Abstract:
In a method for reducing rotational non-uniformities of the crankshaft of a piston internal combustion engine, a movable balancing mass component is coupled to the crankshaft in such a way that the kinetic energy of the mass component increases in phases in which the angular speed of the crankshaft would increase if said crankshaft were not coupled to the balancing mass part, and decreases in phases in which the angular speed of the crankshaft would decrease if said crankshaft were not coupled to the balancing mass component.

Description:
CROSS-REFERENCE 
     This application is the U.S. national stage of International application no. PCT/EP2008/004205 filed May 27, 2008, which claims priority to German patent application no. 10 2007 025 549.9 filed May 31, 2007. 
     TECHNICAL FIELD 
     The invention relates to a method and device for reducing rotational non-uniformities of the crankshaft of an internal combustion piston engine. 
     RELATED ART 
     A characteristic of the rotation of the crankshaft of internal combustion piston engines, in particular reciprocating-piston internal combustion engines, is that this rotation, which is generated by the working cycles of the piston(s) connected to the crankshaft, is non-uniform. This non-uniformity increases when the number of cylinders is reduced, when rotational speeds are low, and at high loads. In practice, it is attempted to counteract these rotational non-uniformities by using large flywheel masses, which are connected, if necessary, with the crankshaft in a manner having a low torsion constant. A further attempt to counteract the rotational non-uniformities is to connect a cam disk with the crankshaft in a torque-proof manner, which cam disk is followed by a follower member that is moveable against a spring force. During phases in which the angular speed of the crankshaft is to be reduced, the follower member is moved towards the spring, so that potential energy is stored therein. This potential energy can be stored back in the rotation of the crankshaft during phases in which the angular speed of the crankshaft is to be increased. Despite all previously-existing measures, engines with few cylinders, e.g., four or fewer cylinders, exhibit comfort disadvantages as compared to high-cylinder engines, which disadvantages adversely affect their acceptability. In view of the increasing importance of low fuel consumption, a so-called “downsizing” of the engines is desired, which downsizing necessarily leads to a reduced number of cylinders due to the minimum volume of the individual cylinders, which minimum volume is dictated by thermodynamic reasons. 
     A drivable shaft having a moment of inertia that is variable in an acceleration-dependent manner is known from DE 196 49 712 C2. An inertial mass, which moves counter to the rotational direction of the shaft from an idle position relative to the shaft to a displaced position when the rotation of the shaft is accelerated, moves flywheel mass elements via transmission means during its displacement movement towards a reduction of the radial distance from the rotational axis; the flywheel mass elements are connected to the shaft in a torque-proof manner and are movably guided approximately radially relative to the rotational axis. The inertial mass is formed as an inertial ring extending concentrically around the shaft, which inertial ring is connected with each flywheel mass element via the transmission means. If, starting at a state in which the flywheel mass elements are spaced a large distance from the shaft, the rotational speed of the shaft increases, the inertial ring lags behind the shaft, so that the distance between the flywheel mass elements and the shaft decreases. Consequently, the moment of inertia of the assembly decreases when the shaft accelerates, which boosts the acceleration of the shaft. When the rotational speed of the shaft decreases, the moment of inertia of the assembly increases, whereby the reduction of the rotational speed is boosted. All in all, this assembly achieves an improved response behavior of the internal combustion engine by reducing the effective moment of inertia. 
     SUMMARY 
     In one aspect of the present teachings, the rotational non-uniformities of the crankshaft can be considerably reduced, so that internal combustion piston engines having a low number of cylinders, e.g., four or fewer cylinders, exhibit a running comfort that is acceptable to demanding customers. 
     In another aspect of the present teachings, methods and devices capable of reducing rotational non-uniformities or torsional fluctuations of the crankshaft of an internal combustion piston engine are disclosed. For example, a movable compensating mass element may be coupled with a crankshaft such that the kinetic energy of the compensating mass element increases during phases in which the angular speed of the crankshaft would increase if it were not coupled with the compensating mass element, and decreases during phases in which the angular speed of the crankshaft would decrease if it were not coupled with the compensating mass element. 
     In another aspect of the present teachings, the compensating mass element is rotatable about an axis and is coupled with the crankshaft such that the angular speed of the compensating mass element increases relative to the angular speed of the crankshaft during phases in which the angular speed of the crankshaft would increase if it were not coupled with the compensating mass element. In addition, the angular speed of the compensating mass element preferably decreases relative to the angular speed of the crankshaft during phases in which the angular speed of the crankshaft would decrease if it were not coupled with the compensating mass element. 
