Large diesel engine and method for determining cylinder pressure in large diesel engine

Abstract

To provide a large crosshead driven diesel engine in which cylinder pressure can be determined in a reliable manner.SOLUTION: A large diesel engine includes: at least one cylinder 2 having a combustion chamber 4 that is defined by a piston 3 reciprocable along a cylinder axis A; and a rotatable crankshaft 9. According to the present invention, in the large diesel engine, the piston is connected via a piston rod 6 to a crosshead 7 having a crosshead pin 71, and the crosshead is connected via a push rod 8 to the crankshaft 9. A hydraulic chamber 10 is provided in the crosshead pin, and the hydraulic chamber is defined by the piston rod. By means of the hydraulic chamber, the piston rod is displaceable relative to the crosshead pin in a direction of the cylinder axis A. The large diesel engine is provided with: a pressure sensor 11 capable of determining hydraulic pressure in the hydraulic chamber; and an evaluation unit 12 capable of determining cylinder pressure in the combustion chamber from the hydraulic pressure.SELECTED DRAWING: Figure 1

Description

Technical Field

[0001] The present invention relates to a large diesel engine described in the preamble of the independent claims of each category, and a method for determining cylinder pressure in a large diesel engine.

Background Art

[0002] A large diesel engine can be designed as a two-stroke engine or a four-stroke engine, for example, as a longitudinally scavenged two-stroke large diesel engine, and is frequently used as a driving engine for ships or in steady operation for power supply of a large generator for generating electric energy, for example. Here, the engine generally operates in continuous operation for a long time, and thus high requirements are placed on operational safety and availability. Therefore, a particularly long maintenance interval, low wear, and economical handling of operating materials are the main criteria for the operator. A large diesel engine usually has cylinders with an inner diameter (bore) of at least 200 mm. Today, large diesel engines with a bore of up to 960 mm or larger are used. Large diesel engines are conventionally operated with heavy fuel oil.

[0003] In view of economical and efficient operation, compliance with exhaust gas standards, and resource conservation, alternatives to fuel heavy oil are currently being sought even in large diesel engines. Here, both liquid fuels, which mean fuels introduced into the combustion chamber in a liquid state, and gaseous fuels, which mean fuels introduced into the combustion chamber in a gaseous state, are used.

[0004] Examples of liquid fuels known as alternatives to heavy oil include other heavy hydrocarbons, alcohols, particularly methanol or ethanol, which remain as residues during the refining of natural oils, gasoline, diesel, or emulsions or suspensions. For example, emulsions known as MSAR (Multiphase Superfine Atomized Residue) are known to be used as fuel. Known suspensions are those made from coal dust and water, and are also used as fuel for large engines. Gaseous fuels include natural gas such as LNG (Liquefied Natural Gas), liquid gases such as LPG (Liquefied Petroleum Gas), or ethane.

[0005] In particular, there are large diesel engines that can operate using at least two different fuels, and in which the engine operates using one fuel or the other depending on the operating conditions or environment.

[0006] An example of a large diesel engine that can operate using two different fuels is a large diesel engine designed as a dual-fuel large diesel engine. This can operate in liquid mode, where liquid fuel is introduced into the cylinders for combustion, and in gas mode, where gas is introduced into the cylinders as fuel.

[0007] Large diesel engines can operate using at least two or more different liquid or gaseous fuels, and are often operated in different operating modes depending on the fuel actually being used. In the operating mode often called diesel operation, combustion of the fuel generally occurs according to the principles of compression ignition or autoignition of the fuel. In the mode often called Otto operation, combustion is carried out by spark ignition of a flammable premixed air-fuel mixture. For example, this spark ignition can be carried out by an electric spark from, for example, a spark plug, or by the autoignition of a small amount of injected fuel, thereby causing spark ignition of another fuel. The small amount of fuel intended for autoignition is often injected into a pre-combustion chamber connected to the combustion chamber.

[0008] Furthermore, a hybrid configuration of Otto operation and diesel operation is also known.

[0009] Within the scope of this application, the term “large diesel engine” refers to an engine capable of operating in at least diesel mode. In particular, the term “large diesel engine” also includes dual-fuel large engines capable of operating in another operating mode, such as Otto mode, in addition to diesel mode.

[0010] Within the scope of this application, the terms “gas mode,” or alternatively “operation in gas mode,” mean the use of only gas or gaseous fuel as fuel for torque-generating combustion. As stated above, while it is possible and quite common to inject a small amount of self-igniting liquid fuel, such as heavy oil, to spark-ignite the premixed air-fuel mixture in gas mode, the torque-generating combustion process is still operated entirely by gas or gaseous fuel.

[0011] This process of spark ignition through the self-ignition of a small amount of liquid fuel is sometimes called pilot injection. This pilot injection is completely unrelated to the injection of liquid fuel into the combustion chamber when a large engine is operated in liquid mode. Pilot injection typically uses a different injection device than that used for injecting liquid fuel in liquid mode. Also, during pilot injection, the small amount of liquid fuel is often not injected directly into the combustion chamber, but rather into at least one pre-combustion chamber connected to the combustion chamber via a duct.

[0012] It is also known that such dual-fuel large diesel engines can be operated in a low-pressure process in gas mode, that is, by introducing gas into the cylinder in a gaseous state with a gas injection pressure of up to 50 bar, preferably up to 20 bar. For this purpose, the cylinder wall is provided with at least one gas inlet, through which gas is introduced into the cylinder in gas mode. In practice, two gas inlets are often provided, located on opposite sides in the diametrical direction with respect to the cylinder axis. The gas inlets are located at a height between the lower and upper reversal points of piston motion, so that gas can be introduced into the cylinder during the upward movement of the piston, provided that no compression occurs in the cylinder, or at least no significant compression occurs.

[0013] Typically, large diesel engines are designed to have the highest possible thermodynamic efficiency at 100% load, i.e., at full load and maximum rotational speed, by optimizing the density ratio, or in other words, the compression ratio, so that they have the best possible balance between consumption behavior and efficiency.

[0014] The compression ratio is a geometrical value that is the ratio of the first volume of the combustion chamber before the air-fuel mixture is compressed to the second volume of the remaining air-fuel mixture in the combustion chamber after it has been compressed.

[0015] Optimizing combustion behavior at 100% load results in the efficiency of large diesel engines being suboptimal at low loads, such as at low to medium pressures.

[0016] Furthermore, in large diesel engines that operate using at least two different fuels, such as dual-fuel large diesel engines, it is desirable to achieve the highest possible efficiency for each of the different fuels.

[0017] For these reasons, large diesel engines are known that can change the compression ratio to optimize efficiency for each load and / or each fuel. Such designs are also called VCR systems (VCR: Variable Compression Ratio).

[0018] Regarding large diesel engines with crosshead drive, for example, EP-A-2687707 indicates that, in order to change the compression ratio, the piston rod, which is installed in the crosshead pin of the crosshead, can be displaced relative to the crosshead pin in the direction of the cylinder axis in order to change the compression ratio. In this way, the compression ratio can be changed. For example, if the piston rod is displaced relative to the crosshead pin in the direction of the combustion chamber, the volume of the combustion chamber at maximum compression will be smaller as a result, and therefore the compression ratio will be larger.

