DE19507082A1 - Orbital lift gear for cam operating system used in internal combustion engine - Google Patents

Abstract

For the cylinder (3) an area-wise guided piston (4) is provided, which together with a bent lever (5) engages on a body (6) effecting the guiding so that an effect on the movement path is produced. The piston remains in drive connection with an oscillating gear component (7) and thereby further forms chambers of alternating space content, whilst the piston is guided and the bent lever runs through its bend position. The piston during its movement in a path is controlled, the shape and dim. of which is predetermined by a link point of a coupling. During the circulation the piston movement along the path is limited by TDC (OT) and BDC (UT), which is defined by the distance of the coupling link point in the dead position angles of the crank.

Description

The invention relates to an orbital jack for combustion in particular engines, the stroke of which rotates eccentrically to the axis of a sun gear Carrier is designed with a crank and a guided for the cylinder in some areas Piston exists, which together with an articulated lever on a mediating the guide Attacks body with the proviso that an action on the movement path of the Piston results, the piston being in driving connection with an oscillating one Gear element stands, and also forms chambers of changing volume, while the piston is guided and the articulated lever passes through its articulated position.

1. State of the art

A transmission of this type is known from PCT application EP 88/00367.

Then a control of the piston movement is proposed, in which the Using split connecting rods instead of the usual one-piece connecting rod is. This is intended to change the kinematics of the machine, with the proviso that the Piston remains much longer in the area of top dead center. This extended Dwell time can be set so that the ignition only takes place when the pressure torque is also exerted on the crankshaft on the piston. By Forces released from combustion represent a gain in usable power. In addition, the combustion process is intercepted more easily, which is particularly useful improved exhaust gas behavior with nitrogen oxides. The occurring mass forces can be difficult due to the special gear kinematics of this machine dominate.

As an in-depth investigation into the application of the known state of the art Technology shows, such gears are still for internal combustion engines unsatisfactory, also because of their limited speed level, MTZ  Motortechnische Zeitschrift 49 (1988) 3. A gearbox equipped with this gearbox was running Machine flawless and soft, but since the construction is relatively large had to be carried out, the maximum speed was inevitably limited to 2300 / min.

2. Task of the invention

The disadvantages described above are to be remedied by further development are, on the one hand the inert masses reduced and on the other hand measures for Efficiency improvement can be created.

Accordingly, the invention is seen in solving the problem, the usable Increase torque while increasing fuel consumption and the amount of exhaust gas reduce and the possible speed level in relation to this and primarily in Compared to self-igniting machines. A flawless and soft operation should be maintained with improvement achieved, whereas that Exhaust behavior should improve above all with NOx. The applicability should thereby be increased that the invention can be constructed accordingly. After all, that is Gearbox according to the invention for work and power machines equally applicable.

 3. The invention

The invention solves this problem by proposing claim 1 and its subclaims.

The innovation achieves this in particular through the appropriate type of crank mechanism construction. The vibrating gear element, which together with a coupling element, the shape of the movement path for the articulation point of the piston primarily specifies, going beyond the shape of the extent of Trajectory is taken into account. Finally, the location of the trajectory also controlled by these elements. The common attitude for those here Interestingly, decisive design leads to a previously unrecognizable Progress.

The one already mentioned, which is decisive for the course of the piston path or its dead center Control of the movement path forms the basis for the performance picture of the invention, according to which not only a single improvement is achieved, but a broad one Spectrum of the result according to the invention. The use of these gear elements,  especially as multi-joints in the form of angle levers, the setting also allows for difficult routes.

As far as reference is made to top dead center (OT), this is equivalent to the Definition of "outer dead center", whereas for "bottom dead center" also the "inner dead center" Dead center "can be selected. It is also necessary to use the terms" swinging Gear element "and" swing arm "to be used synonymously depending on the context.

According to the prior art, the connection mentioned is theoretically one with three Rods composed control joint, in which the four-bar transmission Another fifth articulation point is added eccentrically, which the articulated gear to Five-joint system (here with gearbox degree of freedom = 1) developed. With respect to that Geometric dimensions of the rocker is in the special prior art Connection of the joints and their area of application difficult. Hence the Use of a triangular gear element according to the invention as (angular) Swing arm advantageous because of the restricted swing arm geometry in the special stand technology significantly reduces the possible uses. The swing design also allows an easy way to positively influence the Imbalance or mass pressure behavior of the engine. This applies both with and without special design of the coupling link.

The invention therefore creates the prerequisite for a more differentiated design of the Elements with the objectives as described.

Since the appropriate articulation of the piston to ensure a precise Operating setting, which is sensitive to deformation, acts Use of the swing arm or coupling design according to the invention as multiple joints, especially advantageous in the form of angle levers. This also applies when supporting bodies be provided for stabilization of the transmission, as is the case when for example the piston crown either with a rod or with the triangular tip one firm connection is received, so that the invention use for the purpose of mobility makes of a longitudinal guide of the articulation point (s) or the (swing) board. Latter offers the additional advantage that the articulation points can be adjusted along the edges of the board are, so that the operation of the machine from case to case without changing the board is customizable.

According to the invention, with the appropriate design, the result is an output condition that leads to a significantly increased economy of the machine and above furthermore the reduction of the forces caused by the inertia on the crank or moments achieved.  

In particular, a machine can be designed in which a piston in a cylinder whose axis is the extended connecting line between the axes of rotation of Crankshaft and swing arm cuts at an angle, which at one Crankshaft rotation can preferably run a complete four-stroke process.

Here, the gearbox achieves due to the special gearbox kinematics, in the Starting from the piston crown and with regard to the output, an area more favorable Power transmission function, since the energy is particularly during the engine Expansion clock can advantageously be transferred into mechanical shaft power.

Likewise, you get the opportunity to be very flexible and in particular to the specific process-determining variables, such as pressure and temperature, interactively influence exercise - even during a process.

For example, depending on the design parameters, it allows more or less time Dwell time around the UT a control device that can be timed at a low rate to use the engine (possibly map-oriented) to the most suitable Compression ratio leads. Consequently, there is also the perspective, the maximum keeping the occurring temperature level low increases the thermal efficiency improve and thus come closer to the goal of effectively reducing nitrogen oxides, or to improve the exhaust gas behavior overall.

Furthermore, the comparatively higher piston speed in the work cycle reduces the thermal wall heat losses. The cylinder inclination already indicated is reduced the normal force component occurring on the cylinder wall. This remains from The amount may be higher than that of the special prior art, but the normal wear distribution of the cylinder barrel largely to other causes attributable, especially to high temperatures and pressures, furthermore to Combustion products, material pairing of the sliding partners and their lubrication, while the piston speed is inversely related to its size for the Service life or nominal service life of the cylinder is decisive.

However, the invention also enables the main influencing factors mentioned above Possibilities for improvement, as explained below.

The piston machine according to the invention can be used, for example, as an OTTO or as DIESEL engine work, especially in the latter case, the direct injection too can be done without an antechamber.

You can, in addition to the version with a conventional plunger, like proposed a (possibly additional) piston shape and function of a swivel lend levers. When the same is acted upon, the piston acts as a swivel lever. By applying the pivot lever joint, one comes up with a relatively small one Exertion of force to a remarkably large power conversion within the system. The otherwise  the required straight guidance is not required. In this case, the Leadership of the swivel piston bearing system replaced.

The use of guide bodies is required, regardless of whether one Straight guidance or a positive guidance produced for example by means of gears is provided. This is above all an improved control of individuals Elements of the new gearbox can be executed together.

The invention improves the internal combustion process by making it allowed to "modulate" the engine, i. H. e.g. B. no cylinder deactivation To improve efficiency of other working cylinders, but rather the Use of all cylinders of an engine at lower engine speeds and at high ones mean effective working medium pressures, i.e. better efficiency.

This is achieved by introducing new control parameters, which result from the combination of the orbital lifting gear with a variable charge exchange control device or with a control element that can be selectively tuned to the conventional valve train to let. This results in a further significant reduction in consumption and Pollutant emissions, especially in the partial load range, which is known to be the effective one Efficiency drops sharply because the real cycle process more and more of the ideal process deviates. In addition, the flexibility of the piston machine with respect to it increases Applicability and mode of operation.

As is known, the compression ratio ε is one in this context decisive structural size:

• - The thermal efficiency of the process increases (non-linearly) with increasing ε.
• - The combustion peak temperature increases with increasing ε.
• - The maximum combustion pressure increases with increasing ε.

Because of the dominant importance of the combustion peak temperature for the Nitrogen oxide formation must generally with increasing ε with increasing NOx concentration in the exhaust gas.

Since an increase only after successfully solving specific combustion problems the compression ratio compared to today's values is possible, but especially in the partial load range, the invention provides an application-dependent Spectrum of variation and optimization possibilities available.

Now the time until a certain mixture of air and the fuel to be processed auto-ignite is a physical constant under defined boundary conditions. For an auto-ignition-free operation of a gasoline engine, it is therefore important to avoid the ignition delay time with the time required for the working cycle. For this purpose, either the speed of the engine can be reduced or, and the work of the invention is also concentrated on this, the flame propagation speed can be increased. In addition, the flame paths should be kept short. Since the flame propagation speed can be increased considerably by mixed turbulence, a piston variant according to the invention shows related design features, such as, for example, the squeezing flow realized in a swinging piston, cf. Fig. 2.1f.

