Patent Publication Number: US-2009220353-A1

Title: Axial Piston Compressor

Description:
The present invention relates to an axial piston compressor, especially a compressor for motor vehicle air-conditioning systems, in accordance with the preamble of claim  1 . 
     In the field of compressor drive mechanisms, a trend is beginning to emerge that, in the case of compressors having variable piston stroke, increasing use is being made of tilt plates in the form of a tilt ring, that is to say ring-shaped tilt plates, with a tilt-providing articulation necessary for tilting of the plate being substantially integrated into the ring-shaped tilt plate. For example, there is known, from EP 0 964 997 B1, a compressor in which the stroke movement of the pistons is accomplished by means of engagement—in an engagement chamber—of a ring plate oriented on a slant to the machine shaft. The engagement chamber is provided adjacent to the enclosed hollow space of the piston. For sliding engagement that is substantially free from play in any slanting position of the tilt plate or tilt ring there are provided on both sides, between it and the spherically curved inner wall of the engagement chamber, spherical segments, so-called sliding blocks, so that the tilt ring slides between them as it revolves. 
     The drive is transmitted from the drive shaft to the tilt ring by means of a pin for conjoint movement which is attached to the drive shaft and the spherical head of which engages in a radial bore in the tilt ring, the position of the head of the member for conjoint movement being selected so that its centre-point coincides with that of the spherical segments. In addition, that centre-point is located on a circular line which connects the geometric axes of the seven pistons with one another and, moreover, on a circular line which connects the centre-points of the spherical articulation members of the pistons. By that means, the upper dead-centre position of the pistons is determined and a minimum clearance volume is ensured. At this point it should be noted that the dead centre is constant for all deflection angles, or tilt angles, of the tilt ring and accordingly the afore-mentioned minimum clearance volume is ensured. The head shape of the free end of the member for conjoint movement makes it possible for the inclination of the tilt plate to change by means of the fact that the head of the member for conjoint movement forms a bearing body for a tilting movement of the tilt plate which changes the stroke distance of the pistons. 
     A further precondition for tilting of the tilt plate is the displaceability of its mounting axis in the direction of the drive shaft. For the purpose, the mounting axis is formed by two mounting pins mounted on the same axis on each side of a sliding sleeve, which mounting pins are additionally mounted in radial bores in the tilt plate. For the purpose, the sliding sleeve preferably has mounting sleeves on each side, which span the annular space between the sliding sleeve and the tilt plate in the manner of spokes. 
     The limitation on the displaceability of the sliding sleeve and, as a result, the maximum tilted position of the tilt plate results from the pin for conjoint movement, by virtue of the fact that the latter passes through an elongate hole provided in the sliding sleeve so that the sliding sleeve meets end stops at the ends of the elongate hole. The force for the change in the angle of the tilt plate and, therefore, for regulation of the compressor results from the sum of the pressures acting against one another in each case on each side of the pistons, so that this force is dependent on the pressure in the drive mechanism chamber. In accordance with the prior art, the pressure PC in the drive mechanism chamber can be regulated between a high pressure and a low pressure (inlet pressure) and consequently affects the balance of forces at the tilt plate, which influences the inclination of the latter. The position of the sliding sleeve can moreover be influenced by springs which, in various variants, are likewise included in the prior art. 
     Furthermore, the position of the sliding sleeve, which position governs the delivery output, is also determined by the forces of inertia acting on the tilt plate; the position of the tilt plate, that is to say its angle of tilt or slant, changes with increasing speed of rotation. In the case of modern compressors, the trend is towards using tilt plates having moments of inertia such that they bring about a reduction in the stroke distance of the pistons and therefore a reduction in delivery output when the speed of rotation increases. In this context it is desirable for the angle of tilt to become smaller when the speed of rotation of the compressor increases. 
     However, what is problematic in the arrangement explained hereinbefore is the high Hertzian stress in the region of the head of the member for conjoint movement (point contact/line contact) and the tilt plate (system: sphere/cylinder) and the take-up of the (axial) reaction forces due to the gas force on the pistons and the forces due to the torque to be transferred to the tilt plate. 
     A compressor similar to the compressor known from EP 0 964 997 B1 is known from JP 2003-269330 M, although in that compressor a total of two members for conjoint movement are used. 
     It is important to the kinematics according to the two mentioned publications, that is to say to the kinematics in the case of the subject-matter of EP 0 964 997 B1 and JP 2003-269330 M, that the head of the member for conjoint movement centrally coincides with the centre-point of the sliding blocks of the pistons and that the position of the centre-point of the head of the member for conjoint movement is at the same time approximately tangential to the reference circle of the central axes of the pistons. 
     Added to the afore-mentioned disadvantageous characteristics is the fact that the subject-matter of EP 0 964 997 B1 and of JP 2003-269330 M has a very complicated structural arrangement, which results in a high number of parts and therefore cost, and in addition the mounting by means of two members for conjoint movement is over-determined and therefore susceptible to wear, and the strength of the components, especially due to the fact that a hole is introduced into the shaft, has to be regarded as rather low. 
     A further compressor is known from DE 101 52 097 A1, differing considerably from the subject-matter of the publications discussed hereinbefore. In the case of the subject-matter according to DE 101 52 097 A1, the member for conjoint movement, in particular the spherical head of the member for conjoint movement, is replaced by a hinge pin or spindle. This is, however, integrated into the tilt plate from the outside and fastened using a cup-shaped disc for conjoint movement which is a component of the drive shaft assembly. The subject-matter of DE 101 52 097 A1 also has a complicated structural arrangement; in addition it has to be borne in mind that a large imbalance can come about, depending on the angle of tilt. This encourages wear on the compressor and as a result reduces its service life. 
