GB2431695A

GB2431695A - Internal combustion five-stroke engine with opposed pistons and eccentric gearing

Info

Publication number: GB2431695A
Application number: GB0620985A
Authority: GB
Inventors: Gregory Lawrence Leonard
Original assignee: Individual
Current assignee: Individual
Priority date: 2005-10-28
Filing date: 2006-10-23
Publication date: 2007-05-02
Also published as: GB0521960D0; GB0620985D0

Abstract

An i.c.engine has opposed pistons 5,6 sliding in a cylinder 1, with a working space 2 between inlet ports 3 and exhaust ports 4. Connecting rods 7,8 are attached to crank pins 9,10 on second cranks 11,12 which are attached to rolling eccentric gears 13,14 which roll around stationary eccentric internal gears 15,16. The bearing of each rolling gear 13,14 is in turn driven by a main crank 17,18. With a 3:1 ratio of internal gear to rolling gear, each main crank 17,18 rotates clockwise once, and each second crank 11,12 rotates twice, for the full five-stroke cycle. Alternatively, the ratio may be 3:2. The crank pins execute hypotrochoidal motion. Preferably there exists separate adjustment to the mechanisms driving each piston and to the phase link between those mechanisms; each separate adjustment may be used to control compression ratio (CR), and in combination may be used to control CR, exhaust gas recycling (EGR) and intake capacity. The cylinders may be arranged in a hexagonal formation (fig.16) or a rhomboid formation (fig.17). The invention provides high efficiency through better combustion and prolonged expansion.

Description

* 2431695

I

IMPROVED INTERNAL COMBUSTION ENGINE

TECHNICAL FIELD

This invention refers to the control of compression ratio (CR), exhaust-gas recirculation (EGR) and intake capacity of an internal combustion (IC) engine which has two opposed pistons sliding in a chamber with a working space between inlet and exhaust ports. Each piston has a separate driving mechanism, and the two driving mechanisms are linked to a common power output mechanism. The engine operates with five strokes comprising the normal four stroke cycle and a fallow, or return, stroke between exhaust and induction.

SUMMARY OF THE iNVENTION

This invention adds to the normal advantages of a five-stroke engine by high efficiency through improved combustion, prolonged expansion, and by enabling dynamic variation of CR, EGR and induction volume.

According to one aspect of the invention there is provided an IC engine comprising: * at least one cylinder, having one or more inlet ports to towards one end of the cylinder and one or more outlet ports towards the other end of the cylinder; * a pair of opposed pistons in the cylinder; * a pair of shafts arranged at opposite ends of the cylinder with the axes of the shafts transverse the central axis of the cylinder, the shafts being mechanically coupled for rotation in phase with each other; * a pair of eccentric internal gears normally stationary, each arranged concentric with a respective one of the shafts; * a pair of matching eccentric pinions with bearings, each in rolling mesh with a respective one of the eccentric internal gears; * a pair of cranks, each drivingly connecting the bearing of a respective one of the eccentric pinions to a respective one of the shafts; * a pair of connecting rods, each connecting a pin eccentrically fast with a respective one of the eccentric pinions to a respective one of the pistons, the arrangement being such that rotation of the shafts causes the pistons to approach and recede from each other in a five stroke cycle, the five strokes being: i. induction as the pistons recede from each other with the inlet port(s) open between them, ii. compression as the pistons approach each other with no ports being open between them, iii. expansion as the pistons recede from each other with no ports being open between them, iv. exhaust as the pistons approach each other with the outlet port(s) open between them, and v. fallow in which both pistons move from having the outlet port(s) open between them to having the inlet port(s) once more open between them, with the expansion stroke being greater than the compression stroke for enhanced thermodynamic efficiency.

The engine can be a compression ignition engine with at least one fuel injection point between the inlet and the outlet ports or a spark ignition engine with at least one correspondingly placed spark plug, or an Homogeneous Charge Compression Ignition engine. It is suitable for any fuel.

In accordance with an important preferred feature of the invention, the engine is adjustable by known means, either by adjustment of the rotary position of the or each eccentric internal gear or by adjustment of the phase of the shafts.