However, the most important effect for increasing fuel sales is extraordinarily strong mixture movement, caused by the increasing Approach of the piston to its top dead center between the cylinder head and the Piston base attached squeeze zones pressed cylinder charge.

Distances between the piston and cylinder head can be changed with today's Manufacturing options can also be set to less than 0.2 mm in large series. In addition, the invention provides a device on the swing arm bearing, for. B. in the form of a adjustable eccentric, with which the gearbox can be adjusted if necessary can.

In principle, a certain concentration of the compression volume should be around Spark plug (s) can be reached around, while producing more pronounced Mixture movement.

The increase in the compression ratio has so far led to the more unfavorable In principle, surface / volume ratio increases the concentration unburned hydrocarbons (HC) in the exhaust gas.

According to the invention, however, the measures taken, in particular the combination of a special gear design based on FIG. 1.3 with the piston shape and a sensor system including control device, cause overcompensation and thus counteract the unfavorable trend described above.

In addition, therefore, the motor according to the invention delivers by using the Fuel configuration given the possibility of emaciation significantly lower Hydrocarbon (HC) and carbon monoxide (CO) emissions.

Based on the (not listed here) formal connection to the Calculating the power of a piston engine reveals that a required Driving performance at constant speed through combinations of medium effective pressure and stroke volume can be displayed. For example, the stroke volume reduced, so compensation can be achieved by increasing the Medium pressure.  

In connection with the requirement for a constant compression ratio, there are currently with increasing partial load, however, increasingly flat, disc-shaped Combustion chambers avoided because of the dangers of uncontrolled spontaneous ignition should be. These considerations lead to a long-stroke design of Piston engines, a tendency that is recognizable in new designs and itself can also design according to the invention.

Because the mechanical losses are clearly related to the speed exhibit - the progressive course is determined by the mass forces of the determined oscillating engine parts -, there is a general demand for light Pistons and connecting rods, d. H. also to the smallest possible cylinder units.

The machine according to the invention, in particular according to FIG. 1.3, also solves part of this problem at this point, since it differentiates between the mass of its transmission components on the one hand and their inertial forces involved on the other. The result is not necessarily a machine with lower mass, but rather its reduced inertial mass and its feedback to other engine components.

It is known that the specific fuel consumption at constant speed with decreasing load increases hyperbolic.

As already indicated, the fact that Road vehicles are mainly operated in the partial load range.

Work on the problem outlined above with so-called low-speed concepts want to solve, encounter constructive problems. So the per work cycle is about Increase the amount of heat dissipated from the combustion chamber wall as the speed decreases and thus lead to an increasing loss effect from a certain speed compensated for the described positive effect. The rising leads in the same way Cycle duration of increasing leakage at the piston rings with a negative effect on efficiency. In addition, the reduced mixed turbulence Danger of an uncontrolled spontaneous ignition increases, which is also in Connection with the extended cycle times can lead to knocking problems. About that in addition, these piston engines show a decrease in speed at low loads increasing lability to external disturbances, e.g. B. Turn on an electrical Consumer, which is only possible through appropriately sensitive automatic speed stabilization can be caught.

The task of the invention follows from the complexity of the described relationships, the embodiment variant according to FIG. 1.3 preferably serving as the basis for solving it.

The following is the setting of the invention to the problems mentioned above outlined.

The special state of the art is workable according to the four-stroke principle and as general usual, designed so that the geometric compression and expansion ratio are equal to each other. Therefore, here too, in the case of an OTTO process, that Compression ratio is limited by the knock, which in the Full load operation occurs. This also limits the expansion ratio.

As known from the prior art, an exhaust gas reduces to high Temperature level not only increases the thermal effectiveness, but disadvantageously increases the thermal stress in the cylinder (see also MTZ 49,1988,3).

Compared to the state of the art, the invention, in particular on the basis of FIG. 1.3, in principle allows three different engine configurations:

• 1. The geometric ratio of compression to expansion is equal to one (ε _g = 1) or the suction stroke is equal to the expansion stroke, Fig. 1.3.
• 2. The geometric compression / expansion ratio is greater than one (ε _g <1) or the suction stroke is larger than the expansion stroke.
• 3. The geometric compression / expansion ratio is less than one (ε _g <1) or the suction stroke is smaller than the expansion stroke.

For this and with regard to further features of the invention, reference is made to the drawings. These illustrate the progress of the invention in a fundamental way as well in a special aspect.

The relationships can thus initially be selected geometrically and generally by means of constructive design also constant, e.g. B. 10: 1 or 15: 1.

A control element can now be integrated in the reciprocating piston machines according to the invention, for example a controlled slide in the intake complex of the cylinder, which acts in a control loop-oriented and differentiated manner on the influencing variables outlined in the task objective formulated above, namely particularly also the process sequence. A control element placed in this way can also be installed in conventional reciprocating piston machines, but this leads to disadvantages, as will be shown further below, which the invention avoids by means of the orbital lifting gear based on FIG. 1.3.

The functional mode of operation of the control body is first generalized and described later in connection with a preferred lifting gear according to the invention, without going into structural details for the time being.  

There is therefore the possibility of clocking the control element in such a way that it is timed in the Near the lower piston dead center, or also clearly distant from it, during a each intake stroke is operated to increase the effective compression ratio of an engine change.

Since the time at which it is activated should initially be load-dependent, the motor can for example, run with "pre-compression" or "loading mode". In the case it follows following situation:

In the unregulated case or when the control element is not actuated, the geometric one is then Compression / expansion ratio equal to one, or the expansion stroke is equal to that Intake stroke. Because of the pre-compression, the compression ratio is high "Expansion chamber" (seen in the running direction this corresponds to the gear position the compression stroke) is overcharged or charged, but with the disadvantage that the Temperature during the entire work cycle is at a high level with correspondingly high nitrogen oxide emissions in the exhaust gas. Especially in the case of an OTTO There is also a risk of knocking during the process. To these disadvantages To counteract this, the effective compression / expansion ratio can now be less than One can be set, the amount of the geometric expansion stroke that of geometric intake stroke corresponds. Then the (compression end) pressure or (Compression end) temperature can be regulated to an uncritical level.

In practice, the control element can, for example, by an actuator on the basis of a Sensor signal regulated or retrieved from a map memory device Signals are manipulated, for example by early actuation of the control unit, but especially in front of an inlet valve, so that in this case the control element effective suction stroke reduced. The disadvantage here is that the process is at a cost reduced charge expires.

These disadvantages are prevented according to the invention by a suitable choice of the transmission parameters, in particular by the fact that the geometric compression / expansion ratio is greater than one, or the suction stroke is greater than the expansion stroke, cf. see Fig. 1.3.

To this end, the functional mode of operation of the invention is described below when used described with a control and regulation device, without first referring to structural details enter into.

As already described, the compression or expansion ratios are first geometrically preselectable and then generally through the design permanent, e.g. B. 10: 1 or 15: 1.

The reciprocating piston transmission according to the invention, in particular on the basis of FIG. 1.3, can first be fixed to a specific requirement, for example to the geometric compression ratio of 15: 1, that is to say significantly higher than in conventional OTTO engines (average about 1: 9) that in the case of a geometric expansion ratio, the compression / expansion ratio in the aforementioned case (application-oriented) is 1.5, corresponding to the "Case No. Two: geometric compression / expansion ratio is <1" mentioned above.

There is therefore the possibility, as already shown, of clocking the control element so that it temporally close to the lower piston dead center, or also far away from it, is operated during each intake stroke to adjust the compression ratio of the Motors to adapt to the expansion ratio.

Since the time at which it is activated should be load-dependent, the motor can for example, run with "pre-compression" or "loading mode", so that it works accordingly the following situation results:

In the unregulated case or when the control element is not actuated, the geometric compression / expansion ratio is then greater than one, or the expansion stroke is significantly smaller than the suction stroke. The compression ratio is high, the "expansion chamber" (seen in the direction of travel this corresponds to the gear position after the compression stroke, Fig. 1.3) is overcharged or charged to a certain extent and the engine can be described in a similar way to a conventional pre-compressed engine, but in the regulated case with the advantage that the temperature can run at a significantly lower level during the entire work cycle, without additional measures such. B. Charge air cooling. As a consequence, this leads to a lower thermal load on the components and thus also to a reduction in the nitrogen oxides contained in the exhaust gas, since this also lowers the process or peak temperature. With this control and regulating device, the tendency to knock is also acutely and favorably influenced (as will be shown below), so that therefore the embodiment variant according to the invention, in principle, favors its use in gas engine technology, since the lower compression end temperature has real advantages in terms of Knocking brings with it and therefore higher performance can be driven. For the special application, the reason for the more favorable temperature level in the compression cycle is explained below.