     A further compressor is known from FR 278 21 26 A1, which has a member for conjoint movement extending out from the drive shaft radially and engaging in the tilt plate. In similar manner to the solution according to DE 101 52 097 A1, the tilt plate in this arrangement is also fixed to the member for conjoint movement in radial extension. In this there also lies a central difference from the subject-matter of EP 0 964 997 B1 and JP 2003-269330 AA. Whereas in the latter cases the mounting point of the head of the member for conjoint movement in the tilt plate undergoes relative movement in the guideway (bore) in the tilt plate because the tilt plate performs the rotary movement in an articulation lying on the shaft axis, the rotary movement in the case of the arrangements according to FR 278 21 26 A1 and DE 101 52 097 A1 is accomplished in the lateral articulation of the tilt plate. 
     In the unpublished Patent Application DE 102 00 404 1645 belonging to the present Applicant, there is proposed a member for conjoint movement (consisting of a supporting element which is in contact with the tilt plate, and of a force transmission element which is in operative engagement with the drive shaft and the tilt plate) which member for conjoint movement is displaceably mounted in the drive shaft. As a result, the transmission of force between the head of the member for conjoint movement and the tilt plate can be accomplished optimally (force transmission as a result of area-wise contact). However, the displacement of the member for conjoint movement in the drive shaft can be problematic because high forces have to be taken up there owing to the bending moment and the parts therefore have to be of very rigid construction. This rigid construction causes the compressor to have an increased mass. 
     From DE 103 154 77 A1 there is known a compressor of the tilt plate/member for conjoint movement constructional type wherein the member for conjoint movement does not transfer any torque. This feature in addition also applies to preferred arrangements of DE 102 00 404 1645. The conjoint movement function is restricted to providing support for the piston forces acting axially on the tilt plate, the torque being delivered by further elements independent of the member for conjoint movement. As a result, the forces acting on the member for conjoint movement are lower because, as already mentioned, no torque is transferred. The advantage of this approach lies in the fact that the forces or surface contact pressure due to the forces applied (because of the fact that these forces are relatively low) do not cause any excessive deformation at and in the member for conjoint movement, as a result of which the member for conjoint movement can be of correspondingly lightweight construction and tilting of the tilt plate can be accomplished in a relatively hysteresis-free manner or with relatively little hysteresis. However, a disadvantageous effect can be that the spherical head of the member for conjoint movement is located in a relatively large recess in the tilt plate. As a result, the Hertzian stress can or must be described by a plane/sphere geometric pairing, which is relatively disadvantageous because it causes a high degree of Hertzian stress. 
     From the likewise unpublished DE 10 2005 004 840 belonging to the present Applicant, there is known a compressor which provides an improvement in respect of the problem of surface contact pressure. The subject-matter of DE 10 2005 004 840 includes a support element in engagement with a tilt ring, with line contact arising between the support element and the tilt ring. Compared to the previously described prior art, this constitutes an improvement in respect of the Hertzian stress. A likewise advantageous effect is that, in the case of the subject-matter of DE 10 2005 004 840, a drive moment and a torsional moment are decoupled from gas force support. However, a relatively large recess is necessary in the tilt plate in order to ensure thereby a sufficient length of line contact and to achieve correspondingly low surface contact pressure. The large recess in the tilt plate could, because of the gas forces to be transmitted, result in deformation of the tilt ring and therefore in wear. Furthermore, the down-regulating behaviour of the tilt plate (which is dependent on the moment M sw  due to the particular relevant moment of deviation of the rotating masses and on the interplay between that moment and the moments of the masses moved in translation M K,ges ) and also the imbalance thereof are disadvantageously affected by a large recess. In the case of the subject-matter of DE 10 2005 004 840 the mass of the gas force support means does not affect the moment of deviation. 
     Finally, from the likewise unpublished Patent Application DE 10 2005 018 110 23 belonging to the present Applicant, there is known a compressor wherein the force transmission element is in rotatable and/or radially displaceable articulated connection with the supporting element. As a result there is provided a compressor whose supporting element can take up forces over a large area (which in turn corresponds to low Hertzian stress), whilst an imbalance of the tilt plate due to the mounting and tilting thereof and of further parts associated with the mass-related properties of the tilt plate is low over the entire tilt angle range and the entire speed of rotation range. Although the problem of surface contact pressures is solved in the subject-matter of DE 10 2005 018 110 23 in acceptable manner, the configuration proposed in said Application does, however, require a relatively large size of construction. Determination of the requisite diameters or cross-sectional profile of the force transmission element (for example, by means of a strength calculation) and determination of the possible deformation (because of the possibility of jamming within the guideway) lead to the result that a reduction in the size of construction of a compressor of such a kind would be desirable. The fact that the force transmission element is integrated into the bore in the supporting element and also, for that purpose, a residual wall thickness has to be provided which ensures adequate strength also results in the supporting element having a substantial diameter. A similar problem occurs if the supporting element is integrated into the tilt plate or tilt ring terminally. As a result, this kind of arrangement governs the height of the tilt plate to a not inconsiderable degree. To illustrate the point better, an example will be given as follows: an estimate of the requisite strengths or the dimensions required therefor results in a force transmission element diameter of 8 mm, a supporting element diameter of 14 mm (maximum residual wall thickness 3 mm) and a tilt plate height of 18 mm (maximum residual wall thickness 2 mm). Because the height of the tilt plate substantially influences the bridging region of the piston, which connects the bottom sliding block accommodating means to the top sliding block accommodating means, a more compact mode of construction is desirable. The larger the bridging region, the larger the ensuing bending moments become. Although appropriate dimensioning does prevent this, such a compressor then has a large housing diameter. 
