GEOMETRICAL COMPRESSION RATIO

The present invention relates to a method pertaining to internal combustion engines of the kind defined in the preamble of Claim 1. The invention also relates to an arrangement according to the preamble of Claim 6, for carrying out the method.

In Otto- or Diesel-type engines, the compression ratio is selected generally to meet full load operating conditions. At idle and part load, the compression ratio will be too low, resulting in high fuel consumption and increased exhaust emissions.

To counteract this, it has been proposed to provide an engine of this kind having one or more cylinders having a variable geometric compression ratio, comprising - for each cylinder - a combustion chamber formed between a piston and a cylinderhead and having a second cylinder which communicates with the combustion chamber and which includes a second piston whose end surface forms a movable wall for varying the volume of the combustion chamber. The second piston is connected with drive means to control the position of the second piston in the second cylinder, through the medium of control means.

U.S. 2,420,117 shows an engine of this kind wherein the cylinder housing the second piston, on the side opposite to the combustion chamber, is filled with hydraulic fluid which communicates, via a conduit, with a reservoir having an enclosed gas volume. When the pressure in the combustion chamber exceeds a given value, the second piston is lifted and hydraulic fluid forced into the reservoir. When the pressure decreases, the hydraulic fluid forces the piston down, back to its starting position. This is determined by an outlet to a sump. The fluid is pumped from the sump back to the reservoir, through a pressure relief valve and surplus fluid is returned to the sump.

This is a passive system which is unable to change the geometric compression ratio without large energy losses. The pump is a highly energy-consuming constant flow pump with a pressure relief valve.

U S 2,41 ,217 shows a similar system in which hydraulic fluid is delivered to a second cylinder through a check valve and out through a pressure relief valve in the form of a spring-biased needle valve The valve closes during the expansion stroke by a cam shaft This arrangement is also energy-consuming, in the form of engine energy to move the second piston and pump energy for returning the piston A drawback, common with both of the aforesaid solutions, is that there scarcely will be time to move the second piston in the manner intended

U S 4,987,863 shows an engine having a variable geometric compression ratio which is achieved by changing the vertical position of the second piston so as to provide optimal conditions at different operating conditions The piston communicates with a source of pressure medium via a solenoid-controlled valve The piston is moved outwards and inwards in relation to the combustion chamber, by virtue of the valve being opened when the pressure in the combustion chamber is low and high respectively The second piston remains stationary at constant engine load The valve is controlled by an electronic control system

Because the second piston remains stationary in a lifted position for a relatively long period of time, the lubricant present in the exposed cylinder bore will burn without being replaced with fresh oil before the piston again sweeps over its surface Soot and other deposits are not removed frequently, resulting in heavy wear and, in time, expensive engine repairs

An arrangement for varying the geometric compression ratio of an internal combustion engine is known from U S 5,188,066, for instance This arrangement includes the control of an auxiliary piston which runs in a cylinder that opposes the ordinary cylinder of the engine, with a conventional system comprised of a driven crankshaft, a connecting rod alternatively supplemented with a further connecting rod and a rocker arm A phase shifting device brings the auxiliary piston more or less out of phase with the engine piston and therewith changes the compression ratio

This arrangement with a crank shaft has an expensive and complicated construction which includes many moving parts The arrangement does not react cyclically to signals from the control system, but takes several cycles before changed operating conditions result in a changed geometric compression ratio As

with other known arrangements, it is unable to regulate the geometric compression ratio separately during an individual stroke of the cycle, for instance during the intake and exhaust strokes.

Accordingly, the object of the invention is to provide an improved method for providing a variable geometric compression ratio which is energy-efficient, responds rapidly to changed operating conditions, enables a variable and independent movement of the auxiliary piston within the 4-stroke cycle even during the intake and exhaust strokes, provides good lubrication and prevents wear and soot deposits on the second cylinder and its piston. Another object is to provide an arrangement in Otto and Diesel internal combustion engines which has an inexpensive and simple design with few moving parts and with low operating- and repair costs.

These objects are achieved in accordance with the invention with the inventive method and inventive arrangement defined in following Claims.

According to one particularly advantageous embodiment of the invention, drive means cause the end wall of the second piston to at least temporarily take a lower end position during each 4-stroke cycle or in a 4-stroke cycle that reoccurs at close intervals, irrespective of the level at which the second piston is positioned during the expansion stroke.