The motor according to the invention can be particularly advantageous in critical operating states are operated, especially when there is a risk of knocking, in which the following are then in principle Processes run:

In the regulated case, the control body becomes effective on a process-dependent basis Compression / expansion ratio set to less than one (the geometric The compression / expansion ratio remains unaffected by this and remains greater than One). The effective suction stroke becomes more than the expansion stroke  or less shortened. The compression ratio is reduced, to put it simply, it runs Engine in the partial load range, which mainly occurs in motor vehicles.

The control body is z. B. by an actuator based on a sensor signal regulated or from a signal retrieved from a map memory device manipulated. This regulates the compression ratio in suitable areas, in this case, the control unit in particular determines the effective intake stroke.

Of course, the versatile possibility of exerting influence also allows this Set the compression ratio to the same value as the expansion ratio or approaching the area where a knock is likely to occur or actually sets (knock sensor system required!). The control loop to be provided is then able to feed back the compression ratio via actuators, i.e. specifically to reduce.

If the engine works under partial load, the control element can be influenced in such a way that the compression ratio is reduced, thereby preventing knocking.

As a result of the closed control loop, the control device can adjust the compression ratio increase again if necessary. So you are able to use integrated sensors and Actuators individually adjust the optimal compression ratio, so good ones To achieve process states just before knocking. This will increase the efficiency of the Reciprocating engine according to the invention further improved.

It goes without saying that the possibilities of variation of the facts described above cannot be dealt with exhaustively and that there are other useful combination possibilities inherent in the system, especially in connection with the gear on the basis of FIG. 1.3.

The desired goals are achieved and the present task solved according to the characterizing features of the claims.

The mode of operation of the invention is now once again described in a more general form and subsequently with reference to the idealized pressure-volume diagram (pvD) of a 4-stroke OTTO process according to FIGS. 9.1 to 9.4, for example with the configuration " Geometric compression / expansion ratio <1 ", or suction stroke (with UT2 in Fig. 1.3) greater than expansion stroke (with UT1 in Fig. 1.3).

If the engine is operating at full load, the control element (as described above) is adjusted to a minimum. Because the compression ratio is set too high under these conditions, combustion knocking is initially unavoidable when the engine is started. However, the sensor system provided detects knocking immediately as an input signal in the control loop. The behavior of the control element is then determined by means of the complementary output signal, with the further objective of ending the intake or intake process in the next moment even before the UT (especially UT2 in FIG. 1.3).

Fig. 9.1 shows the beginning of the suction process at point 0, i.e. in the TDC position, cf. Fig. 1.3. This process ends (initially unregulated) at bottom dead center, point 6, gear position UT2 in Fig. 1.3. This is where the compression stroke begins. If the latter is continued conventionally, the dotted curve results. From the upper reversal point of the compression stroke, the mixture, in particular an air / fuel mixture, is (adiabatically) compressed in accordance with the end pressure of point 2. High pressure and high temperature then result in a combustion knock. As described above, the knock is promptly detected by the built-in sensor technology. In the control loop, the control element is operated, which shifts the suction process to point 1 during the suction stroke, with the consequence that the pressure generated in the bottom stroke of the suction stroke (gearbox position UT2 in Fig. 1.3) drops and corresponds to that of point 6 '. The volume inside the cylinder increases or the mixture expands (adiabatically). In the further course, the pressure falls below that of the external atmosphere at point 6 '(UT of the intake stroke), combined with a simultaneous drop in temperature.

The now following compression stroke begins in point 6 ', and in point 1 it is the same Pressure to that of the external atmosphere combined with a simultaneous one Temperature rise. Since the compression stroke mainly starts at point 1 and at TDC at point 2 'is completed, it follows above all that the Compression ratio is reduced because the final compression pressure at point 2 ′ is reduced compared to the compression end pressure in point 2. Since the Compression end temperature is also lower, knock is also suppressed.

Conventionally, the process delivers as useful energy the area enclosed by lines 1-2'-3-4-1 in Fig. 9.1. Inventive motor on the basis of Fig. 1.3 may have a 'extended to the trace 4-4 effective expansion stroke. Then the useful energy corresponds to the area delimited by lines 1-2'3-4'6-1 and is larger than that of the equivalent area 1-4-4'-6 (hatched in Fig. 9.1 or 9.2) with the conventional engine. Provided that the mass of the fuel consumed, which is introduced in particular within points 0 and 1, is unchanged, the useful work of the machine is increased again and also the thermal efficiency. Furthermore, the exhaust gas temperature is lower, so that important engine components are thermally less stressed.

In the case of a pre-compressing engine, combustion knocking can occur earlier due to the higher pressure in the initial phase of the compression stroke. The control loop then further reduces the effective compression ratio accordingly. Point 1 in Fig. 9.1 consequently shifts more to the left. The thermal output is not reduced here. The exhaust gas temperature does not rise either, since the comparatively increased level of the expansion ratio is maintained with a reduced effective compression ratio.

Fig. 9.2 illustrates these relationships in the idealized temperature-entropy diagram (T, sD): The thermal efficiency increases analogously to the shaded area. The increased useful energy results from the reduced heat loss. The marked points correspond to those of the p, vD, Fig. 9.1.

In order to achieve an improvement in efficiency, certain operating states are required of the engine an increase in the compression ratio, especially in part-load operation of the motor. In this state, the wall temperature of the Combustion chamber comparatively low, so that the air volume and useful work can be increased by increasing the control ratio of the compression ratio.

In the previously described process sequences, compression begins at the expense of a reduced charge mass if the geometric relationships of compression and expansion are the same and the motor is regulated in the manner described. This is the case with the general and special prior art. However, the invention can (over) compensate for this disadvantage with the gear configuration initially shown on the basis of FIG. 1.3. This is a particular advantage of the invention, since on the other hand this means that with an increasing load an (adapted overloading) of the expansion chamber becomes possible without further auxiliary devices.

Subsequently, the partial-load operation of the engine according to the invention is developed on the basis of FIG. 1.3, again with the configuration "geometric compression / expansion ratio <1" or intake stroke (with UT2 in FIG. 1.3) greater than expansion stroke (with UT1 in Fig. 1.3), (see here the idealized p, vD according to Fig. 9.3, or T, sD according to Fig. 9.4).

For comparison purposes, a corresponding process for the status of the Technology outlined.

Conventionally, the cylinder filling is carried out in the intake stroke by means of an air quantity regulation, so that the cylinder pressure is now marked by point 01 and point 11 marks the end of the suction stroke. For the reasons already mentioned, the temperature of the intake air is reduced at the same time. At the same time, a charge acceleration occurs which corresponds to the work share of the area delimited by the 00-01-11-1 line or which can also be expressed by the differential pressure of points 00 and 01. In this process, however, the movement of the air dissipates, so that the temperature of the air reaches an almost atmospheric level. The compression stroke now begins at point 11, line 00-1 again representing a measure of the quantity of mixture supplied. At point 22 (position OT in Fig. 1.3) the compression ratio and the compression end temperature correspond to the corresponding full load values, but with a comparatively reduced density and lower combustion speed (represented in pvD by lines 22-33-44). Lines 33-33'-44-33 enclose the (hatched) loss area, corresponding to a low thermal efficiency.

Now, as already described, the cylinder pressure can be regulated to point 02 during the intake stroke according to the invention. At point 111 within the intake stroke, above all the control element ends the intake process, nevertheless the piston ( 4 ) moves further in the direction of the UT. Again, the pressure and temperature of the mixture are reduced (adiabatic expansion). Point 1111 represents the UT in the intake stroke (UT2 in Fig. 1.3). This is immediately followed by the compression stroke along point 111. For reasons already mentioned, the temperature of the mixture at point 111 essentially corresponds to that of point 11, that is to say the atmospheric temperature. It can be seen that the line 111-22, which corresponds to the compression stroke of the engine according to the invention, is longer than the line 11-22, which represents the comparable state in the prior art. The compression ratio of the new engine can thus be increased again. At point 22 (OT) the temperature of the mixture is raised. If the sensors do not detect knocking in this state, the effective compression stroke is automatically extended and the compression ratio increases. In the event that the sensor system registers knocking in the further course of the process, the system regulates the compression ratio to a lower level. Therefore, the combustion speed increases to the maximum possible values, which creates more favorable conditions for the subsequent combustion process. The improvement corresponds to the line 33-33'-44 in p, vD, Fig. 9.3 and dto. In Ts-D, Fig. 9.4. For the expansion stroke essentially apply the relationships shown above, so that the effective gain of work corresponds to the difference area, which is formed from the lines 1-22-33'-4'-6-1 and 00-02-111-1-00 . The latter is essentially a measure of the loss in air volume regulation.

Otherwise, with the engine according to the invention, especially on the basis of FIG. 1.3, OTTO processes or conditions can be realized which are at least close to those in the diesel process. Then more useful work can be gained than with engines that run on the last-mentioned principle.

Because this (four-stroke) engine according to the invention has a significantly smaller one  Allows combustion pressure without loss of performance or deterioration in efficiency, can then with a reduced weight percentage, also the specific Fuel consumption decrease compared to that of diesel engines. Because the latter Removal of (soot) particles z. Z. causes even greater technical difficulties, however, the special motor according to the invention in thermal efficiency Diesel engines is superior, the invention can not only current Comply with emissions regulations, but also offers reserves regarding future tightening in the areas of consumption and exhaust gas legislation.