     Starting from the prior art explained hereinbefore, the problem of the present invention is accordingly to provide a compressor wherein the mounting between the force transmission element and supporting element is so arranged that on the one hand a low degree of deformation in the region of the mounting is ensured whilst at the same time the tilt plate has a height which is as small as possible whilst ensuring consistently low Hertzian stress in the course of force transmission. 
     The problem is solved by a compressor having the features according to patent claim  1 , with preferred developments and embodiments being described in the subordinate claims. 
     A fundamental point of the invention accordingly is that the supporting element is in articulated connection with the force transmission element so as to be displaceable in a radial direction and/or perpendicular thereto, especially in a direction perpendicular to the drive shaft axis. This means, in other words, that the supporting element can be displaceable along an axis or, especially, in a plane. As a result, a smaller force transmission element than in the prior art can be used, because the requisite strength and rigidity is provided by a wide mounting. As a result, the desired low degree of deformation can also be ensured in a relatively small constructional form. Furthermore, the desired low degree of Hertzian stress is also ensured in the course of force transmission. 
     In a preferred embodiment, the supporting element is in the shape of a cylindrical pin, as a result of which optimal Hertzian stress is ensured at the same time as a constructionally simple structure. The supporting element can have a groove with which the force transmission element is in operative engagement. Alternatively the supporting element can also have a pocket-shaped recess, “pocket-shaped” denoting under the terms of the present invention an especially rectangular but also round or elliptical recess in the supporting element. In that case, optionally, at least that end region of the force transmission element which faces the supporting element is formed in the shape of a flat steel part, that is to say having an approximately rectangular peripheral contour. As a result, secure engagement between the supporting element and the force transmission element is ensured at the same time as a low degree of constructional complexity. 
     The force transmission element can furthermore be connected to the drive shaft in a manner which does not allow relative rotation, which ensures that a compressor according to the invention has a simple structure. Depending on the desired or required degrees of freedom, the force transmission element can also of course be rotatably mounted on or in the drive shaft. Optionally, the force transmission element can also be part of the drive shaft or integrated into the latter. In a preferred embodiment, the force transmission element is formed in one piece with the drive shaft. This ensures a functional construction at the same time as a low degree of constructional complexity. 
     Preferably, the supporting element and the force transmission element serve substantially only for providing the pistons with axial support or, that is, for gas force support, whereas an arrangement independent thereof, especially an articulated connection, between the drive shaft and the tilt plate serves substantially only for torque transfer. This ensures decoupling of the drive torque and gas force support. 
     In the case of a compressor according to the invention, preference is given to the tilt plate being pivotally mounted on a sliding sleeve mounted so as to be axially displaceable along the drive shaft, the tilt plate being connected by way of drive pins to the sliding sleeve and/or to the drive shaft. This ensures simple implementation of the decoupling of drive torque and gas force support. The drive pins can be introduced into the sliding sleeve or the tilt plate with a press fit or secured therein or thereon by means of axial securing elements or circlips. The drive pins can project into a recess, which can especially be in the form of a groove in the drive shaft. Optionally, a connecting element, especially in the form of a feather key, is arranged between the drive shaft and the sliding sleeve, which connecting element allows transfer of forces and moments in a radial direction and which is mounted in axially displaceable manner on the drive shaft. That end of the force transmission element which is remote from the supporting element can also project through the drive shaft and into a longitudinal slot in the sliding sleeve in such a way that drive torque is transferred from the drive shaft to the sliding sleeve by means of that end of the force transmission element which is remote from the supporting element. The above-mentioned constructional features ensure reliable decoupling of drive torque and gas force support, as a result of which both the force transmission element and the supporting element can be of appropriately “slim” construction. 
     For secure torque transfer, there can also be provided between the sliding sleeve, or the drive shaft, and the tilt plate some other arrangement, especially in the form of flattened contact surfaces. It should be mentioned at this point that arrangements with a sliding sleeve and also arrangements without a sliding sleeve are both feasible so that torque transfer can take place, depending on the arrangement, firstly between an appropriate flattened region on the drive shaft and a flattened region corresponding thereto in the tilt plate as well as indirectly or directly by means of corresponding flattened regions between the drive shaft and/or sliding sleeve, on the one hand, and the tilt plate, on the other hand. Secure moment transfer is ensured as a result. 
     The supporting element is optionally mounted in the tilt plate in a cylindrical recess, especially a bore. The bore therein can extend perpendicular to the drive shaft axis. This constructional measure also ensures a simple means of producing a compressor according to the invention. 
     Preferably, the supporting element is secured in the corresponding recess, especially the cylindrical recess, in the tilt plate by means of at least one circlip. Additionally and/or alternatively, the supporting element is secured in the corresponding recess in the tilt plate by means of at least one threaded fastening element. The latter can be a grub screw. It should be noted at this point that a combination of a circlip and a grub screw is also feasible for securing the supporting element. In a constructionally preferred embodiment, the at least one fastening element or especially two fastening elements is/are arranged in the recess in the tilt plate. Alternatively or additionally, the at least one fastening element can be arranged in a radially extending (optionally additional) recess in the tilt plate and can extend into a recess, optionally a further groove, in the supporting element, which recess is arranged on the latter. The groove is preferably arranged along the longitudinal axis of the supporting element. As a result, effective fastening is ensured with a low cost outlay. 
