened, the pressure in the combustion chamber 15 and the force required to complete the opening of the exhaust valve 11 drops significantly. At this stage, the downward force on the piston 67 is reduced by shutting off the flow of hydraulic fluid to the auxiliary pressure chamber 73 by the control flange 76 and at the same time connecting the auxiliary pressure chamber 73 to the return line 43 via the opening 77 , to allow the hydraulic fluid from the return line 43 to be supplied to the auxiliary pressure chamber 73 during the rest of the opening stroke, allowing further expansion of the auxiliary pressure chamber 73, thereby avoiding the occurrence of air in the axial passage 75 and in the auxiliary pressure chamber 73 hole.

As the degree of opening of the exhaust valve 11 increases, the flow area of the gap 89 decreases. Thus, the pressures in the main pressure chamber 69 and the auxiliary pressure chamber 79 gradually decrease. At the same time the pressure in the air spring 18 gradually increases, whereby the speed of the exhaust valve 11 decreases steadily until a good balance is established between the forces exerted by the hydraulic and pneumatic medium. Due to the gradual change of the relative fluid pressure, the exhaust valve 11 and the piston decelerate smoothly to a complete stop without any hydraulic shock wave and mechanical contact, see FIG. 8 stage II. Any oscillating movements of the exhaust valve 11 in the vicinity of the fully open position are reduced by the damping effect of the sharply reduced flow area of the gap 89 .

During the opening phase of the exhaust valve 11 , the valve member 81 is closed against the underside of the piston cap 78 due to the action of the spring 80 . The quantity of hydraulic fluid flowing into the compensation chamber 79 ensures the predefined position of the piston cap 78 . The pressure difference between the intermediate pressure chamber 71 and the main pressure chamber 69 and the force of the spring 80 force the piston cap 78 upwards so that a small amount of hydraulic fluid is sucked between the compensation chamber 79 via the annular gap 85 between the piston cap and the piston . In the fully open position of the exhaust valve 11 the pressures of the main pressure chamber 69 and the intermediate pressure chamber 71 are equal and only the spring 80 urges the piston cap 78 upwards. During the opening and fully opening phases of the exhaust valve 11 , the refilling of the compensation chamber 79 causes the piston cap 78 to move slowly upwards relative to the piston 67 .

The exhaust valve 11 closes again when the proportional control valve 25 changes position and connects the opening 70 to the return line 43 . The thrust of the air spring 18 causes hydraulic fluid to pass from the main pressure chamber 69 via the annular gap 89 into the return line 43 . The small flow area in the annular gap 89 ensures a soft start of the return stroke, which has a steadily increasing velocity during the upward movement of the piston 67, which is governed by the steady increase in the flow area of the annular gap 89, see FIG. 8 stage III . Since the pressure in the main pressure chamber 69 is higher than in the intermediate pressure chamber 71 , the discharge via the annular gap 85 will cause the compensating chamber 79 to contract somewhat. The hydraulic fluid in the auxiliary pressure chamber 73 is evacuated via the opening 77 and, when the opening 77 is blocked by the enlarged diameter portion 74 , via the axial channel 75 , the intermediate chamber 71 , the opening 70 and the return duct 43 .

During the final phase of the closing movement, the piston cap 78 is inserted into the damping chamber 87 so that the annular gap 90 formed considerably reduces the available flow area for hydraulic fluid in the damping chamber. The hydraulic fluid in the damping chamber 87 flows out of the damping chamber via the annular gap 90 so that it acts as a braking force on the piston 67 through a corresponding increase in pressure in the compensating chamber 79 to decelerate it, cf. FIG. 8 stage IV. An increase in pressure in the compensation chamber 79 will cause some of the fluid therein to escape through the annular gap 85 . The speed at which the valve head 58 lands on the valve seat before the exhaust valve 11 closes is thus largely determined by the flow area of the annular gap 90 . The ventilation duct 86 and the annular gap 85 more or less facilitate the flow of fluid from the damping chamber 87 .

If the compensation chamber 79 had been fully enlarged during the opening phase of the exhaust valve 11 , the piston cap 78 would occupy a slightly higher position than it was inserted into the damping chamber 87 . Thus, the valve member 81 will abut against the end of the cylinder 66 (bottom of the damping chamber) and open the first flow path to empty the compensating chamber 79 (Fig. 11) so that the piston cap 78 can assume the correct position (Fig. 10).

If the compensation chamber 79 is fully retracted during the return stroke of the exhaust valve 11, the piston cap 78 will occupy a slightly lower position than it is inserted into the damping chamber 87, and the valve member 81 will not abut the end of the cylinder (Figure 12). It is not until the next opening stage that the spring 80 will press the piston cap 78 upwards, so that the compensating chamber 79 will receive the amount of hydraulic fluid flowing out via the annular gap 85 until the valve member 81 abuts against the end of the cylinder 66 ( FIG. 13 ), and Make sure that the piston cap 78 occupies a generally centered position within its axial extent.

Operation of the piston cap 78 in combination with the compensation chamber 79 enables the hydraulic actuator 19 to automatically compensate for dimensional changes due to operation at different temperatures, trimming, ie lapping of valve seats, and manufacturing tolerances. Thus, the valve head 58 will always land softly and precisely on the valve seat.

According to one embodiment of the invention, the hydraulic actuator 19 can also be realized without a compensation chamber as shown in FIG. 7 . This embodiment can be used in engines where compensation for dimensional changes is not too important, for example when the typical hydraulic fluid used as hydraulic fluid operates at 30-60°C.