In the following, the invention is specified in application to a control device and orbital lifting gear, FIG. 10, with particular reference to the structure according to the invention on the basis of FIG. 1.3.

The knock limit control provided on a cylinder-selective basis can, for example, by means of Structure-borne noise sensors or non-modulated or frequency-modulated acceleration sensors are made. Here, the knock sensor signals of a certain Frequency and bandwidth filtered and integrated. They correspond to a value ("Knock sensor integral"), which in this case is a measure of the corresponding Represents combustion noise. The measurement window position or the filtering is seen as a result their special coordination in the event of knocking combustion is the greatest possible Signal increase before. The influence of harmful frequencies or noise should insofar be largely switched off. A self-adaptation of this system cylinder or engine-specific frequency ratios is achieved in that from the cylinder-selective knock sensor integral a reference level is constantly updated. From the The system then learns the normal frequency level of the cylinder adaptively, for example significant changes in operating status, to react or between knocking and to distinguish knock-free combustion. The special detection circuit or The measured variables are evaluated using electronics or logic. This processed the knock sensor signals and operated when necessary, for example when knocking Combustion, according to a predetermined algorithm that in previously discussed topic mentioned actuator or control member. As already described, the latter exerts its influence specifically in or on the intake tract of the engine and can, for example, from a Slide mechanism or a rotating flap in an intake duct. The displacement or the rotation can be done by a sensor or electronically controlled actuator happen.

Depending on the engine status, different program sequences can be run, whereby For example, even complex signals within complicated maps as required the electronic control systems are implemented or processed. It can  simply first on previously defined and saved (pre-) optimized Setpoint maps are used. For example, if the target function is a Consumption optimization in the partial load range of the engine corresponds to the manipulated variable in Control loop influencing the control element in or on the intake tract. The system can So either adjust to a (map) predetermined value or individually Find the most favorable operating state alternately in small steps, one Phase shift are taken into account by the response time of the control member can. A certain optimum can thus be found quickly. The control process is now ended, and the manipulated variable can remain constant until a changed one Engine operating state is present or is desired.

This knock control can be very advantageous in combination with a control device Influencing the special orbital gear geometry can be used. You get thereby an expanded range of applications, because of the, comparatively sluggish and slowly proceeding geometry change on the one hand only a relatively rough specification of the achieve desired ratios and then fine-tune the control algorithm described.

The actual geometric compression / expansion ratio can vary from one Control device to be changed. This is particularly the bearing of the rocker relocated locally. This can be done, for example, by means of a servomotor, which is a Eccentric or a worm gear moved in the desired manner. The Subsequent fixation of the gear is done for example by force and / or Positive locking. Based on this, d. H. of the changed and practically achieved Transmission scheme, i.e. the actual geometric compression / expansion ratio, the compression ratio to the necessary level adjusted, as well as the compression temperature.

As already indicated at the beginning, according to the invention there is the possibility of integrating the rocker into the piston. Then, at point ( 4.5 ), instead of the (plunging) piston ( 4 ) in Fig. 1.3, a (swivel) piston (for example according to Fig. 2.3) can be articulated.

This gear / piston combination then contains the one described above Favorable power transmission function also the previously described possibilities of complex engine management.

The design of the (swivel) piston takes into account not only sufficient mechanical strength in the support cross-sections and in the pin bosses as well as sufficient resistance to thermal fatigue in the area of the combustion chamber bowl, but also the necessary resistance to wear, especially in the first sealing groove: For example, piston profile variants according to the invention enable ( Fig 6 ff) adequate Expansions of. between the piston head and pin bore to prevent to the "stress cracking". In addition, the ratios of (fictitious) bowl diameters to equivalent (fictitious) piston diameters are related to each other in order to limit the stress conditions within the piston.

Against thermal fatigue in the area of the combustion chamber trough or to avoid too High local temperature peaks can, for example, affect the general temperature level lowered, or a special piston cooling (injection cooling) may be provided.

In addition, as already described, using special piston contour shaping directly influence the individual phases before, during and after the Taken incineration.

Because in conventional reciprocating piston engines due to the ignition delay (the combustion begins the pressure of the piston before reaching the top dead center position) Counteracts combustion gases in the initial phase of piston movement negative starting torque on the crank.

In the "special prior art" this "ignition delay" is shortened to the extent that the Piston lingers longer in the area of its top dead center position, the combustion can start at top dead center. This process can be compared to conventional reciprocating piston machines and on the piston side as "quasi-static" describe. The negative starting torque on the crank is here due to the special Gearbox constellation reduced.

According to the invention, the above-described negative influence with regard to the initial torque is said to occur the crank can also be reduced. The energy of the combustion gases should hardly or do not counteract the piston movement at all.

One means of doing this is above all the specially designed ignition and expansion cycle based on, for example, FIG. 2.1ff or also FIG. 4.1. The result should be a shortening of the flame core formation, combined with its faster enlargement during the ignition phase and a faster charge distribution over the (piston) active surfaces. Instead of a "quasi-static" process around the OT ("special prior art"), the process of mixture preparation, ignition and expansion is structured "quasi-dynamically" according to the invention.

An increased warming of the engine does not take place because of the constructive Design of the engine and its management prevents this. Then they point Exhaust gas products also result in a reduced pollutant content.

In addition, the ignition pressure relieves the piston bearing ( 4.3 ) in the initial phase due to the design of the application surface ( 4.1 ) according to the invention (counteracts the centrifugal piston force).

The usually occurring negative torque as a result of the early ignition is suppressed or mitigated according to the invention in particular by the following measures in particular:
When considering what is happening at the same time, both in terms of space and time, especially during the ignition (delay) phase, the primary area of the piston surface ( 4.1 ) - or trough sector ( 41 ) - is essentially concentric due to its piston axis of rotation ( 4.3 ) Contour no harmful torque on the crank.

A favorable design of secondary or ignition chamber section (s) and their connecting transition area can reduce premature overflow into the main combustion chamber during the ignition delay ("space-time solution"), as well as an earlier ignition setting, more favorable burning of the mixture charge and enable their subsequent complete combustion (see, for example, Fig. 2.1 and 2.2).

However, a pronounced one arises when it flows into the main combustion chamber Squeezing flow, so that as a result of the considerably increased spreading of the flame speed also not with a tear of the near-surface flame front on the piston is to be expected (otherwise leads to increased HC formation).

In conclusion, the reciprocating piston machine according to the invention, in particular based on FIG. 1.3, can achieve not only the greater degree of mechanical conversion but also a higher effective thermal efficiency, especially under partial load. The (cold) starting behavior of the engine can also be improved by the engine management system according to the invention. In addition, the amount of exhaust gas is significantly reduced.

Basic details of the invention are explained using the following exemplary embodiment, FIG. 1:

With ( 1 ) the axis of rotation of the output or generally the center of the sun gear is designated, on the eccentrically arranged pin ( 2 ) of which the lower connecting rod eye ( 10.2 ) of a one-piece ( 10.3 ) or multi-piece ( 10.4 ) connecting rod ( 10 ) acts. This crankshaft or eccentric shaft - the preferred direction of rotation is indicated by an arrow - is mounted ( 1.1 ) in a housing ( 30 ) (not shown here). The rocker ( 7 ), in its special design as an angle rocker ( 7.9 ), has three pivot points here. The angle rocker is rotatably mounted in point ( 7.4 ), in particular on the motor housing, and is stationary or also for the purpose of gear adjustment with (not shown) means ( 91 ), e.g. B. an eccentric gear, on and lockable. Depending on the gearbox design, articulation points ( 7.5 , 7.6 ) are located at the same or different radial distance from the rocker bearing ( 7.4 ). As a result of the central and essentially triangular gear body formed in this way, two goals are fundamentally achieved: on the one hand, a stress concentration that is harmful to the structure (with only a single articulation point) is largely avoided or its local concentration is rectified, and on the other hand the flexibility in terms of applicability is effectively improved by increasing gear design parameters. In this embodiment, at point ( 7.5 ), "rocker arm bearing near the piston", either another gear link is articulated, for example a lever or a rod ( Fig. 1.3), or as shown in Fig. 1, the shaft of a (plunger) piston . The force transmission elements or guide-imparting bodies ( 6 ) at the articulation points are generally articulated bolts, for example a conventional piston pin ( 6.1 ), or, according to the embodiment variant in Fig. 1.4 or Fig. 1.5, a multi-articulated transmission element, especially a three-articulated body. The upper connecting rod eye ( 10.1 ) of the one-piece ( 10.3 ) or multi-part ( 10.4 ) connecting rod ( 10 ) then rotatably engages at point ( 7.6 ) of the rocker. At top dead center (TDC) the connecting lines between the rocker bearing ( 7.4 ) and the articulation point ( 7.6 ) on the one hand and crank pin ( 2 ) and - shaft bearing ( 1.1 ) on the other, are at an (articulated) angle ( 5.1 ) to each other and characterize here in their functional sequence the gear element "articulated lever" ( 5 ).