     The force transmission element can, at its end facing the drive shaft, be ring-shaped or sleeve-shaped and, by means of the ring-shaped or sleeve-shaped end, can be mounted on or fastened to the external diameter of the drive shaft or placed over the drive shaft. Alternatively thereto, the force transmission element can also be in operative engagement with a ring-shaped or sleeve-shaped element which is in turn mounted on or in articulated connection with the external diameter of the drive shaft or otherwise in operative engagement with the drive shaft. The force transmission element can, when it is ring-shaped or sleeve-shaped at its end facing the drive shaft, be secured or fastened to the drive shaft in a radial direction by means of a feather key. The same also applies of course if it is a component which is composed of a force transmission element and a ring-shaped element. Furthermore, the force transmission element or the ring-shaped/sleeve-shaped element can optionally be secured to the drive shaft in an axial direction by means of a machine element, especially in the form of a grooved nut or some other axial securing element. Optionally, the ring-shaped or sleeve-shaped end of the force transmission element or the ring-shaped or sleeve-shaped element can have at least one groove extending in an axial direction, in which the device or devices for transfer of the torque engage(s). As already explained, said devices are, in a simple embodiment, drive pins which engage in the corresponding groove. The constructional arrangement described above is a structure which does not impinge on the drive shaft, especially not on the stability of the drive shaft. As a result of the fact that the drive shaft can be of solid construction, on the one hand the radius can be reduced compared to drive shafts through which a bore passes and on the other hand, as a result of the reduction in radius, a weight saving and also a smaller size of construction of a compressor according to the invention can be achieved. 
     Between the supporting element and the force transmission element a length compensation can be provided, which preferably is arranged in an approximately radial direction. In addition to an articulated connection of the supporting element to the force transmission element which allows displacement of the supporting element in a direction perpendicular to the drive shaft axis, this length compensation can also be implemented, alternatively and/or additionally, by a telescopic mechanism or the like. 
     The centre of the articulation or articulated connection between the supporting element and the tilt plate preferably is, for small angles of deflection of the tilt plate, radially further away from the drive shaft central axis than is the centre of an articulation resulting from the articulated connection of the piston(s) to the tilt plate. For large angles of deflection, the geometry is preferably displaced so that the centre of the articulation or articulated connection of the supporting element to the tilt plate moves closer to the drive shaft central axis than the centre of the articulation produced by the articulated connection of the pistons to the tilt plate. Preferably, the two afore-mentioned articulation centres are radially the same distance away from the drive shaft central axis for at least one, especially for exactly one, angle of tilt, or angle of deflection, of the tilt plate. The afore-mentioned structural features ensure that a compressor according to the invention has optimum kinematics. 
     The central axis of the supporting element and/or of the force transmission element forms with the drive shaft central axis, for all angles of deflection of the tilt plate, preferably an included angle which is not equal to 90°. This means, in other words, that the central axis of the supporting element and/or of the force transmission element is arranged at the tilt plate at an angle which for all angles of deflection thereof is greater than 0°. In an optimised constructional form of a compressor according to the invention, the middle angle of deflection of the tilt plate corresponds approximately to half the difference of the maximum angle at which the central axis of the supporting element and/or of the force transmission element is arranged at the tilt plate and the minimum angle at which the central axis of the supporting element and/or of the force transmission element is arranged at the tilt plate. The two afore-described structural arrangements result in an advantageous clearance volume characteristic curve. 
     In a further preferred embodiment, the radial spacing of the articulation between the supporting element and the force transmission element, more precisely the radial spacing of the centre of said articulation, from the drive shaft central axis is, for small angles of deflection of the tilt plate, greater than the radial spacing of the centre of the piston articulations from the drive shaft central axis. The sum of the moments due to the masses moved in translation such as pistons, sliding blocks etc. and due to the masses moved in rotation (tilt plate etc.) is, preferably for all angles of deflection, especially for large angles of deflection and more especially for the maximum angle of deflection of the tilt plate, approximately constant. Preferably, the sum of the moments is constant at 0. This provides the desired regulation characteristic of the compressor according to the invention. 
     In a further preferred embodiment of a compressor according to the invention, the supporting element has, extending in a radial direction, a recess in which the force transmission element is mounted. The recess can especially have an approximately rectangular configuration or rectangular cross-section. Optionally, the cross-section of the recess can, in a radially outer region, become larger in a radial direction towards the outside, whereas it is approximately constant in a radially inner region of the recess. Preferably, those edges of the force transmission element which are located radially on the outside (end face) are arranged in the radially outer region of said recess for any angle of tilt of the tilt plate. In other words, this means that the end-face edges extend beyond the region of the approximately constant recess cross-section. The afore-described structural measures ensure that the force transmission element does not become jammed or blocked in the recess in the supporting element serving for its mounting. 
     In a further advantageous constructional form, a compressor according to the invention has a housing and a drive mechanism chamber substantially defined by the housing, there being arranged between the drive mechanism chamber and the inlet gas side a fluid connection which extends at least partly through the drive shaft. The fluid connection makes possible regulation of the drive mechanism chamber pressure and, accordingly, regulation of the angle of tilt, or angle of deflection, of the tilt plate. The at least partial arrangement of the fluid connection in the drive shaft ensures that the compressor according to the invention has few component parts and is accordingly economical to manufacture. 
     Optionally, the fluid connection comprises, in the drive shaft, at least one recess which extends approximately axially and at least one which extends approximately radially. The recesses can especially be in the form of bores. Preference is further given to at least that end region of the force transmission element which faces the drive shaft being approximately cylindrical, that is to say having an approximately circular cross-section. The force transmission element can be pressed into, or fitted by means of a press fit into, the drive shaft. These measures make possible a structurally simple, yet secure and reliable connection of the force transmission element to the drive shaft. 