[0019] For example, a significant parameter that is crucial in both optimally adjusting the compression ratio depending on the current load and adapting it to the fuel actually being used is cylinder pressure, which refers to the pressure inside the combustion chamber of the cylinder. For this reason, it is common to install a pressure sensor in each cylinder, and the pressure inside the combustion chamber can be measured by the pressure sensor.

[0020] A known solution is to place a pressure sensor as close as possible to the combustion chamber of the cylinder. For this purpose, a continuous bore opening into the combustion chamber is provided in the cylinder cover. This bore has threads near the entrance to the combustion chamber, so that a pressure sensor with male threads can be screwed into the bore. Here, the pressure sensor is placed as close to the combustion chamber as possible, specifically to minimize the length of the bore between the pressure sensor and the combustion chamber. This is because this region of the bore becomes a dead volume where gas or other fuels may burn uncontrollably. Also, in such a dead volume, scavenging by scavenging air during the scavenging process is insufficient, which can lead to methane slip, especially in gaseous mode.

[0021] However, because the pressure sensor is located very close to the combustion chamber, it is exposed to extremely high temperatures, negatively impacting its lifespan. This proximity also leads to combustion residue adhering to the sensor surface. Furthermore, this residue often covers not only the sensor surface but the entire bore opening. Therefore, to ensure reliable and efficient operation of large engines, the pressure sensor needs to be disassembled at regular intervals, particularly to allow for cleaning of the sensor surface. Because the pressure sensor is located very low within the bore of the cylinder cover, it is always difficult to access. A long tool is required to insert into the bore to unscrew or rescrew the pressure sensor from the bore threads. Consequently, these service operations require considerable effort. [Prior art documents] [Patent Documents]

[0022] [Patent Document 1] European Patent Application Publication No. 2687707 [Overview of the Initiative] [Problems that the invention aims to solve]

[0023] Based on this prior art, an object of the present invention is to provide a large crosshead-driven diesel engine that can determine the pressure in the combustion chamber of a cylinder in a reliable manner, thereby preventing the above-described problems from occurring. Furthermore, an object of the present invention is to propose a corresponding method for determining the cylinder pressure in a large diesel engine.

Means for Solving the Problems

[0024] The subject matter of the present invention that meets this object is characterized by the characterizing part of the independent claims of each category.

[0025] According to the present invention, there is provided a large diesel engine comprising at least one cylinder having a combustion chamber defined by a piston that is positioned to be reciprocally movable along a cylinder axis, and a rotatable crankshaft. The piston is connected to a crosshead having a crosshead pin via a piston rod, the crosshead is connected to the crankshaft via a push rod, a hydraulic chamber is provided in the crosshead pin, the hydraulic chamber is defined by the piston rod, and the piston rod can be displaced in the cylinder axis direction with respect to the crosshead pin to adjust the compression ratio using the hydraulic chamber. A pressure sensor capable of determining the hydraulic pressure in the hydraulic chamber and an evaluation unit capable of determining the cylinder pressure in the combustion chamber using the hydraulic pressure are provided.

[0026] In this way, the cylinder pressure in the combustion chamber is measured using the hydraulic pressure in the hydraulic chamber, whereby the piston rod can be displaced to adjust the compression ratio. Since the piston rod confines the hydraulic chamber, the cylinder pressure acting on the combustion chamber side of the piston is transmitted to the hydraulic chamber via the piston rod. The pressure sensor measures the hydraulic pressure acting in the hydraulic chamber, transmits this measured value to the evaluation unit, and the evaluation unit determines the cylinder pressure inside the combustion chamber from this measured value. It is not essential but possible to reduce the pressure in the hydraulic chamber to a lower pressure proportional to the hydraulic pressure in the hydraulic chamber by means of a pressure reducer, detect this reduced pressure by measurement, and determine the pressure in the combustion chamber from this measured value.

[0027] As a result, a pressure sensor located extremely close to the combustion chamber becomes unnecessary. Therefore, all of the above-mentioned problems resulting from the pressure sensor being located extremely close to the combustion chamber, for example, on the cylinder cover, can be avoided.

[0028] One possible embodiment of the present invention is that the pressure sensor is positioned on the crosshead so that it moves together with the crosshead. Thus, the pressure sensor does not move relative to the crosshead, but moves in the same way as the crosshead. The measurement taken by the pressure sensor is then transmitted to a typically stationary evaluation unit. This transmission from the pressure sensor to the evaluation unit can be done wirelessly, for example, with a transmitter on the pressure sensor and a receiver on the evaluation unit. Transmission can also be done by a flexible cable connecting the evaluation unit to a pressure sensor that is movable relative to the evaluation unit. Such a cable may be located, for example, on or inside a toggle lever. Such a toggle lever is located between a stationary engine housing and a crosshead that moves in operation and is typically provided in a large diesel engine to introduce a lubricating or cooling medium into the crosshead or piston rod. Alternatively, the toggle lever can be used to guide a cable that delivers a hydraulic medium into the hydraulic chamber to displace the piston rod.

[0029] In a preferred embodiment, the pressure sensor is located at a measurement point that is fixed relative to the engine housing. In this embodiment, the pressure sensor is therefore fixed in place and does not move during operation of the large diesel engine; that is, the pressure sensor does not move relative to the engine housing, while the crosshead moves relative to the pressure sensor.

[0030] Preferably, a hydraulic transmission device is provided that can transmit the hydraulic pressure in the hydraulic chamber to a pressure sensor. The hydraulic transmission device makes it easy to apply a pressure to a stationary pressure sensor that is equal to or proportional to the hydraulic pressure in the hydraulic chamber within the crosshead.

[0031] Particularly preferred, the hydraulic transmission device has a toggle lever, one end of which is connected to the crosshead and the other end to the engine housing. Here, it is possible to provide a separate toggle lever dedicated to the pressure transmission device, or to use another existing toggle lever for the pressure transmission device.

[0032] The toggle lever of the hydraulic transmission device preferably has a supply section that can deliver the hydraulic medium into the hydraulic chamber. Therefore, it is preferable to use the same toggle lever in the hydraulic transmission device through which the hydraulic medium is introduced into the hydraulic chamber.

[0033] A further preferred approach is for the toggle lever to have a control conduit for a relief valve to discharge the hydraulic medium from the hydraulic chamber.

[0034] Furthermore, it is preferable that the hydraulic transmission device has a pressure reducer to which hydraulic pressure in the hydraulic chamber can be applied, and the pressure reducer reduces this pressure. This has the advantage that it is not necessary to transmit high hydraulic pressure in the hydraulic chamber, and only hydraulic pressure reduced in proportion to the pressure in the hydraulic chamber is transmitted. In operating conditions, the pressure in the hydraulic chamber can actually reach up to several hundred bar (1 bar = 0.1 MPa), for example, 500 to 600 bar (50 to 60 MPa).

[0035] The pressure reducer is preferably located on the crosshead. The pressure reducer moves with the crosshead, and as a result, only the reduced pressure needs to be transmitted from the moving crosshead to a measurement point that is fixed relative to the engine housing.