Adjustment of the shaft phase can be effected by for instance arranging each to carry a drive pinion in mesh with a respective driven gear on a third shaft, the driven gears being rotationally adjustable with respect to each other, or by providing chain drive between the shafts with adjustable tensioners in both runs of the chain between respective sprockets on the shafts.

Adjust facilitates for instance, higher compression for starting of a compression ignition engine or for running on LPG after starting on petrol of a spark ignition engine.

Similarly adjustment can be made of the induction stroke for part load running and of exhaust gas re-circulation, if desired.

The eccentricity of each planetary gear can be chosen to alter the piston movement such as to improve the combustion efficiency, and minimise the maximum gear tooth stress. Alternatively the eccentricity can be reduced to nothing, so that standard circular gears may be used.

The dimensions of the main cranks and the second cranks due to the pin eccentricity, coupled with the initial setting of the shaft phase and the stationary internal gear angles, provides the required five stroke cycle.

The dynamic adjustment of the shaft phase and the stationary gear angles provides for variation of CR, swept volume and EGR. In practice only one of the the adjustments is normally needed for a simple CR change.

This invention also provides for reduced rotational masses by shaping rotating gears, so their face width approximately matches the maximwn cyclic variation in gear stress around their circumference.

According to a second aspect of the invention, the engine is asymmetric in having the eccentric planetary gear arrangement of the first aspect at one end of the or each cylinder and a conventional crankshaft and connecting rod at the other end.

BRIEF DESCRIPTION OF THE DRAWINGS

The first embodiment of the invention is depicted in figures 1:11, and the second embodiment in figures 12:15. Multiple configurations for the two embodiments are shown in figures 16,17.

FIGURE 1 shows a simplified section through the whole engine just before the start of the exhaust stroke.

FIGURE 2 shows the hypotrochoidal curve traced by the crank pin controlling the motion of the exhaust piston of the engine depicted in Figure 1.

FIGURE 3 shows the hypotrochoidal curve traced by the crank pin controlling the motion of the inlet piston of the engine depicted in Figure 1.

FIGURE 4 shows a simplified section through an engine identical to that in figure 1, but with longer connecting rods and a different geared drive; it is shown just before the start of the exhaust stroke. it is shown at approximately 2/3 scale of figure 1. -ç

FIGURE 5 shows the movement within the cylinder of the two piston crowns for one full cycle of the engines in figure 4, with the section through the engine shown for each 90 degrees of main crank movement.

FIGURE 6 shows the gaps between the two piston crowns of the engine in figure 4 when the stationary internal gears and the phase adjustable link between crankshafts are set firstly for maximum power, and secondly for maximum efficiency.

FIGURE 7 shows graphs of the change in compression ratio for each of the three adjustable settings of the engine in figure 4 at its design point setting shown by the movements of figure 5.

FIGURE 8 shows the lines of constant inducted volume within the limits for the range of adjustments of the two stationary internal gears.

FIGURE 9 shows a graph of the typical gear tooth force through a complete full cycle of the engine.

FIGURE 10 shows how the tooth force of figure 9 maps on to the maximum gear stress around the circumference of the rolling spur gear.

FIGURE 11 shows how the spur gear can be trimmed to cut down its weight FIGURE 12 shows a simplified section through the second embodiment of the engine just before the start of the inlet stroke.

FIGURE 13 shows the motion within the cylinder of the two piston crowns of the figure 12 engine.

FIGURE 14 shows the gaps between the piston crowns for settings of minimum and maximum capacity of the engine in figure 12 FIGURE 15 shows graphs of the change in compression ratio for each of the two adjustments at the design setting of the engine in figure 12.

FIGURE 16 shows a multiple arrangement for the engine in figure 4.

FIGURE 17 shows a multiple arrangement for the engine in figure 12.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

As shown in figure 1, the engine comprises a cylinder I with a working space 2 between the inlet ports 3 and the exhaust ports 4. The combustion space lies between the inlet piston crown and the exhaust piston crown 31. The inlet and exhaust pistons 5,6 are driven in the cylinder by connecting rods 7,8 attached to crank pins 9,10.