     In a preferred embodiment, the drive-shaft-related end of the force transmission element has an especially semi-circular or groove-shaped or groove-like recess which is part of the fluid connection between the drive mechanism chamber and the inlet gas side. As a result, the force transmission element can be centrally mounted in the drive shaft, which prevents imbalances, whereas at the same time the fluid connection between the drive mechanism chamber and in the inlet gas side, which as already mentioned hereinbefore serves for regulation of the angle of deflection of the tilt plate, can be produced in a simple manner and especially without additional component parts. 
    
    
     
       The invention will be described hereinbelow with regard to further advantages and features by way of example and with reference to the accompanying drawings, in which: 
         FIG. 1  shows, in an exploded view, a tilt plate mechanism of a preferred embodiment of a compressor according to the invention; 
         FIGS. 2   a+b  show the preferred embodiment according to  FIG. 1  at a minimum angle of deflection of the tilt plate (a) and at a maximum angle of deflection of the tilt plate (b); 
         FIGS. 3   a+b  show, in diagrammatic form, the possibilities for assembly of the tilt plate of a compressor according to the invention; 
         FIGS. 4   a - c  show, in a diagrammatic representation of a tilt plate mechanism according to the preferred embodiment, a tilting cycle; 
         FIG. 5  shows an example of a clearance volume characteristic curve; 
         FIG. 6  is an overview of the moments due to the masses moved in translation and due to the moment of deviation of the tilt plate and of the sum of the resulting moments, in each case in dependence on the angle of tilt of the tilting plate; 
         FIG. 7  shows, in a qualitative representation, the regulation characteristic curve of the preferred embodiment for a particular operating point and various speeds of rotation; 
         FIGS. 8   a - e  show various possibilities for securing the articulated connection of the force transmission element to the supporting element; 
         FIGS. 9   a+b  show, in a sectional view, a further possibility for articulated connection of the force transmission element to the supporting element; 
         FIG. 10  shows, in a partial sectional view, the mechanism according to  FIG. 9 ; and 
         FIG. 11  shows a drive shaft with a force transmission element of a further preferred embodiment of a compressor according to the invention. 
     
    
    
     The preferred embodiment of a compressor according to the invention comprises (not shown in the drawings) a housing, a cylinder block and a cylinder head. Pistons are mounted in the cylinder block so as to be movable back and forth axially. The compressor drive is provided via a belt pulley by means of a drive shaft  1 . The compressor in the present case is a compressor having variable piston stroke, the piston stroke being regulated by the pressure difference defined by the pressures on the inlet gas side and in the drive mechanism chamber. Depending on the magnitude of the pressure difference, a tilt plate in the form of a tilt ring  2  is deflected, or tilted, from its vertical position to a greater or lesser degree. The greater the resulting angle of tilt, or angle of deflection, the greater is the piston stroke and, therefore, the higher is the pressure made available on the outlet side of the compressor. 
     From  FIG. 1  it can be seen that the tilt plate mechanism of the preferred embodiment comprises: the drive shaft  1 ; the tilt ring  2 ; a sliding sleeve  3 , which is mounted axially on the drive shaft  1  against the action of a resilient element in the form of a ring-shaped or helical adjusting or restoring spring  4 ; and also a supporting element  5  and a force transmission element  6 . The supporting element  5  is in articulated connection with the force transmission element  6  so as to be displaceable both radially and also perpendicular thereto (in a direction perpendicular to the drive shaft axis), which means that the supporting element  5  is mounted so as to be displaceable in a plane (and not just along an axis). The supporting element  5  is cylindrical-pin-shaped and has a groove  7 , by means of which the supporting element  5  is in operative engagement with the force transmission element  6 . For the purpose, that end which faces the supporting element  5 , or that end region which faces the supporting element  5 , of the force transmission element  6  is formed in the shape of a flat steel part. This means therefore that said end region of the force transmission element  6  has an approximately rectangular peripheral contour. That approximately rectangularly formed end region is in engagement with the groove  7  of the supporting element  5 . The advantage of the arrangement of the force transmission element  6  and the supporting element  5  and, especially, the mounting of one thereof inside the other lies in the fact that the flat steel part does not need to be made too high; the strength and rigidity (low deformation) is provided by the width of the mounting. In a middle region, the thickness of the force transmission element  6  increases, whereas at its end facing the drive shaft  1  it is sleeve-shaped. With the aid of the sleeve-shaped part  8  of the force transmission element  6 , the latter is mounted on and fastened to the drive shaft  1 . It should be noted at this point that, in the present preferred embodiment, the force transmission element  6  is formed in one piece with, and also of one and the same material as, the sleeve-shaped part  8 . Alternatively, the force transmission element  6  and the sleeve-shaped part  8  could of course be two different component parts (even, optionally, made of different materials). It should furthermore be noted at this point that the force transmission element  6 , or the sleeve-shaped part  8  of the force transmission element  6 , has two recesses in the form of grooves  9 . 