In the example shown, the path ( 8 ) of the piston-side articulation point ( 4.5 ) corresponds to the path of the swing-side articulation point ( 7.5 , 7.6 ): in this case, said path ( 8 ) is circular in shape with stretched positions in the reversal points ( 7.7 , 7.8 ); the piston is guided in the cylinder ( 3 ) with a stroke, the amount of which depends on the offset ( 3.2 ) of the cylinder axis ( 3.1 ) and the angle of inclination ( 3.3 ) and ( 4.8 ). If necessary, counterweights ( 20 ) can be provided as a mass balance, which ensure a smoother running of the engine depending on the application.

For the sake of clarity, the presentation of fuel supply or external mixture preparation systems as well as ignition means and charge exchange elements waived. Conventional systems or valve train (s) can be installed here.

In the case of two-stroke operation of the engine, FIG. 1.1 shows the gear position after the expansion or in the gas exchange stroke. Fig. 1.2 illustrates the same machine essentially after the gas change and during the compression stroke.

The functional group described in the preceding paragraphs is the basic inventive idea of the preferred embodiments of the transmission according to Fig. 1.3 and Fig. 1.4. In particular, these gearbox constellations, the central components of which represent the swing arm and coupling which can be designed in a variety of ways, open up the invention to versatility, not least because of the possible combinations with the piston variants outlined.

Reference list

1 = sun gear (center point) or output
1.1 = crankshaft or eccentric shaft bearing
2 = crank (pin) or eccentric
3 = cylinder
3.1 = cylinder axis (plunger)
3.2 = cylinder axis offset with respect to 7.5 (plunger)
3.3 = resulting angle from 3.2 (plunger)
3.5 = flow influencing element
4 = piston (general)
4.0 = piston crown
4.1 = (essentially concentric) pressure surface of the (swivel) piston
4.2 = working surface of the (swivel) piston angled to 4.1
4.3 = axis of rotation of the (swivel) piston
4.4 = (projected piston side) knuckle surface
4.5 = articulation point on the piston side (usually axis: piston pin)
4.6 = sealing system / limit (piston side): medium gas
4.6.1 = sealing limit in the neck area of the piston
4.6.2 = sealing limit in the front area of the piston
4.6.3 = sealing limit in the side area of the piston
4.7 = (fluid) sealing system / limit (piston side): medium oil
4.8 = angle between piston crown and stem
4.9 = piston shaft or rigid piston rod
5 = articulated lever
5.1 = determining kink angle
6 = leadership mediating body / paddock
6.1 = piston pin
6.2 = connecting element
6.3 = straight guide element
6.4 = eccentric element
7 = swinging gear element / swing arm
7.1 = section of the swing arm
7.2 = the section of the rocker which is essentially parallel to 7.1
7.3 = connecting bridge
7.4 = swing arm mounting (swing arm side)
7.5 = pivot point near the piston
7.6 = pivot point remote from the piston
7.7 = upper extended position of the joint point near the piston
7.8 = lower stretch of the hinge point near the piston
7.9 = angle rocker / lever
7.9.1 = swing arm angle
8 = (motion) path
10 = connecting rod
10.1 = upper connecting rod eye
10.2 = lower connecting rod eye
10.3 = one-piece connecting rod (one-piece connecting rod)
10.4 = multi-part connecting rod (multi-part connecting rod)
11 = (maximum) main combustion chamber volume
12 = (mixture) ignition range
13 = overflow area
14 = flame or flame front (s)
14.1 = elliptical or paraboloidal flame or flame front (s)
20 = mass balance
30 = housing
31 = swing arm bearing on the housing side
41 = (peripheral) trough of the applied area ( 4.1 )
42 = (peripheral) trough of the exposure surface ( 4.2 )
50 = injection system: pilot nozzle ( 50 )
51 = injection system: main nozzle ( 51 )
60 = charge change organ, e.g. B. Inlet valve ( 60 ): mushroom valve
61 = charge change organ, e.g. B. Exhaust valve ( 61 ): rotary slide control or slot control
70 = gearbox parameters, general
71 = geometric compression / expansion ratio ε _g
80 = control device, general
81 = knock limit control device
90 = device for changing position or position, general
91 = Device for changing the position or position of the swing arm

Figure description (Fig. 1 to Fig. 10)

Fig. 1: Schematic gear arrangement, general, unspecific version, in TDC

Fig. 1.1: Like Fig. 1, but in UT

Fig. 1.2: Like Fig. 1, but approx. 215 ° after TDC

Fig. 1.3: Modified embodiment of Fig. 1, in subtitles

Fig. 1.4: Modified design variant of Fig. 1, approx. 30 ° from top or from top

Fig. 1.5: Schematic configuration of the transmission diagram according to Fig. 1.4

Fig. 2: Variants of the piston-lifting gear according to the invention:

Fig. 2.1: as a swing piston system, in the TDC position

Fig. 2.2: Like Fig. 2.1, but with details in the front piston area

Fig. 2.3: Swinging piston variant, single-chamber version

Fig. 2.4: Swinging piston variant, two-chamber version

Fig. 3: Simplified, comparative illustration of combustion processes:

Fig.3.1 : Conventional plunger in a cylindrical cylinder

Fig. 3.2: qualitative spatial flame propagation in a (swivel) piston system according to the invention

Fig. 3.3: variant of Figure 3.2.

Fig. 3.4: Local cross-sectional representation of the flame contour

Fig. 3.5: Schematic combustion chamber volume geometry

Fig. 4: Views of two variants of the reciprocating piston transmission according to the invention

Fig. 4.1: oscillating-piston system as a (quasi -) "long-stroke"

Fig. 4.2: The oscillating piston system as a (quasi -) "short-stroke"

Fig. 5: Comparison: machine of the invention / the "prior art"
Illustrate in this regard:
Solid lines = lifting gear according to the invention
Dash line = machine (s) according to WO 88/08922 or EP 88/00367
Dot line = conventional lifting gear

Fig. 5.1: Representation of the piston acceleration (ak), qualitative

Fig. 5.2: Tangential force curve or torque curve (Mi) on the crank with constant pressure on the piston

Fig. 5.3: the inertial forces _(m M) influence on the torque on the crank or rotational force line

Fig. 5.4: Torque line (M _i ) like Fig. 5.2, but for a complete four-stroke cycle

Fig. 6: ( Fig. 6.1 to Fig. 6.12) design variants, in particular of pistons, based on the orbital lifting gear according to the invention

Fig. 7: ( Fig. 7.1 to Fig. 7.6) design variants, in particular of gear elements, based on the orbital lifting gear according to the invention

Fig. 8: Constructional design of a special orbital lifting gear

Fig. 9 or Fig 9.1 to 9.4 Fig.:. Pressure, volume diagrams (p, VD) or temperature, entropy diagram (T, SD) for machine according to the invention on the basis of Fig 1.3.

Fig. 10: Schematic representation of the control and regulation system.

To further illustrate the invention, reference is made to the figures.

Fig. 1: General, unspecific design, ie non-optimized kinematic dimensions or gear parameters (e.g. reference numerals ( 4.8 ), ( 4.9 ) or ( 5.1 )). Swing arm ( 7 ) designed as an "angle swing arm " ( 7.9 ), here with swing arm angle ( 7.9.1 ) approx. 90 °; The reason for this: "equalization" of the concentrated flow of force. Piston ( 4 ) articulated at the articulation point ( 7.5 ) near the piston, connecting rod ( 10 ) articulated at the articulation point ( 7.6 ) remote from the piston. From this plunger version, an argumentative modification to the swing piston system follows (e.g. Fig. 2ff). A more general representation (without figure) would be a "straight line" in the output, different radii with regard to the (swing) articulation points (raceway radii) ( 7.5 ) and ( 7.6 ) as well as differentiated designs of the articulation points for pistons ( 4.5 ) and connecting rods ( 10.1 ) as well as the swing arm mounting ( 7.4 ).

Fig. 1.1 and Fig. 1.2: outline a principle of operation. In the case of two-stroke operation of the engine, FIG. 1.1 shows the gear position after the expansion or in the gas exchange stroke.

Fig. 1.2 illustrates the same machine essentially after the gas change and during the compression stroke. The preferred direction of rotation is marked with an arrow.

Fig. 1.3: shows a further development of the functional group according to the invention on the basis of Fig. 1. A preferred embodiment of this transmission opens up a variety of uses for the invention, as will be shown in the description. The reference numerals essentially correspond to those in FIG. 1.

Fig. 1.4: represents an alternative transmission scheme according to the invention on the basis of Fig. 1. The coupling ( 6 ) is designed here as a three-joint link and can be driven on a circular path. It connects the crank to the connecting rod and, together with the rocker degenerated into a crank, determines the orbital data of the piston. In the case of special geometric and kinematic parameters, this transmission variant achieves a correlation between maximum gas force in the cylinder and maximum torque on the crank and can also be implemented in practice. The longer dwell time in bottom dead center that occurs in this case can advantageously be used for the gas exchange processes during two-stroke operation of the machine. With certain gear parameters, irregularities in the rotational movement of the output shaft can also be reduced. The reference numerals essentially correspond to those in FIG. 1.