The crank pins are on the second cranks 11,12 which are rigidly attached to rolling eccentric gears 13,14 which roll around stationary eccentric internal gears 15,16. The bearing of each rolling gear is in turn driven by a main crank 17,18.

Also shown is a schematic of a spark plug 21 to initiate combustion if needed. This would also be the position for fuel injectors if required.

In this engine, each main crank rotates clockwise once for the full 5 stroke cycle, and each second crank rotates anticlockwise twice. The engine is shown just before the start of the exhaust stroke The angular adjustment of the stationary eccentric internal gears, and the phase adjustment of the link between the main cranks are by known means, and are not shown in the figures.

Figures 2 and 3 show the required eccentric hypotrochoidal motion of the inlet crank pin 19, and exhaust crank pin 20. The motion is achieved with either a 3:2 or a 3:1 ratio of internal gear to rolling gear. With a 3:1 ratio, the main crank rotates once and the second crank rotates twice for the full five stroke cycle. For a 3:2 gearing, the rotations and lengths of main and second cranks must be switched.

Neither hypotrochoidal curve is necessarily symmetrical with the axis of the cylinder, nor necessarily centred on the axis of the cylinder. The particular asymmetry for each, results in the desired irregular movements for the two pistons.

None of the main and second cranks are necessarily of the same dimensions. The size of each internal gear is fixed by the main crank and the selected eccentricity. For zero eccenthcity the gears are circular, and the internal gear has a pitch circle diameter (PCD) of six times the length of its main crank for 3:2 ratio, or three times for the 3:1 ratio gearing.

Figure 4 shows the same engine as in figure 1, but this time with longer connecting rods and the alternative 3:2 gearing with zero eccentricity. The diagram has been drawn to a smaller scale to fit the page. The engine is at the start of the exhaust stroke as in figure 1. The inlet main crank 24 matches exactly in length and approximately in rotation, the inlet second crank 11 of the engine in figure 1. Similar matches between the inlet second crank 25 and the main crank 17 of figure 1, and between main 28 and second 29 exhaust cranks and the second 18 and main 12 exhaust cranks of figure 1.

Comparison of the engines in figures 1 and 4 shows the advantage of a 3:1 gear ratio being more compact and enabling shorter connecting rods to be used. The disadvantage of the 3:1 gearing is that the peak torque forces which occur shortly after ignition, increase the gear tooth stresses from the 3:2 gearing by a factor of approximately 4*(second crank length/main crank length)2. This peak gear tooth stress can be much reduced by using eccentric gearing such that the peak torque forces occur at the maximum radius of the eccentric gear.

In figure 5 the movements of the inlet piston crown 30 and exhaust piston crown 31 of the engine in figure 4 is shown through one full five stroke cycle; shown separated by Lines marking the starts of induction 34, compression 35, expansion 36, and exhaust 37. The fallow stroke starts at 38 and finishes at 39 which is the start of the next induction. The working space of the cylinder lies between the start of the inlet port 32 and the exhaust port 33. At each 90 degrees of travel of the main cranks, the configuration of the figure 4 engine is shown at approximately a quarter of the scale of the shown working space gap.

Figure 6 shows, with a solid line, how the gap 40 between the pistons varies over the five strokes for the engine shown in figure 4. The angular settings for the internal gears 22, 26 and the phase angle between cranks 24, 28 are such that the engine is operating with a compression ratio of 10 and at maximum intake capacity. The dotted line 41 shows the piston gap after the adjustment of the two internal gears and the phase angle between cranks, to generate the minimum intake capacity. This new setting still maintains the compression ratio of 10, but allows a significant amount of exhaust gas recycling (EGR) as shown by the long dotted line at fallow start 38. The EGR coupled with the reduced piston gap 35 at the end of induction, results in a fresh intake charge of less than half that shown by the line 40. Both curves show prolonged expansion, with the gaps at the end of expansion 37 being greater than the gaps at the start of compression 36.

Figure 7 shows graphs of the change in compression ratio for each of the three possible adjustments for the engine shown in figures 4,5. The design setting is for a compression ratio of 10. One adjustment is made, in turn, with the other two maintained at the design setting.