     Because the internal diameter of the spring  4  is greater than the external diameter of the sleeve-shaped part  8  of the force transmission element  6 , the sleeve-shaped part  8  can, in the assembled state of the tilt plate mechanism, be slid underneath the spring  4 . This means therefore that the sleeve-shaped part  8  is placed over the drive shaft  1  and is fixed on the drive shaft  1  radially by means of the spring  4 . On that side of the force transmission element  6  which is remote from the spring  4  there is then placed over the drive shaft  1  the sliding sleeve  3 , which has a recess  10  corresponding to the force transmission element  6 . The sliding sleeve  3  furthermore has two recesses in the form of bores  11 . The force transmission element  6  and the sliding sleeve  3  are secured on the drive shaft  1  axially by means of a grooved nut  12  (see  FIG. 2 ). For better start-up behaviour of the compressor there is furthermore arranged on the drive shaft  1  a disc spring  23  which ensures that the compressor does not start up with the angle of deflection of the tilt ring  2  being minimal. Furthermore, there are arranged on the drive shaft  2  end-stops in the form of end-stop discs  24 ,  25 , which limit the angle of deflection of the tilt ring. The end-stop disc  24  serves as an end-stop for a minimum angle of deflection, whereas the end-stop disc  25  serves as an end-stop for a maximum angle of deflection of the tilt ring  2 . 
     The supporting element  5  is mounted in a cylindrical recess in the form of a bore  13  in the tilt ring  2 . The bore  13  extends perpendicular to the drive shaft axis. The supporting element  5  is secured in the tilt ring  2  by means of two circlips  14 . 
     It should be noted at this point that the force transmission element  6 , which in the present preferred embodiment is connected to the drive shaft  1  in a manner which does not allow relative rotation, can also, in other embodiments, be in rotatable operative connection therewith. It should furthermore be noted at this point that, as a result of the sleeve-shaped arrangement, or sleeve-shaped part  8 , of the force transmission element  6 , the drive shaft  1  is not penetrated by a hole and accordingly has appropriate stability. The open width of the bore in the tilt ring  2  is at least very slightly larger than the corresponding extent of the force transmission element  6 . 
     In the present preferred embodiment, the mechanism comprising the supporting element  5  and the force transmission element  6  is not intended for transferring the torque from the shaft to the tilt plate in the form of the tilt ring  2 . The mounting locations between the supporting element  5  and the force transmission element  6 , between the force transmission element  6  and the drive shaft  1  and between the supporting element  5  and the tilt ring  2  are not designed for the purpose of transferring torque. Accordingly, a kind of movement transmission function does not fall on the supporting element  5  and the force transmission element  6 , this having been deliberately so decided for reasons of hysteresis, that is to say the tilting of the tilt ring  2  and the transfer of torque are functionally decoupled from one another. The mechanism comprising the force transmission element  6  and the supporting element  5  basically take up the piston forces. For its part, the torque is transferred from the drive shaft  1  to the tilt ring  2  by a tilt-providing articulation (implemented by drive pins  15 ) provided on the drive shaft central axis. The drive pins  15  transferring the torque between the sliding sleeve  3  and the tilt ring  2  are retained, or secured, in the tilt ring by means of circlips  16 . The tilt ring  2  has flattened regions  17 , which correspond to flattened regions  18  on the sliding sleeve  3 . In principle it is also feasible in other embodiments for the sliding sleeve  3  to be omitted and for the torque to be transferred directly between the drive shaft and tilt ring  2  in any desired form (for example, by means of flattened regions on the drive shaft  1  and the tilt ring  2 ). 
     The decoupling of the torque transfer and gas force support makes it possible, in addition to the possibility of making the supporting element  5  and the force transmission element  6  of appropriately small dimensions, to achieve optimised surface contact pressure, especially between the force transmission element  6  and the supporting element  5  and between the supporting element  5  and the tilt ring  2 . As a result thereof and as a result of the mode of construction, specific to the invention, of the supporting element  5  and the force transmission element  6 , or the articulated connection, specific to the invention, between the force transmission element  6  and the supporting element  5 , a compact constructional form of compressor can be achieved. 
       FIGS. 2   a  and  b  show the preferred embodiment of the compressor according to the invention again, in the assembled state, for an angle of minimum deflection ( FIG. 2   a ) and for an angle of maximum deflection ( FIG. 2   b ). In  FIG. 2   a , “V” indicates the position of the grooved articulation, that is to say the position of the articulation formed by the supporting element  5  and the force transmission element  6 , whereas “U” indicates the position of the piston. The configuration, in accordance with the invention, of the supporting element  5  and the force transmission element  6  allows both for the radius U to correspond to the radius V and for U to be smaller than V and also for U to be larger than V. The large number of degrees of freedom results in a low degree of Hertzian stress and also in a low degree of wear because no jamming occurs. 
       FIG. 3   a  shows the assembly of the preferred embodiment of the compressor according to the invention. Because the diameter of the force transmission element  6  including the sleeve-shaped part  8  is larger than the bore in the tilt ring  2 , the tilt ring  2  is placed in a slanting position over the force transmission element  6  and is then moved into a position perpendicular thereto, as a result of which the force transmission element  6  is moved into the recess  13 . If, as in a further preferred embodiment of a compressor according to the invention which is not described in further detail herein, the diameter of the bore in the tilt ring is larger than the diameter of the force transmission element  6  (see  FIG. 3   b  in this regard), the tilt ring  2  can be placed over the force transmission element  6  in a perpendicular orientation and, by means of a sideways movement, brought into operative engagement therewith or with the supporting element  5 . 