[0036] It is even more preferable that the pressure reducer has a high-pressure side and a low-pressure side, with the high-pressure side being connected to the hydraulic chamber via a flow path, and that the pressure at the low-pressure side can be applied to the pressure sensor.

[0037] For example, a pressure reducer is connected to a discharge pipeline for discharging the hydraulic medium from the hydraulic chamber. Here, this pipeline connection is located upstream of a shut-off valve, and the hydraulic medium can be discharged from the hydraulic chamber through this shut-off valve.

[0038] A further preferable approach is to be able to supply the hydraulic fluid to the low-pressure side of the pressure reducer. In this way, the pressure reducer can be returned to its initial configuration.

[0039] Preferably, the large diesel engine is designed as a longitudinally scavenging two-stroke large diesel engine.

[0040] It is particularly preferable that the large diesel engine be designed as a dual-fuel large diesel engine, capable of operating in liquid mode, where liquid fuel is introduced into the combustion chamber and burned, and further capable of operating in gaseous mode, where gas is introduced into the combustion chamber and burned.

[0041] The present invention further proposes a method for determining cylinder pressure in a large diesel engine, comprising at least one cylinder having a combustion chamber limited by a piston positioned to be reciprocally movable along the cylinder axis, and a rotatable crankshaft, wherein the piston is connected via a piston rod to a crosshead having a crosshead pin, the crosshead is connected to the crankshaft via a push rod, a hydraulic chamber is provided at the crosshead pin, the hydraulic chamber is limited by the piston rod, and the hydraulic chamber is used to displace the piston rod axially relative to the crosshead pin to adjust the compression ratio. According to the present invention, the hydraulic pressure in the hydraulic chamber is determined by a pressure sensor, and the cylinder pressure in the combustion chamber is determined by an evaluation unit using the hydraulic pressure.

[0042] In the method according to the present invention, the pressure in the combustion chamber is determined using the hydraulic pressure in the hydraulic chamber. As a result, a pressure sensor located in the immediate vicinity of the combustion chamber is unnecessary. Therefore, all the aforementioned problems caused by placing a pressure sensor near the combustion chamber, for example, on the cylinder cover, can be avoided.

[0043] Additional advantageous strategies and embodiments of the present invention are derived from the dependent claims.

[0044] The present invention will be described in detail below with reference to examples and drawings. [Brief explanation of the drawing]

[0045] [Figure 1] This is a schematic cross-sectional view showing a first embodiment of a large diesel engine according to the present invention. [Figure 2] This is a schematic cross-sectional view showing a second embodiment of the large diesel engine according to the present invention. [Figure 3] This is a schematic cross-sectional view showing a third embodiment of the large diesel engine according to the present invention. [Figure 4] This is a schematic cross-sectional view showing a fourth embodiment of the large diesel engine according to the present invention. [Figure 5] This is a schematic cross-sectional view showing a fifth embodiment of the large diesel engine according to the present invention. [Modes for carrying out the invention]

[0046] The term "large diesel engine" is typically used to refer to the main drive engine of a ship, or for steady operation to power, for example, a large generator for generating electrical energy. Typically, the cylinders of a large diesel engine have a bore of at least 200 mm. The term "longitudinal scavenging" refers to the introduction of scavenging or intake air into the cylinder in the lower end region. Combustion residue, i.e., exhaust gases, is discharged at the upper end of the cylinder.

[0047] In the following description of the present invention, we are referring to a large diesel engine designed as a dual-fuel large diesel engine, that is, an engine capable of operating using two different fuels. Specifically, a dual-fuel large diesel engine can operate in liquid mode, in which only liquid fuel is injected into the combustion chamber of the cylinder. Typically, liquid fuel, such as heavy oil or diesel oil, is injected directly into the combustion chamber at the appropriate moment, where it autoignites according to the diesel principle of autoignition. A large diesel engine can also operate in gaseous mode, in which a gas that acts as fuel, such as natural gas such as LNG (Liquefied Natural Gas), or LPG (Liquefied Petroleum Gas), or ethane, is burned in the combustion chamber via spark ignition in the form of a premixed air-fuel mixture.

[0048] As already stated above, within the scope of this application, the terms “gas mode” or alternatively “operation in gas mode” should be understood as meaning that in this gas mode, a large diesel engine is operated using only gas or gaseous fuel, and optionally a small amount of self-igniting fuel, such as heavy oil or diesel oil, is introduced into the combustion chamber or pre-combustion chamber or multiple pre-combustion chambers solely for spark ignition of the air-gas mixture (pilot injection).

[0049] Specifically, in gas mode, the large diesel engine operates according to a low-pressure method, that is, gas is introduced into the cylinder in a gaseous state, with an injection pressure of up to 50 bar, preferably up to 20 bar. The air-gas mixture is spark-ignited in the combustion chamber according to Otto's principle. Preferably, this spark ignition is caused by introducing a small amount of self-igniting liquid fuel (e.g., diesel oil or heavy oil) into the combustion chamber or pre-combustion chamber or multiple pre-combustion chambers at the appropriate moment, which self-ignites and causes spark ignition of the air-fuel mixture in the combustion chamber.

[0050] The embodiments described herein refer to a large diesel engine designed as a crosshead-driven, longitudinally scavenging, twin-fuel, two-stroke, large diesel engine. Naturally, the present invention is not limited to twin-fuel, large diesel engines, but relates to all types of large diesel engines, i.e., large engines capable of operating at least in diesel mode.

[0051] Figure 1 is a simplified schematic diagram of a first embodiment of a large diesel engine according to the present invention, the entire structure of which is designated by reference numeral 1. In Figure 1, only one cylinder 2 is depicted among the usual multiple cylinders 2 of the large diesel engine 1.

[0052] Inside cylinder 2, piston 3 is positioned in a manner known to itself, so as to be able to reciprocate between a lower reversal point and an upper reversal point along the cylinder axis A.

[0053] The piston 3 has an upper surface 31 that works in cooperation with the cylinder cover 21 to define the combustion chamber 4 in which the combustion process takes place.

[0054] As is known in crosshead drive systems, the piston 3 is connected to a crosshead 7 via a piston rod 6, which in turn is connected to a crankshaft 9 via a push rod 8. As a result, the motion of the piston 3 is transmitted to the crankshaft 9 via the piston rod 6, the crosshead 7, and the push rod 8, causing it to rotate. The rotation of the crankshaft is indicated by an arrow denoted by the reference numeral R. The crosshead 7 is designed in a manner known to itself to convert the linear up-and-down motion of the piston 3 and the piston rod 6 into the rotational motion of the push rod 8, which is supported so as to pivot about a crosshead pin 71 of the crosshead 7.

[0055] The structure and individual components of such a large diesel engine 1, such as an injection system for liquid mode (not shown), a gas supply system for gas mode (not shown), a gas exchange system for supplying scavenging air or intake air, a gas exhaust system (not shown), or a turbocharger system (not shown), are well known to experts in both two-stroke and four-stroke engine designs and therefore require no further explanation here.