This will ensure that the lower bore of the second cylinder will be regularly lubricated and adhering soot particles will be removed, etc.

According to another advantageous embodiment of the inventive method, during each 4-stroke cycle or during 4-stroke cycles that reoccurs at close intervals, the drive means cause the end wall of the second piston to at least temporarily take an upper position which is spaced a distance from the lower end position of the second piston, irrespective of the level at which the second piston is positioned during the expansion stroke, said distance preferably corresponding at least to the outer distance between the piston rings of the second piston.

In this manner, the oil film on the lower bore of the second cylinder is constantly renewed due to the fact that the piston rings regularly move into that part of the

second cylinder which communicates with the oil mist filled crank case.

The invention will now be described in more detail with reference to exemplifying non-limiting embodiments thereof and also with reference to the accompanying 5 drawings, in which

Figure 1 is a cross-sectional, schematic view of a first embodiment of the invention of an internal combustion engine with a symbolically control system and valve arrangement, wherein the engine is shown at a stage where the compression stroke o has just ended at low engine load and

Figure 2 is a corresponding view of a second embodiment that includes an alternative valve arrangement.

5 The Figure 1 embodiment includes a cylinder 1 having a piston 2 which shares a combustion chamber 3 with a smaller cylinder 4 and its piston 5 which is connected to a hydraulic piston 7 through the medium of a piston rod 6. The hydraulic piston 7 runs in a double-acting hydraulic cylinder 8. The two chambers 9, 10 of the hydraulic cylinder 8, separated by the piston 7, are connected via passages 11, 12 0 to a symbolically illustrated and arbitrary directional valve 13. The valve 13 is also connected via a first conduit means1 to a hydraulic pump 15 which maintains the pressure at a set level, and a high pressure accumulator 16. A conduit 17 connects the valve 13 with a low pressure accumulator 19 and to a hydraulic tank 18 via a pressure regulator 29. A suction conduit 30 is connected between the hydraulic 5 pump 15 and the hydraulic tank 18. The valve 13 is controlled by an electronic control unit 20. The electronic control unit 20 receives information from a plurality of sensors, in similarity with a modern electronically controlled fuel injection system. This information may include the engine crank angle, the speed of the engine, knocking combustion, the throttle angle, inlet manifold pressure, coolant o temperature, etc. A position sensor 21 combined with an indicator pin 22 projecting up from the piston 7 delivers information relating to the positions of the piston 5 and the hydraulic piston 7 in their respective cylinders. The valve 13 can be controlled by the electronic control unit 20 so that the piston 5 and the hydraulic piston 7 obtain a balanced accelerating and retarding movement in respective directions. 5 The space in the cylinder 4 filled with an oil mist is connected to a cavity system 24 through the medium of a plurality of passages 23, said system 24 being connected

in turn to the engine crank case via the area surrounding the engine valve mechanism. The illustrated engine also includes an exhaust valve 25 with a tappet 26, and a cam lobe 27. The end surface 28 of the piston 5 is exposed to the combustion chamber 3. The end surface 28 of said piston 5 has a lower end position 28' and an upper end position 28". Although not shown, an oil cooler may be needed to control the temperature of the hydraulic fluid.

The valve 13 has a first valve position 31 in which passage 11 connects with conduit 17 and passage 12 connects with conduit 14, a fully closed position 32, and a second valve position 34 in which passage 11 connects with conduit 14 and passage 12 connects with conduit 17. The valve 13 is operated by a valve control device 36, e.g. a direct acting solenoid or an indirectly acting solenoid, which is able to operate the valve 13 very rapidly via a hydraulic servo in single or multiple stages and with small losses in both the regulating solenoid and the regulated hydraulic fluid.

The hydraulic piston 7 has a larger piston surface 7" exposed to the upper chamber 10 in the hydraulic cylinder 8, and a smaller surface T exposed to the lower chamber 9 in said hydraulic cylinder 8. The pressures in the high pressure accumulator 16 and the low pressure accumulator 19 are chosen so that when the higher pressure acts in the lower chamber 9, where the hydraulic piston 7 has a small exposed piston surface 7, and the lower pressure acts in the upper chamber 10, where the hydraulic piston 7 has a larger exposed piston surface 7", the hydraulic force will nevertheless be greater on the smaller piston surface 7' than on its larger surface 7", while the resultant force is, at the same time, sufficiently great to provide required acceleration and retardation of the piston 5 and the hydraulic piston 7. When the high pressure acts in both the lower chamber 9 and the upper chamber 10, the resultant force from the exposed piston surfaces 7' and 7" will also be sufficiently great to provide required acceleration and retardation of the piston 5 and the hydraulic piston 7. The greater part of the energy used to accelerate the piston 5 and the hydraulic piston 7 is recovered upon retardation thereof and stored in the accumulator 16. This phenomenon is referred to as a hydraulic pendulum. Only a small amount of energy is lost due to flow losses and friction, among other things. The energy loss is corresponded by a net flow with potential drop between the input and output of the hydraulic pump.