       FIG. 4   a  shows, in diagrammatic form, a tilt plate unit of a compressor according to the invention for large angles of deflection;  FIGS. 4   b  and  4   c  furthermore show details of  FIG. 4   a  for angles of deflection of the tilt ring  2  which differ from  FIG. 4   a  (medium and small angles of deflection). The compressor kinematics take into account the position of the sliding blocks of the pistons by means of the centre at C and the position of the supporting element  5  at B. The spacing between C and B is a snapshot view which is dependent on the angle of deflection. Large angles of deflection (see  FIG. 4   a ) result in a position of the tilt ring  2  wherein the centre B of the supporting element  5  in the tilt ring is located to the inside of the circular cylinder b on which the piston central axes are located. At a medium angle of deflection (see  FIG. 4   b ), the centre B of the supporting element  5  in the tilt ring  2  coincides with the centre of the piston articulation or the circular cylinder on which the centres of the piston articulations are located. Finally, in  FIG. 4   c  (which corresponds to a small angle of tilt), the centre B of the supporting element  5  is located radially to the outside the circular cylinder b, on which the piston central axes are located. In the preferred embodiment, it is therefore possible both for B and C to coincide and for it to be the case that B comes to be located to the left or right of an axis b, which passes through C. In detail, the reference letters in  FIGS. 4   a  to  4   c  have the following meanings:
     A articulation (centre) of the tilt ring  2  on the drive shaft guideway;   B articulation (centre) of the supporting element  5  in the tilt ring  2 ;   C centre of the piston articulation (sliding block) for the piston which is located in the upper dead centre position;   D point of intersection of the axis of the supporting element  5  and the drive shaft axis;   E point of intersection of the drive shaft central axis and its perpendicular to the centre C;   F point of intersection of the drive shaft central axis and its perpendicular to the centre B;   G point of intersection of the axis b and the axis e;   a drive shaft central axis;   b central axis of the piston and of the cylinder for the piston (cylinder) which is located in the upper dead centre position (five, six or seven pistons are usually used);   c centre-line of the tilt ring  2 , on which line the centres B and C are (preferably) located;   d supporting element central axis and/or force transmission element central axis;   e perpendicular from the drive shaft central axis to the centre B;   f perpendicular from the drive shaft central axis to the centre C;   α angle of tilt of the supporting/force transmission element (constant; structurally determined);   β angle of tilt of the tilt ring  2 .   
     If the tilt ring  2  is tilted further relative to a starting position, the centre-point of the articulation B moves towards the drive shaft central axis a; as a result, the line B-D becomes shorter. The degree of freedom required for the variation in the length of the line is obtained from the articulated connection, in accordance with the invention, of the supporting element  5  to the force transmission element  6 . The line B-D is the hypotenuse of the triangle BDF. The catheti D-F and F-B of the right-angled triangle likewise become shorter. The line D-F is of great importance to the clearance volume, which normally should be minimised. 
     Also of importance is the (right-angled) triangle BCG. If the centre-point of the articulation B moves towards the drive shaft A, the triangle BCG becomes larger (converse effect to the triangle BDF, which becomes smaller). In addition to the line D-F, the line C-G is also of importance to the clearance volume. In relation to the two afore-mentioned lines, two contrary effects occur on tilting of the tilt ring  2 , which can be advantageously used for mutual compensation, provided that the parameters are appropriately selected. In that context it is advantageous if the supporting element  5  and/or the force transmission element  6  is/are arranged at an angle α (which is not equal to 0°) in order to produce an effect at all (triangle BDF), which acts counter to the effect due to the triangle BCG. The effect due to BCG can be influenced only—it can never be ruled out (as is also the case with BDF) because the mounting points C and B move relative to one another depending on the angle of tilt. The resulting clearance volume for the particular arrangement is proportional to the plot of the “CG+DF” curve.  FIG. 5  shows the information relating to the clearance volume characteristic curve for an angle of the gas force support means (composed of the supporting element  5  and the force transmission element  6 ) of 9° included between the latter and a perpendicular to the drive shaft central axis. The diagram of  FIG. 5  is to be regarded solely as an example because other clearance volume characteristic curves may be desirable depending on the application. In principle, by appropriately selecting parameters it is possible to obtain a very different behaviour depending on the requirements of the user, although in most cases minimisation of the clearance volume is desirable. 
       FIG. 6  shows the moment distribution of the preferred embodiment of the compressor according to the invention, from which it can be seen that the sum of the moments due to the masses moved in translation and of the moments due to the moment of deviation of the tilt plate almost balances out over the entire tilt angle range of the tilting plate or tilt ring  2 . Attention should be drawn especially to the equalisation of moments for high angles of tilt between 16° and 18°. This moment distribution results in a regulation characteristic curve as is shown in  FIG. 7  for a particular operating point and for various speeds of rotation n. From  FIG. 7  it can be seen that the regulation characteristic curve is very similar for different speeds of rotation n, which results from the optimised moment distribution of the compressor. 
       FIGS. 8   a  to  e  finally show various possibilities for securing the supporting element  5  in the corresponding recess (bore  13 ) in the tilt ring  2 . Securing the supporting element  5  in the bore  13  is necessary especially because of the centrifugal forces that come into effect. In the above-described preferred embodiment, securing is performed by means of two circlips  14 . As an alternative thereto, as indicated in  FIG. 8   a , securing by means of a combination of one circlip  14  and a grub screw  19  which engages in a corresponding thread provided in the bore is also feasible. Of course it is also feasible for the supporting element  5  to be secured on both sides of the bore by means of a grub screw. Because the bore has a disadvantageous effect on the inertia of mass of the tilt ring  2 , the additional introduction of the mass of the grub screw  19  (which has a large mass compared to a circlip  14 ) can advantageously influence the inertia of mass of the tilt ring  2 . It is feasible for the major part of the bore  13  not occupied by the supporting element  5  to be plugged with the grub screw. Alternatively or additionally, stoppers at the ends of the bore are also feasible. These can be made of a material that differs from the material of the tilt ring  2 , especially of a heavier material, in order by that means to compensate the missing inertia of mass. 