[0056] Since this is sufficient for understanding the present invention, only one of these components, the outlet valve 5, is shown in Figure 1. The gas supply system for the gas mode typically has two gas inlets (not shown) through which the gas that acts as fuel in the gas mode is introduced into the cylinder 2. The two gas inlets are preferably located in the wall of the cylinder 2, and particularly preferably located on opposite sides in the diametrical direction from each other, and positioned approximately midway between the upper and lower reversal points with respect to the axial direction defined by the cylinder axis A.

[0057] Furthermore, an engine control system 100 is provided to control the functions and operation of the large diesel engine 1. In modern large diesel engines 1, the engine control system 100 is an electronic system that can adjust, control, or adjust all engine and cylinder functions, particularly injection (start and end of injection) in both gaseous and liquid modes, as well as the operation of the outlet valve 5.

[0058] In the embodiment of the longitudinal scavenging two-stroke large diesel engine described herein, a scavenging slit (not shown) is typically provided in the lower region of each cylinder 2 or cylinder liner, which opens and closes periodically by the movement of the piston 3 in the cylinder 2, thereby allowing scavenging air supplied from the turbocharger to the intake receiver (not shown) under intake pressure to flow into the cylinder 2 through the scavenging slit as long as the scavenging slit is open. The cylinder head or cylinder cover 21 is provided with an outlet valve 5 located approximately in the center, which allows combustion gases to be discharged from the cylinder 2 into a gas exhaust system (not shown) after the combustion process. The gas exhaust system directs at least a portion of the combustion gases to the turbine (not shown) of the turbocharger, and the turbocharger's compressor supplies intake air to the intake receiver under intake air pressure.

[0059] One or more fuel injection nozzles (not shown) are provided to introduce liquid fuel into the combustion chamber 4 of cylinder 2 in liquid mode, and these are located, for example, near the outlet valve 5 of the cylinder cover 21. For example, heavy oil or diesel oil can be burned as the liquid fuel in liquid mode.

[0060] For gas supply or introduction in gas mode, a gas supply system known by itself is provided, the gas supply system having a gas inlet (not shown). The gas inlet is preferably designed as a gas inlet valve equipped with a gas inlet nozzle.

[0061] The piston rod 6 is positioned such that its lower end in the figure is inside the crosshead pin 71. Furthermore, a hydraulic chamber 10 for adjusting the compression ratio is provided inside the crosshead pin 71. The compression ratio is a geometric value which is the ratio of the first volume of the combustion chamber 4 before the scavenging air or air-fuel mixture is compressed to the remaining second volume of the combustion chamber 4 after the scavenging air or air-fuel mixture is compressed. The first volume is the volume of the combustion chamber 4 immediately after the outlet valve 5 is closed, i.e., at the moment when compression begins as the piston 3 moves upward. The second volume is the volume of the combustion chamber 4 at the time of maximum compression of the scavenging air or air-fuel mixture. This is essentially the volume of the combustion chamber 4 at the start of the combustion process.

[0062] Devices for changing the compression ratio are known from existing technology, such as EP-A-2687707, so a detailed explanation is not necessary here.

[0063] In the large diesel engine 1 according to the present invention, the hydraulic chamber 10 is designed to adjust the compression ratio so that the entire piston rod 6, located within the crosshead pin 71, can be displaced relative to the crosshead pin 71 in the direction of the cylinder axis A. For this purpose, the piston rod 6 is designed and positioned to limit the hydraulic chamber 10. The piston rod 6 can be displaced relative to the crosshead pin 71 by introducing a hydraulic medium, such as hydraulic fluid, into the hydraulic chamber 10, or by discharging the hydraulic medium from the hydraulic chamber 10. As shown in Figure 1, the piston rod 6 and piston 3 can thus be displaced upward and downward relative to the crosshead pin 71. When the piston rod 6 is displaced upward in the figure by introducing a hydraulic medium into the hydraulic chamber 10, the second volume, i.e., the volume at maximum compression, decreases, and the compression ratio increases. When the piston rod 6 is displaced downward in the figure by discharging the hydraulic medium from the hydraulic chamber 10, the second volume, i.e., the volume at maximum compression, increases, and the compression ratio decreases. Therefore, the compression ratio can be continuously adjusted between a minimum and a maximum value using the hydraulic chamber 10.

[0064] This allows the compression ratio of the large diesel engine 1 to be optimally adjusted for each load under which it is operated, resulting in the most efficient combustion process possible for each load. Furthermore, the dual-fuel large diesel engine can also be further optimized for each fuel in each case. For example, the large diesel engine 1 can operate at a lower compression ratio in gas mode than in liquid mode.

[0065] Furthermore, the compression ratio can be optimally adjusted according to other operating conditions or parameters, such as the temperature of the scavenging air (supply air), the methane number of the gas acting as fuel in gaseous mode, or other operating parameters.

[0066] According to the present invention, a pressure sensor 11 capable of determining the hydraulic pressure within the hydraulic chamber 10 is provided, and an evaluation unit 12 capable of determining the cylinder pressure in the combustion chamber 4 using the hydraulic pressure.

[0067] The upper surface 31 defines the boundary of the combustion chamber 4. Through the piston 3 and piston rod 6, the cylinder pressure in the combustion chamber is transmitted to the hydraulic medium in the hydraulic chamber 10, changing the hydraulic pressure in the hydraulic chamber 10. Therefore, the cylinder pressure in the combustion chamber 4 can be determined from the hydraulic pressure in the hydraulic chamber 10.

[0068] According to the first embodiment depicted in Figure 1, the pressure sensor 11 is located at a measurement point that is fixed relative to the engine housing 200. In Figure 1, the engine housing 200 is indicated by line 200. The engine housing 200 has a base on which, for example, a sliding track for the crosshead 7 is located, and is stationary in the sense that it does not move relative to the space in which the large diesel engine 1 is located, apart from vibration, during the operation of the large diesel engine 1. In contrast, the drive components, namely the piston 3, piston rod 6, crosshead 7, push rod 8, and crankshaft 9, move relative to the engine housing 200 during the operation of the large diesel engine 1, i.e., are not stationary. The terms "stationary" and "not stationary" should be understood in the following senses: Stationary components do not move relative to the engine housing 200 during the operation of the large diesel engine 1. Not stationary components move relative to the engine housing 200 during the operation of the large diesel engine 1. Therefore, for example, piston 3 is a non-stationary component because it undergoes a stroking motion.

[0069] A hydraulic transmission device 13 is provided to transmit the hydraulic pressure in the hydraulic chamber 10 to a pressure sensor 11 that is positioned stationary. The pressure transmission device 13 has a hydraulic pipeline H, one end of which is connected to the hydraulic chamber 10, and the other end which applies pressure to the pressure sensor 11. The hydraulic pipeline H is depicted as a dashed line in Figure 1.

[0070] The hydraulic transmission device 13 further has a toggle lever 14, which is mounted in a manner known to itself to a fixed point 141 that is fixed relative to the engine housing 200, with one end on a non-stationary crosshead 7 and the other end on a fixed point 141. Therefore, the toggle lever 14, also known as a pivot lever or rocker arm, has one end that can follow the movement of the crosshead 7 and the other end that is attached to a fixed point 141 that is stationary relative to the engine housing 200. The hydraulic line H can be located on or within the toggle lever 14.