Figure 8 shows the lines of constant inducted volume for the engine shown in figures 4,5, within the limits for the range of adjustments of the two stationary internal gears. The link phase being adjusted to keep a compression ratio of 10. The boundary is formed partly by geometrical constraints of keeping the pistons apart, and within the working space of the cylinder during compression and expansion. The rest of the boundary is due to keeping the CR at a constant 10, and the EGR to a maximum of 40%. The design point is marked, and it has the motion as shown in figure 5. The minimum capacity point marked on the negative limit on the inlet internal angle gives rise to the piston gap 41 as shown in figure 6; this setting has prolonged expansion as well as the highest EGR, and has the maximum efficiency. Similarly, the maximum capacity point marked on the positive limit on the inlet internal gear angle gives rise to the piston gap 40 in figure 6; this setting, with negligible EGR, delivers the maximum power but still at high efficiency due to the prolonged expansion.

The working gas pressures during compression, combustion, expansion and fallow result in a cyclic variation in the torque and the resulting gear tooth stresses as shown in figure 9. The start of induction 34, compression 35, expansion 36, exhaust 37, fallow 38 and end of fallow 39 are shown. The maximum gear tooth stress can be minimised by use of eccentric gears.

The profile shown is that of the non-eccentric inlet gearing for the maximum power of figure 5 motion. The tooth stress profile for the exhaust gearing, and the subtly different results for the 3:1 gearing are broadly similar. The rolling gear on both 3:2 and 3:1 ratios will have three rotations for the full profile of figure 9. When the maximum stresses are mapped onto the circumference of the spur gears, the result is shown in figure 10. To reduce the weight of the rotational masses, the spur gears can be trimmed, so that the face width is a maximum only in the region where the stresses are a maximum. A typical cut gear is shown in figure 11.

A second embodiment of the engine is now described, in which only the inlet piston is driven by a crank pin with eccentric hypotrochoidal motion; the exhaust piston is driven by a single crank. The resulting engine will still follow the same five stroke cycle, but will only have two mechanisms for controlling CR and capacity. This second embodiment of the engine is shown in figure 12 without the angular adjustment of the stationary internal gear, and the phase adjustment of the link between the main cranks as these are by known means.

In figure 12, the inlet port 3 is controlled by the inlet pistonS which is driven by a geared mechanism to give the required eccentric hypotrochoidal motion to the crank pin. The main crank 17 rotates in a clockwise direction once for the full cycle.

The exhaust piston 6 is driven by a single crank 42 which rotates clockwise once for the full cycle.The engine is shown at the start of induction, with the motions 30,31 of the inlet and exhaust pistons shown in figure 13 for the full five strokes. Again the limits of the combustion space and the start of inlet and exhaust ports 32, 33 are shown, as are the starts of induction 34, compression 35, expansion 36, exhaust 37 and fallow 38. The end of fallow 39 is also the start of the next induction. Figure 14 shows the gaps between the pistons for the full cycle, with the solid line 43 representing the maximum intake capacity, and the dotted line 44 the minimum capacity. As can be seen, both profiles have extended expansion with the lines 37 at the end of expansion being longer than lines 35 at the start of compression. The extra work extracted during expansion more than makes up for the pumping losses during the fallow stroke.

Figure 15 shows graphs of the change in compression ratio for the two possible adjustments.

The design setting is that shown in figure 13, centred on a compression ratio of 10, and each adjustment is made, in turn, with the other maintained at the design setting.

Multiple configurations of the first embodiment are quite sensible in that any geared crank can drive up to three pistons with identical motion in cylinders spaced at 120 intervals around the gears. With the 720 travel of the faster moving crank for a full cycle of the engine, it means that the other two cylinders are 240 and 480 delayed after the first.