The first embodiment of the invention operates in the following manner. At the end of the exhaust stroke, the piston 5 and the hydraulic piston 7 are located in their respective lower end positions. The electronic control unit 20 computes which 5 geometric compression ratio is suitable for each individual compression stroke. This computation is based on a program stored in the electronic control unit 20, and on the basis of information received by the electronic control unit 20 from its respective sensors. During the intake stroke, the valve 13 is in position 34 controlled by the electronic control unit 20, so that hydraulic fluid will flow from the o hydraulic pump 15 and the accumulator 16 into the lower chamber 9 of the hydraulic cylinder 8, through the conduit 14 and the passage 11 , therewith forcing the piston 5 and the hydraulic piston 7 upwards from their lower end positions. At the same time, hydraulic fluid will flow from the upper chamber 10 of the hydraulic cylinder 8 to the accumulator 19, through the passage 12, the valve 13 and the 5 conduit 17. When a certain pressure is reached in the accumulator 19, the remaining hydraulic fluid passes to the hydraulic tank 18, via the pressure regulator 29.

The piston 5 and the hydraulic piston 7 are accelerated and potential energy from 0 essentially the accumulator 16 is converted to kinetic energy of the piston 5 and hydraulic piston 7. When the piston 5 and the hydraulic piston 7 have reached a certain position, the valve 13 switches to valve position 31, wherewith the hydraulic piston 7 forces hydraulic fluid out from the upper chamber 10 of the hydraulic cylinder 8 to the conduit 14 and the accumulator 16 via the passage 12 and the 5 valve 13. At the same time, hydraulic fluid flows from the accumulator 19 to the passage 11 and the lower chamber 9 of the hydraulic cylinder 8 via the conduit 17 and the valve 13. The piston 5 and the hydraulic piston 7 are retarded to a stationary mode and the kinetic energy thereof is converted to potential energy in the accumulator 16. The valve 13 is then brought to its closed position 32. The o piston 5 and the hydraulic piston 7 have now been moved through a distance, that corresponds to the optimal geometric compression ratio at the current operating conditions, said distance being determined by the control unit 20.

During the compression stroke and the expansion stroke, the valve 13 is in its 5 closed position 32 such that the piston 5 and the hydraulic piston 7 remain locked in their positions. The valve 13 is caused to take position 31 at the earliest when

the exhaust valve 25 is opened, so that hydraulic fluid flows from the hydraulic pump 15 and the accumulator 16 to the upper chamber 10 of the hydraulic cylinder 8 via the conduit 14, the valve 13 and the passage 12. At the same time, hydraulic fluid flows from the lower chamber 9 of the hydraulic cylinder 8 to the accumulator 19, via the passage 11 , the valve 13 and the conduit 17. When the accumulator 19 is filled, remaining fluid passes to the hydraulic tank 18 via the pressure regulator 29. The piston 5 and the hydraulic piston 7 are accelerated and potential energy from essentially the accumulator 16 is converted to kinetic energy of the piston 5 and the hydraulic piston 7. When the piston 5 and the hydraulic piston 7 have reached a certain position, the valve 13 is switched to valve position 34, so that the hydraulic piston 7 will force hydraulic fluid from the lower chamber 9 of the hydraulic cylinder 8 to the passage 14 and the accumulator 16, via the passage 11 and the valve 13. At the same time, hydraulic fluid flows from the accumulator 19 to the passage 12 and to the upper chamber 10 of the hydraulic cylinder 8, via the passage 17 and the valve 13. At the end of the exhaust stroke, the piston 5 and the hydraulic piston 7 have been retarded to a stationary mode and their kinetic energy has been converted to potential energy in the accumulator 16. The piston 5 and the hydraulic piston 7 are immediately accelerated in the opposite direction during the subsequent intake stroke.