[0071] Preferably, the hydraulic transmission device 13 has a pressure reducer 15 for reducing the hydraulic pressure in the hydraulic chamber 10 to a lower pressure applied to the pressure sensor 11. The pressure reducer 15 is preferably located on the crosshead 7 and fixed, and as a result moves with the crosshead 7. The pressure reducer 15 is located in the hydraulic pipeline H such that the high-pressure side is flow-connected to the hydraulic chamber 10 and applied at the pressure in the hydraulic chamber 10. On the low-pressure side, the pressure reducer 15 is connected to the pressure sensor 11, and the reduced pressure is applied to the pressure sensor 11. The advantage of the pressure reducer 15 is that the hydraulic pressure acting in the hydraulic chamber 10 does not need to be fully transmitted from the crosshead 7 to this fixed measurement point where the pressure sensor 11 is located; only the reduced pressure needs to be transmitted. This pressure reduction via the pressure reducer 15 is taken into consideration in the evaluation unit 12 when determining the cylinder pressure.

[0072] Preferably, the pressure reducer 15 is positioned horizontally as depicted in Figure 1, so that the acceleration effect of the vertically moving crosshead 7 does not affect the pressure transmission within the pressure reducer 15. The pressure reducer 15 has, for example, a pressure piston, which has two pressure-pressurized planes, the two planes of different sizes, thereby enabling pressure reduction. The pressure piston is positioned to be displaced horizontally, i.e., perpendicular to the direction of movement of the crosshead 7 or piston rod 6.

[0073] The pressure sensor 11 is signal-connected to the evaluation unit 12. This signal connection is realized, for example, by a first signal line S1. In Figure 1, the signal connection is depicted by a dashed line. The evaluation unit 12 is signal-connected to the engine control system 100. This signal connection is realized, for example, by a second signal line S2. The cylinder pressure determined by the evaluation unit 12 is transmitted to the engine control system 100 via the second signal line S2, so that the engine control system 100 can determine the cylinder pressure in the combustion chamber 4 of the cylinder 2.

[0074] The evaluation unit 12 stores a calculation method, which, using this method, determines the cylinder pressure in the combustion chamber 4 of the cylinder 2 from the pressure or signal measured by the pressure sensor 11. The calculation method takes into account the pressure reduction by the pressure reducer 15. To ensure the most accurate determination of the cylinder pressure from the measurements transmitted by the pressure sensor 11, the calculation method may also take into account other values, such as dynamic forces, including centrifugal force or centrifugal power resulting from the motion of the piston 3. Such dynamic forces resulting from the motion of the piston 3 and piston rod 6 may depend, for example, on the rotational speed at which the large diesel engine 1 is actually operated.

[0075] A first reservoir 16 for the hydraulic fluid is provided to compensate for possible leakage of the hydraulic fluid, for example, leakage that may occur in the toggle lever 14 or the pressure reducer 15. The first reservoir 16 is connected to the hydraulic pipeline H via a replenishment line 17, which opens into the hydraulic pipeline H at a point located between the fixed point 141 and the pressure sensor 11, i.e., the opening of the replenishment line 17 is located at the stationary part of the hydraulic pipeline H. A first check valve 171 is provided in the replenishment line 17 between its opening into the hydraulic pipeline H and the first reservoir 16, and the check valve 171 is configured to allow the hydraulic fluid to flow from the first reservoir 16 into the hydraulic pipeline H, but not from the hydraulic pipeline H into the first reservoir 16. Furthermore, the hydraulic medium introduced from the replenishment pipeline 17 into the hydraulic pipeline H can be used, for example, to return the pressure reducer 15 to its initial position by moving the pressure piston provided in the pressure reducer 15 to the right side in the figure, i.e., to the high-pressure side. In addition, a second throttling device 175 is provided in the replenishment pipeline 17 between the first check valve 171 and the first reservoir 16.

[0076] Replenishing the hydraulic fluid from the first reservoir 16 into the hydraulic pipeline H or resetting the pressure reducer 15 is preferably performed when the piston 3 is at or near its lower reversal point in motion. At this stage of the piston 3's periodic motion, the pressure in the combustion chamber 4 acting on the upper surface of the piston 31 is minimal, and therefore the hydraulic pressure in the hydraulic chamber 10 is also minimal. The pressure of the hydraulic fluid present in the first reservoir 16 is so high that when the piston 3 is in the region of its lower reversal point, the hydraulic fluid can flow into the hydraulic pipeline H through the replenishment pipeline 17. For example, the pressure of the hydraulic fluid in the first reservoir 16 is a few bar, e.g., less than 10 bar (1 MPa), or about 4-5 bar (0.4-0.5 MPa). This pressure is selected to be sufficiently high so that when the piston 3 is in the region of its lower reversal point, the hydraulic fluid can flow into the hydraulic pipeline H through the check valve 171. When the pressure reducer 15 returns to its initial position, the check valve 171 closes. Subsequently, as the piston 3 compresses, the hydraulic pressure in the hydraulic chamber 10 increases, and therefore the pressure in the hydraulic pipeline H also increases. Of course, it is in principle possible to supply the hydraulic medium into the first reservoir 16 at high pressures of up to 100 bar, for example, but a pressure of less than 16 bar is preferred because, in that case, the first reservoir 16 and its supply components are not considered a high-pressure system and can therefore be designed with a single wall.

[0077] Figure 2 schematically illustrates a second embodiment of the large diesel engine 1 according to the present invention. In the following description of the second embodiment, only the differences from the first embodiment will be described in more detail. The other descriptions of the first embodiment apply to the second embodiment in a similar or analogous manner. In the second embodiment, the same parts or parts with equivalent function as in the first embodiment are designated by the same reference numerals.

[0078] In the second embodiment, the toggle lever 14 of the pressure transmission device 13 is used not only for the hydraulic pipeline H but also for other purposes. Therefore, in the second embodiment, the toggle lever 14 is not designed as an independent toggle lever 14 used exclusively for the pressure transmission device 13, but the toggle lever 14 also works for other purposes. In the second embodiment, the toggle lever 14 is also used to introduce a hydraulic medium into the hydraulic chamber 10 and to actuate the pressure release of the hydraulic chamber 10.

[0079] The toggle lever preferably has a feed section Z that can supply hydraulic fluid to the hydraulic chamber 10. The feed section Z is designed, for example, as a conduit located in or on the toggle lever 14. For better understanding, the feed section Z is depicted with a solid line and can be distinguished from the hydraulic conduit H which serves as the pressure measuring element.

[0080] The supply unit Z is connected to the hydraulic chamber 10 via a flow path on one end and can be connected to a second reservoir 23 for the hydraulic medium via a setting device 22 on the other end. In the second reservoir 23, the hydraulic medium is present at a significantly higher pressure than in the first reservoir 16, for example, 30 times the pressure in the first reservoir 16. The pressure in the second reservoir 23 is selected so that the hydraulic chamber 10 can become pressure sufficient to lift the piston rod 6 and piston 3 toward the combustion chamber 4, i.e., upward in the figure, in order to adjust the compression ratio.