     In another aspect of the present teachings, the amount of the increase and decrease of the angular speed of the compensating mass element is preferably changeable relative to the angular speed of the crankshaft. 
     In another aspect of the present teachings, a guide coupler is preferably rotatably supported relative to the crankshaft in a manner such that the rotational axis of the guide coupler is movable or displaceable relative to the rotational axis of the crankshaft. The guide coupler may be coupled, via at least one connecting member, with a driver extension that is connected to the crankshaft in a torque-proof manner, i.e. such that the driver extension and the crankshaft rotate together in a fixed relationship. The guide coupler is also preferably connected with the compensating mass element via at least one additional connecting member. The axes of all hinge connections between the connecting members, the guide coupler and the driver extension are preferably parallel to the rotational axis of the crankshaft. In a further optional embodiment, an imaginary line connecting the rotational axis of the guide coupler and the rotational axis of the crankshaft, as viewed along the rotational axes, is approximately aligned or substantially coincides with the line of movement of a piston connected to the crankshaft of the internal combustion piston engine. 
     In another aspect of the present teachings, energy and/or power may be drawn off from the crankshaft and stored in the compensating mass element during phases in which the crankshaft supplies a high torque. During phases in which the crankshaft supplies little torque or must be driven itself, because the piston(s) connected to the crankshaft is(are) performing compression work or discharge work, the energy stored in the compensating mass element is returned to the crankshaft. By directly storing kinetic energy in the compensating mass element, no complex energy conversions are necessary and such an embodiment can operate with high efficiency. 
     The invention will be explained below, exemplarily and in further detail, with the assistance of schematic figures. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIGS. 1 and 2  depict end views of an embodiment of the present teachings in different operating states, 
         FIGS. 3   a ) to  3   c ) depict representative torque curves of a 4-cylinder, four-stroke, spark-ignition engine operated according to one aspect of the present teachings, 
         FIG. 4  depicts a table that shows the main orders of the alternating torques that are preferably compensated for different types of engines, 
         FIG. 5  depicts a schematic end view of an embodiment modified with regard to  FIG. 1 , 
         FIG. 6  depicts a schematic view of a piston-cylinder unit, 
         FIG. 7  depicts a schematic end view of a modified embodiment of  FIGS. 1 and 2 , 
         FIG. 8  depicts a schematic diagram of an embodiment of an adjusting device and 
         FIG. 9  depicts a sketch for explaining a mass compensator. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     According to  FIG. 1 , which shows an end view of an inventive device, a crankshaft  10  of an internal combustion piston engine having a known structure comprises a radially-protruding driver extension  12  that is connected in a torque-proof manner with, or is formed integrally with, the crankshaft  10 . A compensating mass element  14  is supported on the crankshaft  10  adjacent to the driver extension  12 , which compensating mass element is advantageously balanced with regard to the rotational axis A of the crankshaft  10 . 
     Axially adjacent to the compensating mass element  14 , a bearing shield  16  (shaded; the axial arrangement is not visible in  FIG. 1 ) is supported on a not-illustrated engine housing so as to be pivotable about an axis B that is spaced from the rotational axis A of the crankshaft. The bearing shield  16  has a through-hole  20 , through which the crankshaft  10  extends, so that the above-described component assembly can be disposed at a terminal end of the crankshaft or in a central region of the crankshaft. The pivot axis B is parallel to the rotational axis A of the crankshaft. A guide coupler  22  with a radially protruding arm  24  is rotatably supported on the bearing shield  16  that is pivotable relative to the crankshaft. 
     The radial arm  24  is connected to the driver extension  12  via a connecting member  26  and is connected via a further connecting member  28  with the compensating mass element  14  in a hinge  29 . Elements, such as driver extension  12  and connecting members  26  and  28 , that connect the crankshaft  10  to the compensating mass element  14  may be referred to generally as a “coupling device.” Advantageously, the connecting members  26  and  28  are coaxially hinged to the arm  24  of the guide coupler  22 . The connecting member  28 , for example, projects into a radial slot (not illustrated) of the compensating mass element  14  and is supported therein by a pin. The pivot axes of the hinges, about which the connecting members are pivotable relative to each other and are pivotable relative to the driver extension  12 , the guide coupler  22  and the compensating mass element  14 , are parallel to each other and are parallel to the axes A and B. 