The Figure 2 embodiment differs from the Figure 1 embodiment solely with regard to the control system for the piston 5 and the hydraulic piston 7 and is solely described with regard to the differing components and functions.

An engine-driven hydraulic variable displacement pump 40 comprises a cam disc

40a, a pump plunger 40b, a cylinder 40c, a normally open valve 40d, a solenoid 40e and a pump plunger return spring (not shown). The pump is connected with a conduit 44 and the passage 11 via a check valve 42, said passage 11 leading to the lower chamber 9 of the hydraulic cylinder 8. The conduit 44 also connects with a high pressure valve 46 and with the passage 12 that leads to the upper chamber

10 of the hydraulic cylinder 8. The high pressure accumulator 16 and a pressure sensor 48 are connected to the conduit 44. The upper chamber 10 of the hydraulic cylinder 8 is connected to a conduit 54 and to the low pressure accumulator 19 via a passage 50 and a low pressure valve 52, and is also connected to an accumulator 58 and to the suction side of the pump 40 via a pressure regulator 56. The pressure level of the accumulator 58 is higher than atmospheric pressure and

lower than the pressure in the low pressure accumulator 19 The electronic control unit 20 is connected to the pressure sensor 48, to a control servo 60 for the high pressure valve 46, and to a control servo 62 for the low pressure valve 52, and controls the hydraulic pump 40 so as to obtain a constant hydraulic pressure from its outlet The control program in the electronic control unit 20 bases control of the valves 46, 52 on an essentially constant hydraulic pressure from the pump 40, among other things

As the cam disc 40a rotates, the pump plunger 40b is caused to move reciprocatingly in the cylinder 40c As the pump plunger 40b moves towards the cam disc 40a, the valve 40d is open and the cylinder 40c fills with hydraulic fluid taken from the accumulator 58 As the pump plunger 40b moves in the opposite direction, hydraulic fluid is pumped back to the accumulator 58 provided that the valve 40d is open If the valve 40d is closed, hydraulic fluid is forced out through the check valve 42 and to the high pressure side of the system This enables the electronic control 20 unit to control the amount of fluid pumped out from the pump 40, and therewith the pressure on the high pressure side, to be controlled very accurately by means of the solenoid 40e and the valve 40d This means that less energy will be wasted compared to a constant flow pump that includes a pressure relief valve

The arrangement according to the second embodiment operates in the following manner When switching between an exhaust stroke and an intake stroke, the piston 5 and the hydraulic piston 7 are in their respective lower end positions During the intake stroke, the valve 52 is opened by the electronic control unit 20 so that hydraulic fluid will flow from the upper chamber 10 of the hydraulic cylinder 8 to the accumulator 19, via the passage 50, the valve 52 and the conduit 54 When a certain pressure has been reached in the accumulator 19, remaining hydraulic fluid will pass to the accumulator 58 via the pressure regulator 56 At the same time, hydraulic fluid flows from the hydraulic pump 40 and the accumulator 16 to the lower chamber 9 in the hydraulic cylinder 8, via the conduit 44 and the passage 11 , wherewith the piston 5 and the hydraulic piston 7 are forced upwards from their lower end positions

The piston 5 and the hydraulic piston 7 are accelerated and potential energy from essentially the accumulator 16 is converted to kinetic energy of the piston 5 and the

hydraulic piston 7. When the piston 5 and the hydraulic piston 7 have reached a certain position, the valve 52 closes and the valve 46 opens, wherewith the hydraulic piston 7 forces hydraulic fluid from the upper chamber 10 of the hydraulic cylinder 8 to the conduit 44 and the accumulator 16, via the passage 12 and the valve 46. At the same time, hydraulic fluid flows from the accumulator 16 to the lower chamber 9 of the hydraulic cylinder 8, via the conduit 44 and the passage 11. The piston 5 and the hydraulic piston 7 are retarded to a stationary mode and their kinetic energy is converted to potential energy in the accumulator 16. The valve 46 closes. The piston 5 and the hydraulic piston 7 have now been moved through a distance that corresponds to the optimal geometric compression ratio at the current operating conditions, said distance being determined by the control unit 20.