[0081] The second reservoir 23 and setting device 22 are stationary relative to the engine housing 200.

[0082] The setting device 22 is responsible for adjusting the pressure in the supply section Z so that a desired compression ratio can be obtained. The setting device 22 is designed, for example, as a proportional valve that is signaled to the engine control system 100 via a third signal line S3, so that the engine control system 100 can activate the setting device 22. When attempting to raise the piston rod 6 and piston 3 to increase the compression ratio, the engine control system 100 activates the setting device 22, i.e., a proportional valve, so that pressure is generated in the supply section Z, and thus in the hydraulic chamber 10, to raise the piston 3 to the desired position.

[0083] A second check valve 25 is further provided within the supply unit Z, located between the toggle lever 14 and the hydraulic chamber 10, i.e., not located in the stationary part of the supply unit Z. The second check valve 25 is preferably fixed to the crosshead 7. The second check valve 25 prevents the hydraulic medium from flowing back from the hydraulic chamber 10 into the supply unit Z.

[0084] The setting device 22 is designed, as depicted in Figure 2, to connect the supply unit Z to a second reservoir 23 which has a higher pressure, or to completely shut off the flow path connection between the supply unit Z and the second reservoir 23. This state depicted in Figure 2 means that the setting device 22, which is preferably designed as a proportional valve, is completely closed.

[0085] When attempting to increase the pressure in the hydraulic chamber 10 to change the compression ratio, the supply unit Z is flow-connected to the second reservoir 23 via the setting device 22, and the pressure is increased to the desired value via proportional control. Depending on leakage or periodic consumption, the fully open flow cross-sectional area of ​​the proportional valve will increase or decrease, for example. Therefore, if a large amount of hydraulic fluid is required to fill the hydraulic chamber 10, the flow cross-sectional area in the setting device 22 will increase accordingly, and as a result, more hydraulic fluid will be able to flow into the hydraulic chamber 10 during that part of the work cycle when the piston 3 is in the region of its lower reversal point.

[0086] Parallel to the connection from the supply unit Z to the second reservoir 23, the first reservoir 16 is also flow-connected to the supply unit Z via a backup conduit 172 and a third check valve 174. The third check valve 174 is positioned and designed so that the hydraulic medium can flow from the first reservoir 16 to the supply unit Z, but not in the reverse direction from the supply unit Z to the first reservoir 16. Furthermore, a first throttling mechanism 173 is provided between the third check valve 174 and the first reservoir 16 in the backup conduit 172. The backup conduit 172 serves to ensure that even when the setting device 22 is completely closed and therefore the supply unit Z is disconnected from the second reservoir 23, the hydraulic medium in the supply unit Z will not be depleted, preventing an undesirable dry-running condition.

[0087] In all operating conditions, that is, even in operating conditions where piston rod lifting is not expected to occur, for example, during operation at a constant compression ratio, the backup pipeline 172 ensures that the supply unit Z is at least flow-connected to the first reservoir 16, and that there is no possibility of complete depletion due to leakage, for example. Here, the second check valve 25 prevents the hydraulic medium from flowing back from the hydraulic chamber 10 to the supply unit Z.

[0088] The pressure reducer 15 in the hydraulic pipeline H is flow-connected to the hydraulic chamber 10 on its high-pressure side. For this purpose, the pressure reducer 15 is flow-connected to the discharge pipeline 101 on its high-pressure side, and the discharge pipeline 101 functions to discharge the hydraulic medium from the hydraulic chamber 10.

[0089] Parallel to the pressure reducer 15, the discharge pipe 101 is flow-connected to a relief valve 26, which discharges the hydraulic fluid from the hydraulic chamber 10. Here, the relief valve 26 is designed as an operable two-way valve, and its control port 27 is flow-connected to the low-pressure side of the pressure reducer 15 via the hydraulic pipe H. Therefore, the relief valve 26 is operated by hydraulic pressure. Consequently, approximately the same amount of pressure as the pressure present on the low-pressure side of the pressure reducer 15 acts on the control port 27 of the relief valve 26.

[0090] As depicted in Figure 2, when the relief valve 26 is in its open position, the flow path connection between the discharge pipe 101 and the drain 28 is opened, allowing the hydraulic medium to flow from the hydraulic chamber 10 through the discharge pipe 101 and the relief valve 28 to the drain 28, and through the drain 28, the hydraulic medium can flow into a container 29, such as a tank or a water basin. When the relief valve 26 is in its closed position, the relief valve 26 closes the flow path connection between the discharge pipe 101 and the drain 28, preventing the hydraulic medium from flowing from the hydraulic chamber 10 to the container 29.

[0091] In this way, the discharge pipe 101, which is at the same pressure as the inside of the hydraulic chamber 10, is connected to the hydraulic pipe H via two parallel branches for pressure exchange. Specifically, on one side, it is connected via a pressure reducer 15 whose high-pressure side is connected to the discharge pipe 101 and whose low-pressure side is connected to the hydraulic pipe H, and on the other side, it is connected via a relief valve 26 whose control port 27 is connected to the hydraulic pipe H.

[0092] Thus, as long as the pressure in the control port 27 is high enough to keep the relief valve 26 in its closed position, the hydraulic fluid cannot flow out of the hydraulic chamber 10 through the drain 28. In order to discharge the hydraulic fluid from the hydraulic chamber 10, the pressure in the control port 27 is released, which opens the relief valve 26 and allows the hydraulic fluid to flow out through the relief valve 26 and the drain 28.

[0093] In the second embodiment, the hydraulic pipeline H thus further serves as a control pipeline for the relief valve 26 to discharge the hydraulic medium from the hydraulic chamber 10.

[0094] In the second embodiment as well, the hydraulic pipeline H is connected to the first reservoir 16 via a replenishment pipeline 17 on its stationary side, and the replenishment pipeline 17 opens into the hydraulic pipeline H at a point located between the fixed point 141 and the pressure sensor 11. A first check valve 171 is provided in the replenishment pipeline 17, and this check valve 171 is positioned so that the hydraulic medium can flow from the first reservoir 16 to the hydraulic pipeline H, but cannot flow from the hydraulic pipeline H to the first reservoir 16.

[0095] Furthermore, the hydraulic pipeline H is flow-connected to the relief pipeline 30 in its stationary region, and the relief pipeline 30 is provided with an operable switching element 31, which is designed here as a shut-off element, and the switching element 31 selectively opens and closes the flow through the relief pipeline 30. For example, the switching element 31 is designed as an operable shut-off valve, for example, as a two-way valve. The switching element 31 is signal-connected to the engine control system 100, for example, via a fourth signal line S4, so that the engine control system 100 can operate the control element 31.

[0096] In this embodiment, the relief pipeline 30 opens into another container 32, such as a tank or a water basin, which may be the same as the container 29.

[0097] As long as the switching element 31 blocks the passage of the relief pipe 30, the pressure inside the hydraulic pipe H is reduced via the pressure reducer 15 and is proportional to the hydraulic pressure in the hydraulic chamber 10, making it possible to determine the cylinder pressure in the combustion chamber 4 using the pressure sensor 11 and the evaluation unit 12. Furthermore, the discharge pipe 101 is kept closed via the relief valve 26.