     By pivoting the bearing shield  16  about the axis B of the bearing mounted on the engine housing, the spacing between the rotational axis C, about which the guide coupler  22  rotates, and the rotational axis A of the crankshaft can be changed. An adjusting device, identified as a whole by  40 , is provided for pivoting the bearing shield  16 , which adjusting device  40  includes an adjusting member  42  connected to an arm  44  in a hinged manner, which arm  44  is rigidly connected to the bearing shield  16 . Approximately in the direction of extension of the arm  44 , the bearing shield  16  has another arm  46  on its opposite side, which is rigidly connected therewith and is pivotable about the pivot axis B. The adjusting member  42  can be moved in the vertical direction according to  FIG. 1  in a known manner using a hydraulic cylinder or an electric motor, wherein the moveability is advantageously limited by stops  48  and  50 . 
     The above-described device can be constructed in a highly compact manner with an interleaved construction. The compensating mass element  14  is rotatably supported on the crankshaft between the guide coupler  22  and the driver extension  12  and coaxially with the rotational axis A of the crankshaft. The guide coupler  22  is supported so as to be rotatable about the rotational axis C that is defined by the bearing shield  16 . Axially adjacent to the guide coupler  22 , the arms  44  and  46  of the bearing shield  16  extend. It is understood that the eccentric pivotability of the bearing shield  16  relative to the crankshaft and its adjustment can also be provided by other constructions, e.g., by supporting the bearing shield directly with a pin in the engine housing, which pin is disposed eccentric to the rotational axis of the crankshaft. 
       FIG. 1  illustrates the state of the assembly wherein the rotational axis C of the guide coupler  22  is maximally spaced from the rotational axis A of the crankshaft  10 . Advantageously, the connecting line between the rotational axes A and C is parallel to the moving direction of the not-illustrated pistons of the internal combustion engine, which are connected to the crankshaft  10 , and is aligned in the illustrated view with the line of movement of the pistons. Minor directional changes of the connecting line during pivoting of the bearing shield  16  about the pivot axis B need not be considered. The arrangement of the driver extension  12 , the radial arm  24  and the connecting members  26  and  28  is such that the connecting members  26  and  28  pass through their extended position twice during one revolution of the crankshaft  10  or during the revolution of the guide coupler  22  that is eccentric therewith, so that the compensating mass  14  of the crankshaft  10  advances and lags, respectively, twice during one revolution, i.e. rotates at a higher and a lower speed, respectively, than the crankshaft. The advancement and lag, respectively, can be adapted to the particular requirements by the geometrical arrangement of the connecting members. 
       FIG. 2  shows the state of the assembly of  FIG. 1  wherein the rotational axes A and C coincide; that is, the state of maximal eccentricity according to  FIG. 1  has been shifted to the state of minimal eccentricity in which the guide coupler  22  rotates coaxially with the crankshaft  10 . In this state of minimal eccentricity, the positioning of the driver extension and the connecting members  26  and  28  and the radial arm  24  relative to each other remains constant, so that the compensating mass element  14  always rotates at the same angular speed as the crankshaft  10 . 
     In contrast thereto, in the position according to  FIG. 1 , the compensating mass element  14  must be cyclically accelerated and decelerated, so that the crankshaft must apply additional torque to the compensating mass element during phases of rotation in which the compensating mass element is accelerated, and the crankshaft receives an additional torque from the compensating mass element during other phases of rotation, in which the angular speed of the compensating mass element is reduced relative to the crankshaft. 
       FIG. 3   a ) shows an exemplary torque curve of a charged, 4-cylinder, spark-ignition engine that is operated in the four stroke process and rotates at a speed of, for example, 1000 rpm at full load. The abscissa indicates the rotational position of the crankshaft in degrees, and the ordinate indicates the torque from the crankshaft, which torque is acting in the drive train of the vehicle. The torque fluctuations are larger at increasing loads and become increasingly noticeable in an unpleasant manner in the drive train when the rotational speed decreases, whether it is due to acceleration fluctuations of the entire vehicle or due to noises or vibrations, all of which are perceptible by the vehicle occupants in a comfort-reducing manner. 
     Part b) of  FIG. 3  shows the torque that is producible by the above-described first embodiment and that can be applied by the compensating mass  14  to the crankshaft  10 . As can be seen, a negative torque is necessary for accelerating the compensating mass element during phases in which the crankshaft generates a high positive torque and, conversely, the compensating mass element supplies a positive torque to the crankshaft in the phases during crankshaft deceleration in which the crankshaft experiences a negative torque due to the work being performed by the pistons. In summary, by equipping a known internal combustion engine with a device according to the first embodiment, an overall torque, which is illustrated in  FIG. 3   c ), results at the crankshaft, wherein the torque fluctuations of the crankshaft are considerably reduced as compared to the torque fluctuations of  FIG. 3   a ). Hence, a very comfortable, even speed of the internal combustion piston engine is achieved, even when it has a relatively low number of cylinders. It is pointed out that  FIG. 3  shows the torque fluctuations. In particular in the illustrations of  FIGS. 3   a ) and  3   c ), the curves must be shifted upwards by the average torque that is supplied by the crankshaft to the vehicle. By pivoting the bearing shield  16 , the device can be adjusted to the respective compensation requirements. At high fluctuations, it is therefore operated with high eccentricity and/or with a large distance between the rotational axes A and C. At low fluctuations, the eccentricity is decreased. 