During the compression stroke and the expansion stroke, the valves 46, 52 are closed, so that the piston 5 and the hydraulic piston 7 remain locked in their respective positions. The valve 46 is opens at the earliest when the exhaust valve 25 is opens, so that hydraulic fluid will flow from the hydraulic pump 40 and the accumulator 16, the conduit 44 and the valve 46 to the upper chamber 10 of the hydraulic cylinder 8 through the passage 12. At the same time, hydraulic fluid flows from the lower chamber 9 of the hydraulic cylinder 8, and the passage 11 to the accumulator 16, via the conduit 44. The piston 5 and hydraulic piston 7 are accelerated and potential energy from essentially the accumulator 16 is converted to kinetic energy of the piston 5 and the hydraulic piston 7. When the piston 5 and the hydraulic piston 7 have reached a certain position, the valve 46 closes and the valve 52 opens so that the hydraulic piston 7 will force hydraulic fluid from the lower chamber 9 of the hydraulic cylinder 8 to the accumulator 16, via the passage 11 and the conduit 44. At the same time, hydraulic fluid flows from the accumulator 19 to the passage 50 and to the upper chamber 10 of the hydraulic cylinder 8, via the conduit 54 and the valve 52. At the end of the exhaust stroke, the piston 5 and hydraulic piston 7 have been retarded to a stationary mode and their kinetic energy have been converted into potential energy in the accumulator 16. The piston 5 and the hydraulic piston 7 immediately begin to accelerate in the opposite direction during the following intake stroke of the engine.

Consequently the second embodiment illustrated in Figure 2 operates in a similar manner as the first embodiment shown in Figure 1 including said hydraulic pendulum. Potential energy is converted to kinetic energy and then back to

potential energy, in the same manner as with a mechanical pendulum The Figure 2 embodiment with constant high pressure in the lower chamber 9 of the hydraulic cylinder 8 and alternating high and low pressure in the upper chamber 10 and subsequent to adapting the size of the piston surfaces 7' and 7" in respective chambers 9, 10 and respective pressures in the high and low pressure sides of the system provides essentially balanced acceleration and retardation forces and therewith lower losses than the first embodiment Furthermore, the second embodiment is a closed loop system having an accumulator 58 instead of a hydraulic tank 18, which also results in lower losses A closed loop system can also be used in the first embodiment

It is very important that the piston 5 will return to its lower end position 28' with great accuracy The position sensor 21, 22 is therefore arranged to deliver a signal to the electronic control unit 20 when the piston is in its lower end position 28' The signal is used by the electronic control unit to calibrate the time dependent movement of the piston 5 and the hydraulic piston 7 Since the piston 5 returns to its lower end position 28', soot deposits, etc are removed from the smaller cylinder bore 4 and a satisfactory lubrication and sealing of the piston rings is ensured

At idle, engine braking and at low engine load respectively when the combustion- governed positional changes of the piston 5 are nonexistent or small, it may be suitable to cause the piston to effect a stroke movement at least at regular intervals so as to guarantee satisfactory lubrication and functioning of the piston rings and also to replace hydraulic fluid in the hydraulic cylinder 8

In some cases there may be a need to increase the amount of residual exhaust gases, so as to change the chemical composition of the engine exhaust gases In such cases, the piston 5 can be controlled so that its return to its lower end position 28' will be terminated during the intake stroke The piston 5 may also be caused to return so that prior to moving to its lower end position 28', it first moves in the opposite direction and then turns immediately and moves to its lower end position 28' This mode of operation can reduce the occurrence of unburned hydrocarbons in the engine exhaust gases

The oil mist enclosed in the cylinder 4 flows through a plurality of passages 23 to and from the cavity system 24 in response to the reciprocating movement of the

piston 5. The cavity system 24 is, in turn, connected to the crank case via the area surrounding the engine valve mechanism. This enables the oil mist used to lubricate the cylinder 4 and the piston 5 to be replaced. The valve tappet 26 and the cam lobe 27 also contribute to replacement of the oil mist as they move.

It will be evident that the invention is not restricted to the illustrated exemplifying embodiments thereof and that modifications can be made within the scope of the inventive concept as defined in the following Claims. The invention also includes modifications evident to the person skilled in this art, for instance various known variations of the hydraulic piston control system. For instance, a certain displacement between the various phases of the second piston 5 may be arranged. The idling position of the piston 5 is not necessarily the lower end position 28' of the piston. The position sensor 21 may alternatively be disposed within the hydraulic cylinder 8, such that the indicator pin 22 does not pass through the cylinder wall. Other position indicators may be used, preferably contactless indica¬ tors.

The cylinders of the described engine and of the engine defined in the following Claims are upstanding cylinders, although it will be understood that the cylinders may be orientated in any position whatsoever within the scope of the invention.