Fig. 1.5: shows an example of a possible physical-technical design or structural design of the schematic transmission of Fig. 1.4, with position in the TDC. The gear elements are drawn rotated to each other for better understanding. The coupling ( 6 ) here has three rotary joints, the axes of which are offset eccentrically and spatially from one another. For reasons of disassembly / assembly, the coupling can be expediently made in two parts and can be coupled, for example, along the axis of symmetry of FIG . The latter is possible here - in contrast to conventionally built crankshaft segments - because operational bending moments in this gear section are limited to a minimum by the triple bearing ( 1.1 , 7.4 ) of the coupling ( 6 ). A complex axial securing against displacement can be omitted. The connection therefore essentially has the task of preventing a relative rotation of the transmission components. In certain cases, a synchronization of the two output shaft segments may also be appropriate, for example by means of a gear transmission. If necessary, the coupling can have a mass balance ( 20 ) in the form of a recess. For the reasons mentioned above, the housing can also be in two parts on the line of symmetry, while in this case the connecting rod consists of two identical sub-segments and these are only indirectly connected to one another, each being designed in one piece. With this "double connecting rod machine", the one-piece connecting rods are simply pushed onto the coupling joints during assembly. The usual screw connection of the connecting rod halves is omitted. The articulation point 4.5 on the piston side is located near the center of the piston and not, as is customary, in the peripheral peripheral zone of the piston. This on the one hand homogenizes the stress distribution within the piston and on the other hand creates a favorable prerequisite for the design of (piston) troughs and / or (valve) pockets. If necessary, the output shaft ( 1 ) can be equipped with heavy metal inserts - as a mass balance ( 20 ) - and / or have counterweights (not shown). Both plain and roller bearings can be used to support the individual gear components, with or without oil lubrication. The bearings do not have to be divided under any circumstances. Overall, the transmission is also characterized by simple (dis) assembly and is also inexpensive to manufacture. The economy also increases because there is no need for an expensive crankshaft.

This transmission can also be used advantageously in a piston compressor, for example will. This reduces what is necessary for the operation of the compressor Drive power clearly.

Fig. 2.1: outlines a design implementation of the articulated lever ( 5 ) and the piston-integrated rocker ( 7 ). The angled surfaces ( 4.1 ) and ( 4.2 ) can be seen, with a relative angularity of almost 90 ° here. The (conventional) connecting rod ( 10.4 ) is split on the lower connecting rod eye ( 10.2 ), the output is via a crankshaft ( 1 ) or ( 1.1 ). A peripheral trough ( 41 ) is available for the provision of a space which favors the mixture ignition and permits an intensive and directed (squeezing) flow in the direction of the ignition means or enables charge stratification. Furthermore, a schematic sealing limit ( 4.6 ) on the gas and piston side is shown.

Fig. 2.2: The trough ( 41 ) in the front piston area ( 4.1 ) intensifies charge mixing on the one hand over a larger crank angle range and, in conjunction with an additional trough ( 42 ) in the piston area ( 4.2 ) on the web side, enables a variant of the charge distribution or stratification. Ignition means (not shown) (e.g. spark plugs) are located in the cylinder ( 3 ) or in the engine block. The following injection scheme is also shown: A pilot nozzle ( 50 ) in connection with (not shown) ignition means and a main nozzle ( 51 ) achieve either sequential or simultaneous injection for the purpose of an optimized ignition and combustion process. An overflow area ( 13 ) following or overlapping with the ignition area ( 12 ) is advantageous.

As an example of the (schematically represented) charge change, in particular for the 2-stroke operation of the engine, a direct current flushing with an inlet mechanism ( 60 ), in particular a mushroom valve and an outlet mechanism ( 61 ), especially a rotary slide valve control or slot control is provided. This results in an asymmetrical control diagram (see corresponding arrow directions for inflow and outflow and charge flow). A (fluid) sealing limit ( 4.7 ), in particular an oil sealing system, made in one or more parts, causes "deliberate" leaks to the raceway or sealing body lubrication at the butt corners. The schematically indicated elastic element ( 4.7 ), in particular a mechanical spring, ensures the necessary contact pressure, especially during the starting process. An additional element ( 3.5 ) in the contour, for example concave, convex or both, in particular swirl generator or turbulence and vortex generator or also tear-off contour / edge, shapes the flow in the chamber according to the principles described above. It can also serve as an inlet slot in 2-stroke operation (symmetrical control diagram).

Fig. 2.3: shows a single-chamber swivel piston. The (piston-side) rocker bearing is provided with a through hole, if necessary, to reduce weight. The material is e.g. B. aluminum injection molding or an AL alloy. Ceramic inserts can be located in the piston face area, which allow a favorable influence on the heat transfer behavior (e.g. insulation). In addition, sealing strip carriers ("armor" or reinforcement of the sealing limit) can be made by comparatively harder materials such. B. cast "austenitic cast iron", resulting in an increase in wear resistance.

Fig. 2.4: shows a two-chamber swivel piston, the second chamber ensuring in particular a "clean" pre-compression or fresh charge supply. Reinforced articulation points ( 4.5 ) or connecting bars ( 7.3 ) result in structural stiffening of the piston.

Fig. 3.1: exemplarily and qualitatively outlines the spatial flame spread of a plunger engine, with a lateral ignition area ( 12 ) in the TDC environment, at three different times or crank degrees. The main combustion chamber is essentially formed by the piston ( 4 ) (bottom, 4.0 ) and the cylinder ( 3 ) and in this case is essentially cylindrical. It can be seen that, in the course of the combustion cycle, an increasingly larger proportion of the cylinder head surface is successively recorded. However, the flame front ( 14 ) must also increasingly follow the piston moving away from the TDC with increasing speed in the Z direction and into the depth of the combustion chamber. There is a risk here that more or less large zones form directly above the piston, particularly towards the end of the flame, which zones can no longer (more) be reached from the flame front ("quench zone formation"). Quench zone volumes and HC formation are directly related here, ie with increasing quench zone volumes, the proportion of unburned hydrocarbon also increases. At high air ratios, the flame may go out prematurely and intensify this process, although the quench zones primarily contain a fresh mixture of fuel and air.
Flame contour: essentially spherical.

Fig. 3.2: The spatial flame propagation of a (swivel) piston designed according to the invention with ignition area at the level of the peripheral trough ( 41 ) is shown qualitatively. The (temporally variable) main combustion chamber contour is essentially approximated by the two admission surfaces ( 4.1 ) and ( 4.2 ), which are angled to one another here, and also by the cylinder ( 3 ) (with its side walls). The maximum combustion chamber volume ( 11 ) spanned in this way is shown in dash-dot lines in FIG. 3.5 and in this case results approximately in a cuboid.

Usually the burning rate is that ignited by the ignition means (candle) Mixtures much lower than the gas velocity. The latter is according to the invention with flow-giving contours increased (especially through those described above Measures), which are larger than conventional measures Crank angle should enforce the advantageous flow conditions. Developed by this the flame is not circular or spherical, but spreads out ellipsoidal or ellipsoidal or paraboloidal, i.e. conforming to the contour Chamber geometry. This results in a faster ignition, increased  Flame spreading speed, shorter burn times, displacement of the Combustion or flame center of gravity in areas of more favorable space / pressure conditions and thus increased torque. In addition, the heat loss is over the combustion chamber walls lowered.

Fig. 3.3: like Fig. 3.2, but with a different flow and / or "single ignition".

Fig. 3.4: Cross-sectional view: shows ellipsoid-like or ellipsoidal or paraboloid-shaped flame contour, conforming to the combustion chamber geometry. Among other things, this results in the faster ignition described above and increased flame propagation speed. The spatial flame contour spread is "quasi two-dimensional", since one of the three dimensions becomes comparatively smaller in spatial-temporal relation.

Fig. 3.5: (Swivel) piston designed according to the invention with the main combustion chamber contour formed essentially by the two angled acting surfaces ( 4.1 ), ( 4.2 ) and by the cylinder ( 3 ) (with its side walls). The maximum combustion chamber volume ( 11 ) spanned in this way is shown in dash-dot lines and in this case results approximately in a cuboid.

Fig. 4.1: "Quasi-Langhuber"; Design of an orbital lifting gear according to the invention, wherein the output takes place via an eccentric shaft ( 2 ), on the eccentric pin of which a one-piece connecting rod ( 10.3 ) is mounted, which with the piston ( 4 ) via the piston-side articulation points ( 4.5 ), ( 4.5 ') in the sections ( 7.1 ) and ( 7.2 ) of the rocker are non-positively connected with a conventional bolt connection.

It can be seen that the eccentric pin ( 2 ) moves in the area between the parts ( 7.1 ) and ( 7.2 ) of the rocker arm or the piston-side articulation points ( 4.5 ), ( 4.5 ′) and the pin edge areas thereby the plane of the axis of rotation ( 4.3 ) projected piston-side knuckle surface ( 4.4 ) (here: right-angled); Purpose: to reduce the weight or reduce the influence of mass on the output torque; also minimizes installation space; Another target measure in this regard: The center distance (= distance between the center of the eyes) of the connecting rod should be minimized. The relevant bearings on the housing side must not be less than a minimum radial distance from one another (bearing dimension, strength).