       FIG. 4  shows in tabular form the main orders of the alternating torques shown in  FIG. 3   a ), which must be compensated, for internal combustion engines having different numbers of cylinders and different operating principles (two-stroke process, four-stroke process). 
     The arrangement shown in  FIGS. 1 and 2  is suitable for reducing torque fluctuations at the crankshaft for the second order of a 2-cylinder two-stroke engine or a 4-cylinder four-stroke engine. 
     In the second embodiment according to  FIG. 5 , the eccentric movement and/or the stroke movement of the guide coupler  22  relative to the crankshaft  10  is converted into a relative rotation between the compensating mass element  14  and the crankshaft  10  by pivoting the connecting member  52 . This pivoting takes place once during each revolution of the crankshaft, so that the embodiment according to  FIG. 5  is suitable for reducing torque variations at the crankshaft for the first order in a 1-cylinder two-stroke engine or a 2-cylinder four-stroke engine. 
     Devices according to the present teachings can also be used for compensating torque fluctuations and/or rotational non-uniformities at the crankshaft of other engines, e.g., 3-cylinder engines, in which, for example, the driver extension  12  is not rigidly connected with the crankshaft, but rather is connected with the crankshaft via a speed-increase gear unit or speed-decrease gear unit, as a result of which the driver extension rotates at a speed suitable for compensating torque fluctuations, e.g., in a 3-cylinder engine having a rotational speed increased by the factor 1.5 in comparison to the embodiment of the respective device for a 2-cylinder engine. 
       FIG. 6  schematically shows a piston  60  of a cylinder of a reciprocating-piston internal combustion engine, which is connected with the crankshaft  10  via a piston rod  62  in a known manner. During four-stroke operation, only one working stroke takes place in a known manner during two revolutions of the crankshaft; a driving torque MK is applied to the crankshaft  10  by the piston  60  in the working stroke. The inclination of the piston rod  62  relative to the vertical line results in that the piston  60  is braced on the cylinder wall by lateral forces. These lateral forces can be compensated by applying a counter moment FN×s(t) to the engine block. 
       FIG. 7  shows an example of how such a counter moment is applied to the engine block. For this purpose, the assembly according to  FIG. 1  is augmented in the illustrated example such that the compensating mass element  14  is provided with a circumferential toothing  64 , and a further compensating mass element  66  is supported on the engine housing so as to be rotatable about an axis D, which is spaced from the rotational axis A of the crankshaft; the compensating mass element  66  is formed with a circumferential toothing  68  whose diameter is half the diameter of the circumferential toothing  64 . Thus, the other compensating mass element  66  rotates at a speed that is twice the speed of the compensating mass element  14 , wherein it fully experiences the rotational non-uniformities of the compensating mass element  14 . In this way, the additional compensating mass element  66  can be used for generating a counter moment that compensates the lateral piston forces, according to  FIG. 6 , in the illustrated example for a 4-cylinder four-stroke engine. 
     At this point, it is again pointed out that the compensating mass element  14  and the additional compensating mass element  66  are advantageously balanced with regard to their respective rotational axes. They can be formed as disk-shaped or in other ways. The additional compensating mass element  66  may be used for a known two-shaft compensation of the second order for the free inertial forces. In contrast to the conventional dampening of rotational non-uniformities using a flywheel, according to one aspect of the present teachings the angular speed of the compensating mass element is systematically changed substantially in phase with the angular speed of the crankshaft. Consequently, the moment of inertia of the compensating mass element can be smaller than the moment of inertia of conventional flywheels, so that the engine responds well to position changes of the accelerator pedal. 
       FIG. 8  schematically shows a particularly simple embodiment of an adjusting device  40 . The arm  44  of the bearing shield  16  ( FIG. 1 ) is connected in a hinged manner to a shaft of a piston  72  of a double-acting piston cylinder unit, which shaft serves as an adjusting member  42 ; the pressure chambers of said piston cylinder unit are connected with each other via two conduits  76  and  78 . A one-way valve  80  and  82  is arranged in each of the conduits  76  and  78 , wherein the one-way valves  80  and  82  act oppositely to each other. Further, a shutoff valve  84  and  86  is disposed in each respective conduit. 