Although the invention has been described in detail for purposes of illustration, it is to be understood that such detail is for the purpose of explanation only and variations will be apparent to those skilled in the art without departing from the scope of the invention.

Claims (24)

1. A crosshead type large-scale two-stroke diesel engine (1), comprising: crankcase bracket (4), which supports the crankshaft (3) and the cylinder bracket (5) mounted on said crankcase bracket (4); A plurality of cylinders (6), which are supported by the cylinder bracket (5), each cylinder is provided with at least one fuel injector (23) and at least one exhaust valve (11); a hydraulic actuator (19) associated with each said exhaust valve (11); a common fuel rail (40) having one or more accumulators (42) connected thereto; a high pressure fuel pump (33) supplying fuel at high pressure to said common fuel rail (40), each of said fuel injectors (23) operates on fuel from said common fuel rail (40); and proportional valves associated with each cylinder (6), The proportional valves thereby control the flow of fuel from the common fuel rail (40) to the respective fuel injectors (23). 2. The engine of claim 1, wherein fuel injection timing, volume of injected fuel, and shape of an injection pattern are controlled by said proportional valve. 3. An engine as claimed in claim 1 or 2, wherein said hydraulic actuator (19) is operatively connected to said common fuel rail (40). 4. An engine as claimed in claim 3, wherein for each of said cylinders (6) there is associated therewith a control valve which controls fuel from said common fuel rail (40) to a corresponding hydraulic actuator (19) FLOW. 5. The engine of claim 4, wherein said control valve is an on/off valve. 6. The engine of claim 4, wherein said control valve is a proportional valve. 7. An engine as claimed in claim 5 or 6, wherein said proportional valve and said control valve are combined into one integral valve with a single spool valve. 8. The engine according to claim 7, wherein said integral valve comprises a valve body and an electromagnet (44) for controlling said spool valve, whereby said electromagnet (44) is thermally insulated from said valve body. 9. The engine according to claim 8, wherein a layer of heat insulating material (45) is arranged between the electromagnet (44) and the valve body. 10. A crosshead type large two-stroke diesel engine (1), comprising: crankcase bracket (4), which supports the crankshaft (3) and the cylinder bracket (5) mounted on said crankcase bracket; A plurality of cylinders (6), which are supported by the cylinder bracket (5), each cylinder is provided with at least one fuel injector (23) and at least one exhaust valve (11); a hydraulic actuator (19) associated with each said exhaust valve (11); a common fuel rail (40) having one or more accumulators (42) connected thereto; a high pressure fuel pump (33) supplying fuel at high pressure to said common fuel rail (40), each of said fuel injectors (23) operates on fuel from said common fuel rail (40); and proportional control valves associated with each cylinder (6), The proportional control valves thereby control the flow of fuel from the common fuel rail (40) to the respective hydraulic actuators (19). 11. The engine according to claim 10, wherein the opening and closing timing of the exhaust valve (11) and the opening degree of the exhaust valve (11) are controlled by corresponding proportional control valves. 12. An engine as claimed in claim 10 or 11, wherein the opening and closing curve of the exhaust valve (11) is determined by the characteristics of the hydraulic actuator (19). 13. An engine as claimed in claim 10 or 11, wherein the opening and closing curve of the exhaust valve (11) is determined by the proportional control valve. 14. An engine as claimed in claim 10, wherein said proportional control valve operates as an on/off valve when controlling said hydraulic actuator (19). 15. The engine of claim 10, wherein the flow of fuel from said common fuel rail (40) to said fuel injectors (23) is controlled by an on/off valve. 16. The engine of claim 10, wherein the flow of fuel from said common fuel rail (40) to said fuel injectors (23) is controlled by a proportional valve. 17. The engine of claim 15, wherein said proportional control valve and said on/off valve for controlling the flow of fuel to said fuel injector (23) are combined into an integral valve with a single spool valve . 18. Use of a proportional valve to control the flow of fuel from a common fuel rail (40) of a crosshead type large two-stroke diesel engine (1) to a fuel injector (23) and/or a fuel running component (19). 19. Proportional valve use as claimed in claim 18, wherein the fuel operating component is a hydraulic actuator (19). 20. The use of a proportional valve as claimed in claim 19, wherein said proportional valve operates as an on/off valve. 21. An electrically controlled valve for controlling the flow of fuel from a common fuel rail (40) of a crosshead type large two-stroke diesel engine (1) to one or more fuel-running or fuel-consuming engine components (19, 23) Flow, the electric control valve includes a valve body and an electromagnet (44), characterized in that the electromagnet (44) is insulated from the valve body. 22. The electrically controlled valve of claim 21, wherein said electrically controlled valve is a proportional valve. 23. An electrically controlled valve as claimed in claim 21 or 22, wherein said electrically controlled valve has at least three positions: a central position which connects the two hydraulic fuel running or fuel consuming engine components (19, 23) to the return flow Conduit (43); a first non-central location wherein a first of said two hydraulic fuel operating or fuel consuming engine components is connected to a high pressure fuel source (41), said two hydraulic fuel operating or fuel consuming engine components a second of the components (19, 23) is connected to said return conduit (43); and a second non-central location wherein said second of said two hydraulic fuel running or fuel consuming engine components is connected to The high pressure fuel source (41) is connected and the first of the two hydraulic fuel operating or fuel consuming engine components (19, 23) is connected to the return line (43). 24. The electric control valve according to claim 21, wherein said electromagnet (44) is thermally insulated from said valve body by a layer of insulating material (45), said insulating material (45) being a ceramic material.

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