[0098] For example, if the goal is to discharge the hydraulic fluid from the hydraulic chamber 10 in order to lower the compression ratio, the engine control system 100 activates the switching element 31 via the fourth signal line S4 so that the switching element 31 opens the relief line 30. This releases the pressure in the hydraulic line H, and in particular the pressure on the control port 27 of the relief valve 26 decreases. This opens the relief valve 26, and the hydraulic fluid can flow out of the hydraulic chamber 10 via the discharge line 101 and the drain 28.

[0099] Here, the second throttling device 175 in the replenishment line 17 prevents a sufficient amount of hydraulic fluid from flowing in from the first reservoir 16, and as a result the pressure in the hydraulic line H drops as desired, and the relief valve 26 switches to the open position.

[0100] Figure 3 schematically illustrates a third embodiment of the large diesel engine 1 according to the present invention. In the following description of the third embodiment, only the differences from the first and second embodiments will be described in more detail. The other descriptions of the first and second embodiments apply to the third embodiment in a similar or analogous manner. In the third embodiment, the same parts or parts with equivalent function as in the first and second embodiments are designated by the same reference numerals.

[0101] In the third embodiment, as in the second embodiment, the hydraulic connection that serves to relieve pressure in the hydraulic chamber 10 is used as the hydraulic pipeline H for pressure measurement, i.e., the toggle lever 14 of the pressure transmission device 13 is also used to activate the pressure relief in the hydraulic chamber 10.

[0102] However, unlike the second embodiment, in the third embodiment, the hydraulic pipeline H, which additionally acts as a control pipeline for the relief valve 26, can be optionally connected to either the first reservoir 16 or the second reservoir 23. For this purpose, the switching element 31 is designed in the third embodiment as a switching valve, e.g., a three-way / two-way valve, with its outlet connected to the relief pipeline 30, which is flow-connected to the hydraulic pipeline H. Of the two inlets of the switching element 31, one is flow-connected to the first reservoir 16, which is at a lower pressure, via the replenishment pipeline 17, and the other is flow-connected to the second reservoir 23, which is at a higher pressure, via the supply pipeline 231. Furthermore, a fourth check valve 232 is provided in the supply line 231, and the fourth check valve 232 is positioned and designed to allow the hydraulic medium to flow from the second reservoir 23 to the switching element 31, but to prevent the hydraulic medium from flowing back into the second reservoir 23 through the supply line 231.

[0103] When the switching element 31 is in the switching position shown in Figure 3, the hydraulic pipeline H is flow-connected to the first reservoir 16 via the relief pipeline 30 and the replenishment pipeline 17. When the switching element 31 is in the other switching position, the hydraulic pipeline H is flow-connected to the second reservoir 23 via the relief pipeline 30 and the supply pipeline 231.

[0104] The switching position of the switching element 31 depicted in Figure 3 is preferably used only when the VCR system is not operating. When the hydraulic pipeline H is flow-connected to the first reservoir 16, only the lower pressure in the first reservoir 16 is applied to the control port 27 of the relief valve 26. Depending on how large the selected low pressure in the first reservoir 16 may be, it may not be sufficient to keep the relief valve 26 in the closed position. Therefore, the switching position of the switching element 31 depicted in Figure 3 is preferably selected only when the VCR system is not operating.

[0105] When the VCR system is operating or when the VCR system is being activated, the switching element 31 is placed in a switching position in which the hydraulic pipeline H is flow-connected to the second reservoir 23, which has a higher pressure, via the relief pipeline 30 and the supply pipeline 231. Here, a higher pressure is applied to the control port 27 of the relief valve 26, so that it is reliably returned to its closed position and held there. The fourth check valve 231 also prevents backflow of the hydraulic medium when the pressure in the hydraulic pipeline H becomes greater than the pressure in the second reservoir 23, so that the cylinder pressure can be determined via the hydraulic pipeline H using the pressure sensor 11 and the evaluation unit 12.

[0106] In the third embodiment, it is not essential but possible to provide a check valve 171 in the replenishment pipeline 17.

[0107] The third embodiment has the advantage that, even when the VCR system is not operating, the hydraulic line H is connected to the first reservoir 16 via a flow path, so the toggle lever 14 or the hydraulic line H remains filled with the hydraulic medium and cannot be depleted.

[0108] Figure 4 schematically illustrates a fourth embodiment of the large diesel engine 1 according to the present invention. In the following description of the fourth embodiment, only the differences from the first, second, and third embodiments will be described in more detail. The other descriptions of the first, second, and third embodiments apply to the fourth embodiment in a similar or analogous manner. In the fourth embodiment, the same parts or parts with equivalent functions as those in the first, second, and third embodiments are designated by the same reference numerals.

[0109] In the fourth embodiment, pressure measurement is performed via a supply unit Z that delivers the hydraulic medium to the hydraulic chamber 10. This means that in the fourth embodiment, the supply unit Z is used as a hydraulic pipeline H for pressure measurement. In Figure 4 of the fourth embodiment, the hydraulic pipeline H for pressure measurement is depicted with a dashed line, and therefore the supply unit Z is also depicted with a dashed line.

[0110] In Figure 4, the control pipeline for operating the relief valve 26 is depicted with a solid line and designated by reference numeral L.

[0111] In the fourth embodiment, since the supply unit Z is used as the hydraulic pipeline H for pressure measurement, the pressure reducer 15 is flow-connected to the supply unit Z on the low-pressure side. The pressure reducer 15 is connected to the supply unit Z at a low-pressure point located between the toggle lever 14 and the second check valve 25 with respect to the flow direction. The pressure reducer 15 is flow-connected to the hydraulic chamber 10 on the high-pressure side, so that the high-pressure side of the pressure reducer 15 is applied at the pressure in the hydraulic chamber 10. The pressure reducer 15 is located parallel to the second check valve 25.

[0112] The supply unit Z is used as a hydraulic pipeline H for pressure measurement in the non-stationary region of the supply unit Z. A branching section 142 is provided in the stationary region of the supply unit Z, where the hydraulic pipeline H branches off from the supply unit Z and extends toward the pressure sensor 11. From this branching section 142, the supply unit Z extends toward the setting device 22. As already described in relation to the second embodiment, the setting device 22 is designed to optionally connect the supply unit Z to a second reservoir 23 with a higher pressure, or to completely close the flow path connection between the supply unit Z and the second reservoir 23, as depicted in Figure 4. This state depicted in Figure 4 means that the setting device 22, preferably designed as a proportional valve, is completely closed.

[0113] To enable pressure measurement through the hydraulic pipeline H, a fifth check valve 221 is provided at the flow path connection between the branch 142 and the setting device 22. The fifth check valve 221 is positioned and designed to allow the hydraulic medium to flow from the setting device 22 to the branch 142, but to prevent the hydraulic medium from flowing back from the branch 142 to the setting device 22.

[0114] In a manner similar to that described based on the third embodiment, the control pipeline L can be optionally connected to either the first reservoir 16 (lower pressure) or the second reservoir 23 (higher pressure) to actuate the relief valve 26.