     Due to the eccentric bearing of the bearing shield  16 , alternating forces act on the arm  44  during each revolution of the compensating mass element  14 , which forces are directed upwards or downwards according to the figures. Depending on the opening of one of the shutoff valves  84  or  86 , a hydraulic fluid located in the piston cylinder unit  74  can only flow from one pressure chamber into the other one, so that a displacement in the one or the other direction takes place by correspondingly controlling the shutoff valves  84  and  86 , without the need for an external source of pressurizing medium. It is understood that a suitable re-filling mechanism for hydraulic fluid ensures that the conduits and the piston cylinder unit are constantly filled with hydraulic fluid and are free from air. 
     With the help of  FIG. 9 , a fourth embodiment is explained that shows how, according to another aspect of the present teachings, not only rotational non-uniformities of the crankshaft can be compensated, but also how a mass compensation can be carried out in a 1-cylinder two-stroke engine or in a 2-cylinder four-stroke engine. The compensating mass element of  FIG. 9 , which is denoted with 14, corresponds to the compensating mass element  14  of the second embodiment according to  FIG. 5 , whose remaining components are not illustrated in  FIG. 9  for the sake of simplicity. The compensating mass element  14  has a circumferential toothing  64  that meshes with a circumferential toothing  68  of another compensating mass element  88 . Reference numerals  90  and  92  denote unbalanced masses of the compensating mass elements; that is, in contrast to the above-described embodiment, the compensating mass element  14  is not balanced relative to its rotational axis, but rather has a specific unbalanced mass  90 . The unbalanced masses  90  and  92  are preferably equal and arranged such that they are each located simultaneously in the vertical position according to  FIG. 5 , i.e. they simultaneously pass the top dead center and the bottom dead center, and are disposed in opposite phase in their horizontal positions. As was already explained with the help of  FIG. 5 , this arrangement is also preferably used for compensating rotational non-uniformities of the crankshaft of a 1-cylinder two-stroke engine or 2-cylinder four-stroke engine. When the unbalanced masses are arranged relative to the piston of a 1-cylinder two-stroke engine or the pistons of a 2-cylinder four-stroke engine such that they are located in the bottom dead center when the piston(s) is/are located in the top dead center, a mass compensation can be achieved for said engines with the arrangement according to  FIG. 5 . 
     The present teachings, which have been described above in an exemplarily manner, can be modified in various ways. In the embodiments according to  FIGS. 1 ,  2 ,  5  and  7 , the adjusting device  40  may, for example, comprise a hydraulic cylinder whose hydraulic fluid is pressurized fuel, e.g., diesel fuel diverted from a common rail system, so that no separate hydraulic pressure source is required. The hydraulic cylinder is controlled in a known manner using valves that are actuated by an electronic control device, dependent on the torque fluctuation compensation required at the respective load point of the internal combustion engine. The described pivoting of the bearing shield  16  can be replaced by a linear movability of the bearing shield in a coulisse that is affixed to the engine housing. The present teachings are suitable for use in reciprocating-piston internal combustion engines as well as in rotary-piston internal combustion engines. 
     Instead of, or in addition to, the adjustability of the bearing shield  16 , the compensating mass element  14  can be formed such that its moment of inertia  14  is changeable, e.g., by providing the compensating mass element with two mass bodies whose spacing is changeable. By changing the moment of inertia, the device can be adapted to the respective to-be-compensated rotational non-uniformities. 
     A device having the compensating mass element and the coupling with the crankshaft may be disposed variously, for example, at an end of the crankshaft or—in engines with plural cylinders—between the pistons. Several devices according to the present teachings can be provided on one crankshaft. 
     REFERENCE NUMBER LIST 
     
         
         
           
               10  crankshaft 
               12  driver extension 
               14  compensating mass element 
               16  bearing shield 
               20  through-hole 
               22  guide coupler 
               24  radial arm 
               26  connecting member 
               28  connecting member 
               29  hinge 
               40  adjusting device 
               42  actuator 
               44  arm 
               46  arm 
               48  stop 
               50  stop 
               52  connecting member 
               54  arm 
               56  arm 
               58  connecting member 
               60  piston 
               62  piston rod 
               64  circumferential toothing 
               66  compensating mass element 
               68  circumferential toothing 
               72  piston 
               74  piston-cylinder unit 
               74  conduit 
               76  conduit 
               80  one-way valve 
               82  one-way valve 
               84  shutoff valve 
               86  shutoff valve 
               88  compensating mass element 
               90  unbalanced mass 
               92  unbalanced mass