The piston has structurally localized relief contours for the same reasons.

The (favorable) effect of these design variables is reflected, inter alia, in the curve of FIG. 5.3.

Furthermore, here: the loading surface ( 4.2 ) essentially forms the extension of the imaginary chord of the loading surface ( 4.1 ), which is essentially concentric with the axis of rotation ( 4.3 ).

Otherwise: gas exchange, seal, flow conditions can be constructed similarly as was shown in Fig. 2ff.

Fig. 4.2: "Quasi-short stroke"; Variant of an orbital lifting gear according to the invention, the output taking place analogously to FIG. 4.1, drawn with the hinge point ( 7.5 ) near the piston in the upper extended position.

Compared to Fig. 4.1, there is the possibility of reducing the dead volume; this gives you more flexibility for the design of flow formers (troughs, nozzles,...). Compared to conventional two-stroke engines, a smaller stroke-bore ratio is permissible insofar as the course of the "useful stroke" (cf. Fig. 5.2) allows flushing, inlet and outlet slots to be advanced (or with regard to the crank angle) . Although the surface / volume ratio is less favorable ("Kurzhuber"), due to the special spatial and temporal sequence of the process cycles, higher heat losses are not necessarily to be expected.

The piston is lighter. The machine has a further reduced mass influence with respect to the torque line; The curve is similar to Fig. 5.3 or (still) cheaper.

Furthermore, here: the exposure area ( 4.2 ) falls in comparison. to Fig. 4.1 was much greater. In contrast, the applied area ( 4.1 ) i. trained to the latter much smaller.

Furthermore: The flow formers described above can preferably be located on or in the wall area of the cylinder ( 3 ) as well as in the equivalent piston area ( 4.2 ).

Otherwise: gas exchange, seal, flow conditions, for example similar to that in Fig. 2.2.

Fig. 5: The representations are based on realistic comparative calculations with the same stroke volume and constant angular velocity of the shaft ( 1 ), as well as a function of the crank angle (° KW) or the time (t) (exception: orbital lifting gear based on Fig. 1.3: here both the stroke volume and the angular velocity of the shaft are about half the size of the other machines).

Fig. 5.1: The following applies to the (swivel) piston machine according to the invention: the piston acceleration is analogous to the angular acceleration about the axis of rotation ( 4.3 ).

In the "special prior art", WO 88/08922 and EP 88/00367 (dashed curve course), the course of the acceleration curve with its additional turning points simultaneously signals load change points and associated unfavorable load peaks. It can be seen that here the piston just reaches its acceleration maximum with a comparatively large lever arm on the crank and, as a result, develops the (less favorable) curve shown in FIG. 5.3 as a result of the oscillatory inertia forces.

With the "special state of the art", the longer residence time of the piston in the TDC can be one Increasing the amount of heat dissipated via the combustion chamber wall per work cycle cause (in OT, the surface / volume ratio is known to have a maximum; consequently, the heat-dissipating area is the largest under this aspect).

In addition, due to the high pressure conditions prevailing in OT and the comparatively longer partial cycle duration increasing leakages on the piston rings result in an equally negative effect on efficiency.

In contrast, the measures taken according to the invention can be an ignition and Expansion cycle that largely avoids these disadvantages. The process of Mixture preparation, ignition and expansion is made possible by faster mixture distribution structured "quasi-dynamically" via the effective surfaces (instead of "quasi-static" process in OT in the "special prior art"), cf. Description.

Fig. 5.2: Pressure-time diagram or pressure-crank angle diagram. The additionally shown dash-dotted line for the (swivel) piston machine according to the invention corresponds to the piston travel (swivel angle). The maximum tangential force of the internal combustion engines according to Fig. 1.4 and Fig. 4.1 is 60 ° crank angle, at least as in the "special prior art" (the conventional reciprocating machine achieves its maximum value at approximately 75 ° crank angle), while the variant according to the invention Fig. 1.3 developed the maximum tangential force at a later time as "special prior art".

The dash-dotted line representing the piston travel (or swivel angle) of the piston machine according to the invention shows that at approximately 60 ° KW the piston of the machine according to FIG. 4.1 has covered approximately 1/3 of its stroke, ie essentially the "special state of the art Technology "corresponds, while the piston of the machine according to FIG. 1.3 and the conventional reciprocating piston machine each covered about 50% of their stroke.

The machine according to FIG. 1.4 develops its maximum torque even earlier, so that maximum gas power in the cylinder and maximum torque can be correlated to the crank.

Fig. 5.3: Mass tangential pressure diagrams at constant speed. Motors according to the invention have a softer curve with the smaller amplitudes. The proportion of unbalance generated by inertial forces is reduced according to the invention. Comparatively less additional effort is also required for balancing the machine. The course of the curve also indicates that the "special state of the art" makes it more difficult to master inertial forces, the higher the engine speed, the more it is mastered.

Fig. 5.4: Pressure-time diagram (pressure-crank angle diagram), similar to Fig. 5.2, but for a complete four-stroke cycle. The assessment criterion or the benchmark is the maximum tangential force achieved by the "special prior art", ie essentially after about 1/3 of the respective piston stroke. At about half the displacement and half the speed of the output shaft as in the comparison engines (general and special "state of the art"), the internal combustion engine according to the invention achieves a higher torque on the basis of FIG. 1.3 after about 1/3 of the piston stroke at the same time these machines. In addition, it can be seen that the internal combustion engine according to the invention has further potential for savings, in this connection especially with regard to the work required for the further cycles (curve shape with lower amplitudes, smaller pitch angles at the turning points).

Fig. 6: with the exception of Fig. 6.4, Fig. 6.9 and Fig. 6.0, which represent "quasi-long stroke" - only "quasi-short stroke" are shown. This representation is to be understood as a schematic overview. In addition to the different geometrical design, in particular the pistons, there are variations with regard to the sealing limit and the stroke space designed in this regard.

A conversion of Fig. 6.1 into a practically functioning machine is not sensible due to the concentrated power flow in the axis area ( 4.5 ).

Fig. 6.2: The concentrated flow of force is "equalized" by changing the swing arm geometry, angular swing arm (s) ( 7.9 ).

Fig. 6.3: Swing arm integrated in the piston, partial areas ( 7.1 , 7.2 ).

Fig. 6.4: driven via the eccentric shaft (2), with respect to FIG 6.3 modified sealing boundary in the neck area of the piston (4.1) and corresponding thereto hub chamber geometry..

Fig. 6.5: Variation of the sealing limits ( 4.6.2 and 4.6.3 ) and their use as charge exchange control units .

Fig. 6.6: similar to Figure 6.5, but with a reduced distance between the piston bottom (4.0) and Kolbenanlenkstelle (4.5)..

Fig. 6.7: like Fig. 6.5, structural variation, formation of a special trough geometry ( 4.1 ).

Fig. 6.8: piston with a comparatively simplem structure, reduced spectrum of uses.

Fig. 6.9: Piston geometry for a two-chamber system and "clean" pre-compression.

Fig. 6.0: like Fig. 6.9, but in the version as a combined "quasi short / long stroke".

FIG. 7 embodiment variants, in particular of gear elements, on the basis of the orbital lifting gear according to the invention.

Fig. 7.1: Between the piston crown ( 4.0 ) and piston skirt ( 4.9 ), a linearly displaceable, guiding element ( 6 ) is integrated with essentially one degree of freedom, for example roller or plain bearing or guide. Purpose: to avoid piston tipping.

Fig. 7.2 outlines the functioning of this element schematically by means of the mutually displaced triangles.

Fig. 7.3: Variant of the piston guide; Prevent the piston from tilting. Movement sequence for three successive time periods t1 to t3 (upper picture = OT, lower picture = UT, middle picture = intermediate position). The piston guidance within the cylinder (axis 3.1 ) is mediated by means of coordinated bodies, whereby an articulation point ( 7.5 ) that forms a unit that cannot be moved with the rocker arm has a trochoidal outer contour and the articulation point on the piston side ( 4.5 ) is determined by the corresponding profile. Rollers can also be installed in the guide (upper picture).

Fig. 7.4 outlines the invention with a Watt's straight guide. Two rockers ( 7 a, 7 b) are connected to each other via an element ( 6.2 ), on which the piston ( 4 ) is articulated at the point ( 4.5 ), and in this way convey the desired (straight) guidance .

Fig. 8: Constructional design of a special orbital lifting gear.

Fig. 8.1 and Fig. 8.2 differ only in the manner in which the guiding body ( 6 ) is implemented at the articulation point ( 7.5 ) of the angle rocker ( 7.9 ) near the piston.

Fig. Instructs 8.1 of respective point a rectilinear guide (6.3), similar to the loop of crank engines, while in

Fig. 8.2 essentially an eccentric element ( 6.4 ) is used.

Fig. 8.3 shows an extended variant of the piston engine, which is arranged on an axis and opposed cylinders in which the pistons move back and forth. The straight guidance provided is very advantageous at this point, since on the one hand this enables a "clean" pre-compression in a single cylinder by means of double-acting pistons (two-cycle) and on the other hand the relative speed of the guiding / sliding partner is significantly reduced compared to conventional crank loop motors.