[0115] Here, the filling or replenishment of the hydraulic medium into the hydraulic chamber 10 also occurs when the piston 3 is in the region of its lower inversion point, because this is where the pressure in the hydraulic chamber 10 is lowest throughout the operating cycle. The fifth check valve 221 and the second check valve 25 are opened to allow the hydraulic medium to flow into the hydraulic chamber 10 in a manner known to itself, for example, to compensate for any loss due to leakage. The pressure reducer 15 is brought to its initial position by the hydraulic medium, for example, by displacing the pressure piston of the pressure reducer 15. When compression in the cylinder 2 begins during the compression stroke of the piston 3, the pressure in the hydraulic chamber 10 increases, thereby closing the second check valve 25. As soon as the inflow of the hydraulic medium into the hydraulic chamber 10 stops, the fifth check valve 221 also closes. The pressure reducer 15 increases the pressure in the hydraulic pipeline H in proportion to the hydraulic pressure in the hydraulic chamber 10, and therefore in proportion to the cylinder pressure. The pressure in the hydraulic pipeline H is detected by a pressure sensor 11 on the stationary side of the hydraulic pipeline H and transmitted to an evaluation unit 12 for determining the cylinder pressure in the combustion chamber 4 of the cylinder 2.

[0116] Figure 5 schematically illustrates a fifth embodiment of the large diesel engine 1 according to the present invention. In the following description of the fifth embodiment, only the differences from the embodiments described so far will be described in more detail. The descriptions of the embodiments described so far will apply to the fifth embodiment in the same or similar manner. In the fifth embodiment, the same parts as in the embodiments described so far, or parts with the same function, are designated by the same reference numerals.

[0117] In the fifth embodiment, the pressure sensor 11 is positioned on the crosshead 7 so that the pressure sensor 11 moves together with the crosshead 7. Thus, in the fifth embodiment, the pressure sensor 11 is not stationary, but moves relative to the engine housing. The value determined by the pressure sensor 11 is then transmitted, though not essential, typically to a stationary evaluation unit 12. This transmission from the pressure sensor 11 to the evaluation unit 12 can be done via a wireless signal connection, for example, via a transmitter on the pressure sensor 11 and a receiver on the evaluation unit 12. Transmission can also be done by a first signal line S1 designed as a flexible cable, connecting the evaluation unit 12 to the pressure sensor 11 which is movable relative to the evaluation unit 12. Such a cable can be located, for example, on or inside a toggle lever. This may be one of the toggle levers in a large diesel engine 1, for example, located between a stationary engine housing 200 and a crosshead 7 that is moving in operation, in order to introduce a lubricating or cooling medium into the crosshead 7 or piston rod 6. For guiding the first signal line S1, a toggle lever can also be used to supply hydraulic fluid to the hydraulic chamber 10 in order to displace the piston rod 6. Naturally, in the fifth embodiment, the pressure reducer 15 can also be provided in the hydraulic pipeline H between the hydraulic chamber 10 and the pressure sensor 11.

Claims

1. at least one cylinder (2) having a combustion chamber (4) defined by a piston (3) arranged for reciprocal movement along a cylinder axis (A); A rotatable crankshaft (9) and A large diesel engine comprising: the piston (3) connected via a piston rod (6) to a crosshead (7) having a crosshead pin (71); the crosshead (7) connected to the crankshaft (9) via a push rod (8); a hydraulic chamber (10) provided in the crosshead pin (71), the hydraulic chamber (10) being defined by the piston rod (6); and the piston rod (6) being displaceable in the direction of the cylinder axis (A) relative to the crosshead pin (71) to adjust the compression ratio, 1. A large diesel engine, characterized in that it is provided with a pressure sensor (11) capable of determining the oil pressure in the oil pressure chamber (10) and an evaluation unit (12) capable of determining the cylinder pressure in the combustion chamber (4) by means of said oil pressure.

2. 2. A large diesel engine according to claim 1, wherein the pressure sensor (11) is positioned on the crosshead (7) so as to move with the crosshead (7).

3. 2. A large diesel engine according to claim 1, wherein the pressure sensor (11) is provided at a measurement point in a fixed position relative to the engine housing (200).

4. 4. A large diesel engine according to claim 3, further comprising a hydraulic pressure transmission device (13) capable of transmitting the hydraulic pressure in the hydraulic pressure chamber (10) to the pressure sensor (11).

5. 5. A large diesel engine according to claim 4, wherein the hydraulic transmission device (13) comprises a toggle lever (14) connected on the one hand to the crosshead (7) and on the other hand to the engine housing (200).

6. 6. A large diesel engine according to claim 5, wherein said toggle lever (14) has a supply (Z) capable of supplying hydraulic medium to said hydraulic chamber (10).

7. 7. A large diesel engine according to claim 6, wherein said toggle lever (14) has a control line (H) for a relief valve (26) for draining said hydraulic medium from said hydraulic chamber (10).

8. 7. A large diesel engine according to claim 6, wherein the hydraulic transmission device (14) comprises a pressure reducer (15) capable of applying the hydraulic pressure in the hydraulic chamber (10) and reducing this pressure.

9. 9. A large diesel engine according to claim 8, wherein the pressure reducer (15) is located on the crosshead (7).

10. 9. A large diesel engine according to claim 8, wherein the pressure reducer (15) has a high pressure side and a low pressure side, the high pressure side being fluidly connected to the hydraulic chamber (10), and the pressure on the low pressure side can be applied to the pressure sensor (11).

11. 9. A large diesel engine according to claim 8, wherein the pressure reducer (15) is fluidly connected to a discharge line (101) for discharging the hydraulic medium from the hydraulic chamber (10).

12. 11. A large diesel engine according to claim 10, wherein hydraulic medium can be supplied to the low pressure side of the pressure reducer (15).

13. 13. A large diesel engine according to any one of claims 1 to 12, designed as a longitudinally scavenged two-stroke large diesel engine.

14. 14. A heavy-duty diesel engine according to claim 13, designed as a dual-fuel heavy-duty diesel engine capable of operating in a liquid mode, in which liquid fuel is introduced into the combustion chamber (4) for combustion, and also capable of operating in a gas mode, in which gas is introduced into the combustion chamber (4) for combustion.

15. 1. A method for determining cylinder pressure in a heavy-duty diesel engine, said heavy-duty diesel engine comprising: at least one cylinder (2) having a combustion chamber (4) defined by a piston (3) arranged for reciprocal movement along a cylinder axis (A); A rotatable crankshaft (9) and The piston (3) is connected to a crosshead (7) having a crosshead pin (71) via a piston rod (6), the crosshead (7) is connected to the crankshaft (9) via a push rod (8), a hydraulic chamber (10) is provided in the crosshead pin (71), the hydraulic chamber (10) is defined by the piston rod (6), and the piston rod (6) is displaceable in the direction of the cylinder axis (A) relative to the crosshead pin (71) to adjust the compression ratio, 10. The method of claim 9, wherein the hydraulic pressure in the hydraulic chamber is determined by a pressure sensor (11) and an evaluation unit (12) determines the cylinder pressure in the combustion chamber (4) by means of the hydraulic pressure.

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