Figure 16 shows one possible multiple configuration, with each inlet and exhaust crank driving two cylinders. The hexagonal arrangement has alternate inlet and exhaust gears; the inlet ports 3 and the exhaust ports 4 are marked, as is the angular position after the start of induction. With this arrangement, pairs of cylinders will be in synchronisation, and regular power strokes every 240 . If the inlet and exhaust cranks were set to rotate in opposite directions, with all the angles reversed precisely, then identical piston movements can be achieved. This time however, we would have triples of cylinders in synchronisation, with the two triples firing 240 apart and thus a slightly unbalanced firing pattern.

it is entirely feasible for the geometry of inlet and exhaust cranks to be identical, in which case the two hypotrochoidal curves would be the same, although at different orientations. If that is the case, then one crank can drive up to three inlet pistons and up to three exhaust pistons, with the inlet ends of cylinders spaced at 240 , and the exhaust ends of cylinders similarly. The orientation of the different sets of cylinders would match the different orientation of the two eccentric hypotrochoidal curves for inlet and exhaust gears.

Multiples of the second embodiment can similarly show the 120 spacing around the inlet gearing. Spacing around the single crank at the exhaust end can be at any arbitrary angle. If however 60 spacing is chosen, then a rhomboid formation of four cylinders is possible as is shown in figure 17. If both inlet main cranks rotate in the same direction and both exhaust cranks rotate in the opposite direction, then the firing order is 0 , 240 , 360 , 600 , 720 regular but slightly uneven. If the two inlet mains cranks rotate in opposite directions, and the two exhaust cranks rotate in opposite directions, then the firing order becomes 0 , 120 , 240 , 3 60 , 720 ... quite uneven.

While the foregoing is a description of the preferred embodiments of the invention, it will be understood that the invention is not limited to the particular embodiments shown and described herein, but that various changes and modifications may be made without departing from the scope or spirit of this invention as defined by the following claims.

Claims

CLAIMS

1. An internal combustion engine including one or more cylinders (1) each of which contains two opposed pistons (5,6) and a working space (2) between inlet and exhaust ports (3,4); the pistons are mounted to reciprocate in a respective cylinder (1) and are pivotally connected to connecting rods (7,8); the connecting rods (7,8) are pivotally connected to crank pins (9,10); the crank pins trace eccentric hypotrochoadal curves (19,20) resulting from geared assemblies (11,13) and (12,14) which are driven against a stationary eccentric gears (15,16) by main cranks (17,1 8); a transmission system linking the two cranks (17,18); the engine performs a five stroke cycle comprising the normal internal combustion engine four stroke cycle and a fifth fallow, or return, stroke between exhaust and induction.

2. An engine as claimed in claim 1 wherein one geared assembly and its respective stationary gear (12,14,16) or (11,13,15) is removed and the crank pin (9) or (10) is fast with the respective main crank (17) or (18); the respective crank pin tracing a circle for the full five stroke cycle of the engine.

3. An engine as claimed in Claims 1,2 wherein the eccentricity of any internal gear (15,16) is reduced to nothing and becomes an ordinary circular gear.

4. An engine as claimed in Claims 1,2,3 wherein the stationary gear (15) associated with the inlet piston (5) is made rotatable so the the resulting hypotrochoidal curve (19) traced by the crank pin (9) can be realigned as the engine is running, enabling control of compression ratio (CR).

5. An engine as claimed in Claims 1,2,3 wherein the stationary gear (16) associated with the exhaust piston (6) is made rotatable so the the resulting hypotrochoidal curve (20) traced by the crank pin (10) can be realigned as the engine is running, enabling control of CR.

6. An engine as claimed in Claims 1,2,3 wherein the transmission system linking the main cranks (17,18) is made adjustable so that one crank may be advanced or retarded whilst the engine is running, enabling control of CR.

7. An engine as claimed in Claims 1,2,3,4,6 wherein the stationary gear (15) and the transmission system are both adjusted whilst the engine is running, enabling control of intake volume, Exhaust Gas Recycling (EUR) and CR.

8. An engine as claimed in Claims 1,2,3,4,5,6 wherein the stationary gears (15,16) and the transmission system are all adjusted whilst the engine is running, enabling control of intake volume, EGR and CR.

9. Any combination of the engines as claimed in Claims I to 8 wherein a crank pin is used to drive more than one piston in an adjacent cylinder.

lOAn engine with the particular configuration of Figure 16 11.An engine with the particular configuration of Figure 17