Other reference symbols serve as orientation and otherwise correspond to those previous sections.

. FIG. 9: Figs. 9.1 to 9.4 show Fig pressure, volume Diagamme (p, VD) or temperature, entropy diagram (T, SD) for machine according to the invention on the basis of Fig 1.3.. Detailed explanations can be found in the description.

Fig. 10: shows a simplified diagram of the control system, in particular for machine according to the invention on the basis of Fig 1.3.. The dashed lines on the blocks "actuator / control element" and "ε _g <1, ε _g = 1, ε _g <1" indicate that a change that is independent of the control loop (possibly additional), for example "manual", change of the functional elements concerned can be made. The main task of the control loop is to bring the real, actual operating state of the machine as close as possible to the conditions required for an optimal cycle. More detailed explanations can be found in the description.

Claims (28)

1.Orbital lifting gear for internal combustion engines in particular, the stroke of which is designed with a crank ( 2 ) eccentrically rotating driver to the axis of a sun wheel ( 1 ) and for the cylinder ( 3 ) there is a regionally guided piston ( 4 ) which together with an articulated lever ( 5 ) on a guide mediating body ( 6 ) with the proviso that there is an effect on the path of movement, the piston being in drive connection with a vibrating gear element ( 7 ), and thereby also forming chambers of changing volume, while the piston is guided and the articulated lever passes through its articulated position, characterized in that
that the piston ( 4 ) is controlled during its movement in a path, the shape and dimension of which is predetermined by an articulation point of the coupling ( 6 ),
and that during the revolution the piston movement along the path is limited by the top dead center (OT) and the bottom dead center (UT), which is defined by the distance of the coupling articulation point in the crank dead center angles,
the gear element belonging to the kinematic chain and the coupling ( 6 ) indirectly or directly adjacent executing an oscillating movement relative to the crank ( 2 ). 2. Orbital lifting gear according to claim 1, characterized in that a hinge of the coupling ( 6 ) is designed to be pivotable on a track or capable of being driven. 3. Orbital lifting gear according to claim 2, characterized in that the coupling ( 6 ) is multi-articulated, in particular two or three articulated. 4. Orbital lifting gear according to the preamble of claim 1, characterized in that the vibrating gear element ( 7 ) is designed as an angle lever ( 7.9 ) with an articulated rod being arranged in a radial extension from the pivot point. 5. Orbital lifting gear according to at least one of claims 1 to 4, characterized in that the transmission in top dead center (TDC) has a gear element, the spatial position of which is characterized by its coupling points, in such a way that from them the essentially shortest possible distance between the piston articulation point ( 4.5 ) and the pivot axis ( 7.4 ). 6. Orbital lifting gear according to at least one of claims 1 to 5, characterized in that a triangular plate is formed on the coupling ( 6 ), at one corner of which a bolt-shaft which is firmly connected from the piston head and is connected there is articulated. 7. Orbital screw jack according to at least one of claims 1 to 6, characterized in that a device ( 91 ) exists which enables the gear element which is mounted in and / or lockable to be positioned in the housing region ( 7.4 ). 8. Orbital lifting gear according to claim 1 and at least one of the preceding claims, characterized in that the piston articulation point ( 4.5 ) and the pivot point ( 7.5 ) of the rocker are pivotable about a common axis. 9. Orbital lifting gear according to claim 8, characterized in that in the action surface of the piston ( 4 ) a trough ( 41 ) extending in the peripheral direction is introduced, such and to such an extent that the top dead center position (TDC) passes over it becomes. 10. Hoist gear according to claim 1 and at least one of the preceding claims, characterized in that the vibrating gear element consists of two substantially parallel parts ( 7.1 and 7.2 ) at a distance from one another, which via articulation points ( 7.5 ) or ( 7.5 ') and / or ( 7.6 ) or ( 7.6 '), between which a gear element, in particular with a connecting rod eye ( 10.1 ), is articulated. 11. Hoist gear according to at least one of claims 8 to 10, characterized in that
that the rocker ( 7 ) is an integral part of the piston ( 4 ) and has a articulation point which is connected to an upper connecting rod eye ( 10.1 ) in an articulated and non-positive manner,
whereas a lower connecting rod eye ( 10.2 ) acts on the crank ( 2 ) in such a way
that a limited angular movement along a curved path, in particular an arc, arises. 12. Lifting gear according to claim 11, characterized in that the pivoting piston ( 4 ) has an action surface ( 4.1 ) which is designed to be concentric with its axis of rotation ( 4.3 ) substantially. 13. Lifting gear according to claim 12, characterized in that the piston ( 4 ) has two loading surfaces ( 4.1 ) and ( 4.2 ) which are arranged at an angle to one another and essentially form the piston head. 14. A lifting gear according to claim 1 and at least one of the preceding claims, wherein the output takes place via an eccentric shaft, on the journal of which a connecting rod is mounted, which is in a non-positive connection with the piston, characterized in that
that the eccentric pin ( 2 ) moves in the area between the parts ( 7.1 and 7.2 ) of the oscillating gear element,
and thereby intersects a plane of the knuckle surface projected in the axial direction of the pivot axis ( 7.4 ) on the swing arm. 15. A lifting gear according to claim 1 and at least one of the preceding claims, wherein the output takes place via a shaft, the cheek and eccentric pin form an undivided unit and a connecting rod is mounted on the pin, which is non-positively connected to the piston, characterized that the connection between piston ( 4 ) and eccentric ( 2 ) is made by a one-piece connecting rod ( 10.3 ). 16. Lifting gear according to claim 1 and at least one of claims 2 to 5, characterized in that the extended connecting line of the shaft center ( 1 ) and rocker bearing ( 7.4 ) with the cylinder axis ( 3.1 ) intersect at an angle. 17. A lifting gear according to claim 1 and 2, characterized in that the piston articulation point ( 4.5 ) is attached to a substantially symmetrically designed four-bar gearbox with an approximate straight line. 18. Lifting gear according to claim 1 and 2, characterized in that the connection of the rocker ( 7 ) and piston ( 4 ) by means of an eccentrically acting part ( 6.4 ) which has a curved bearing surface which centers on an axis which is unchangeable with respect to the piston is, and has a further curved bearing surface, which is fixed and centered on the axis of the rocker pivot point ( 7.5 ), the axes may fall within the respective other area delimited by their bearings. 19. Lifting gear according to claim 1 and 2, characterized in that between the piston crown ( 4.0 ) and piston skirt ( 4.9 ) a guide-imparting, linearly displaceable element ( 6 ) is integrated with essentially one degree of freedom. 20. Lifting gear according to claim 1 and 2, characterized in that the rocker ( 7 ) has a pivot point ( 7.5 ) which has a substantially curved, in particular trochoidal outer contour and is in engagement with a corresponding inner contour of the piston-side pivot point ( 4.5 ) . 21. A lifting gear according to claim 1 and 2, characterized in that the piston shaft ( 4.9 ) via a crank-loop-like gear member, in particular link guide, with the swing-side articulation point ( 7.5 ) is in positive connection. 22. A lifting gear according to claim 7, characterized in that the functional mode of operation of the device ( 91 ) is integrated in a control circuit ( 80 ). 23, lifting gear according to claim 22, characterized in that the control circuit ( 80 ) belongs to a sensor system, which continuously detects the operating state of the motor (dis - /) and transfer it to an electronic system and the latter can have an (evaluation) logic which control over the operation of a control member ( 71 ), which is mounted in or on the supply tract of a cylinder. 24. lifting gear according to claim 1 and 2, characterized, that specially designed in or on the cylinder-side combustion chamber Contours are located during the compression and / or Expansion stroke the flow pattern of the cylinder filling significantly to shape. 25. Hoist gear according to claim 1 and 2, characterized in that the output ( 1 ) is made in several parts, the divided segments of which are connected to one another by a gear, in particular a synchronizing gear. 26. A method for changing the compression and / or expansion ratio of a piston machine with a lifting gear according to claim 1 and at least one of the preceding claims, with the step of providing a device for (on /) adjustment and / or change of the geometric compression and / or or expansion ratio, characterized by the following process steps:
Generating incremental changes in position of the gear member mounted in ( 7.4 ) on the basis of known and / or predetermined load status values of the orbital lifting gear. 27. The method according to claim 26, with the method steps of providing an actuator and a control member which is attached in or on the supply tract of a cylinder, characterized by the following method step:
Generation of incremental operating states of the control member by the actuator based on known and / or predetermined values in response to changes in location of the gear member mounted in ( 7.4 ). 28. The method according to claim 27, with the step of providing a control loop, in which at least one sensor and at least one actuator are integrated, which detect or change the operating state of the cylinder, characterized by the following steps:
Specification and / or determination of the current geometric compression and / or expansion ratio,
Detection of the operating state in the cylinder and / or on the housing by means of the sensors,
Further processing of the data signals obtained according to a logic within (evaluation) electronics; and
Manipulation of the control element by an actuator or actuator in response to a change in state with / without exceeding specified limit values.