       FIG. 8   b  shows a further possibility for securing the supporting element  5 . In this alternative, a through-hole through the tilt ring  2  is not provided but rather just a blind hole  20 . After the supporting element  5  has been put in place, it is secured by means of a circlip  14 . Alternatively thereto, of course, it is again possible to provide the opening with a grub screw and/or an appropriate stopper. 
     As can be seen from  FIG. 8   c , it is also feasible for the supporting element  5  not to have an uninterrupted groove but rather just a kind of pocket  22 , that is to say a recess located in the central region of the supporting element. As a result, there is basically no longer any need for a securing mechanism as such and the bore  13  can, if required, be plugged at the sides with bungs of any desired material. An arrangement of such a kind is to be recommended especially when assembly in accordance with  FIG. 3   b  is or can be carried out. 
     As can be seen from  FIG. 8   d , different shapes of groove or pocket also come into consideration. The recess or pocket  22  shown in  FIG. 8   d  is, for example, produced by simple means using a disc milling cutter, which ensures simple manufacture. 
       FIG. 8   e  finally shows a further possibility for securing the supporting element. In this case, there is produced in a radial direction in the tilt ring  2  a further, threaded recess, into which a grub screw  21  is introduced. The grub screw projects out beyond the radially outer rim of the bore  13  and engages in a corresponding groove provided on the supporting element  5  so that the supporting element  5  is secured against slipping in the bore  13 . 
       FIGS. 9   a  and  9   b  show a further preferred embodiment of a tilt plate mechanism of a compressor according to the invention. Like the tilt plate mechanism shown in  FIG. 1 , this mechanism comprises the tilt ring  2  and also the sliding sleeve  3  which is axially mounted on the drive shaft  1 . Details will be given hereinbelow only of those features which distinguish the mechanism according to  FIGS. 9   a  and  9   b  from the mechanism according to  FIG. 1 . It should be noted at this point that the tilt plate mechanism in  FIG. 9   a  is shown for a minimum angle of deflection of the tilt ring  2 , whereas  FIG. 9   b  shows the tilt plate mechanism at a maximum angle of deflection of the tilt ring  2 . 
     In contrast to the embodiment according to  FIG. 1 , the supporting element  5  (in this respect see also, especially,  FIG. 10 ) does not have a groove-shaped or pocket-shaped recess but rather an approximately rectangular recess  26 , which extends through the whole of the supporting element  5  in a radial direction. The force transmission element  6 , which as already mentioned hereinbefore is formed like a flat steel part at its radially outer end, engages in the recess  26  and accordingly forms the articulated connection of the force transmission element  6  with the supporting element  5 . 
     In a radially outer region (indicated by arrows  27  in  FIGS. 9   a  and  9   b  and arrows  28  in  FIG. 10 ), the cross-section of the recess  26  becomes larger towards the outside radially (in this respect see the cross-sections in  FIGS. 9   a  and  9   b ), from which it can be seen that the cross-section in said radially outer region flares in an approximately V-shape, whereas in a radially inner region it is approximately constant (in this respect see also, especially,  FIGS. 9   a  and  9   b ). The radially outer end-face edges of the force transmission element  6  are, for all angles of tilt of the tilt ring, arranged in the radially outer region (in this respect see  FIGS. 9   a  and  9   b ). This means that the edges  29  project out from the region of the approximately constant recess cross-section at all angles of the tilt plate. As a result, jamming of the edges  29  with the supporting element  5  is avoided at every operating point of the compressor. Friction and hysteresis in the region of the mechanism are reduced thereby, which in use results in a low degree of wear and a low degree of heat generation. 
       FIG. 11  shows a further preferred embodiment of a compressor according to the invention, wherein in this preferred embodiment the force transmission element is not part of a sleeve or not connected to a sleeve but rather is pressed into the drive shaft. As is also the case in the preferred embodiment described in  FIG. 1 , the force transmission element is, at its radially outer end facing the supporting element, formed in the shape of a flat steel part. However, at the end facing the drive shaft or in a region facing the drive shaft, the force transmission element  6  is formed in a cylindrical shape, although in the region of the drive shaft, that is to say where the force transmission element is pressed into the drive shaft, said element has an approximately semi-circular or groove-like recess  30 . 
     The recess  30  in the force transmission element  6  is part of a fluid connection  31  between the drive mechanism space or drive mechanism chamber of the compressor and the inlet gas side. This fluid connection serves for regulation of the pressure in the drive mechanism chamber and, accordingly, for regulation of the piston stroke. The fluid connection extends through the drive shaft  1  and comprises, in addition to the recess  30 , a recess in the drive shaft, which recess extends approximately axially and is in the form of a bore  32 , and a recess which extends approximately radially and which is likewise in the form of a bore  33 . It should be noted at this point that an arrangement of such a kind is not limited to the connection of two chambers, especially the above-mentioned chambers, but that as a result thereof fluid connections between any desired chambers or volumes or regions are possible. 
     At its inlet gas end, the bore  32  opens into a cylindrical recess  33  (bore). An arrangement of such a kind ensures that oil is separated out well in the fluid connection between the drive mechanism chamber and the inlet gas side. On the one hand, as a result of the centrifugal forces, the oil that is present in the drive mechanism chamber, wherein oil mist lubrication takes place, is spun out of the connection in the direction of the drive mechanism chamber. On the other hand, the oil that reaches the region of the axially extending bore  32  is separated out on the wall of the bore  32  by means of the centrifugal forces and it can then flow back to the bore  33 . Further bores in the drive shaft in a radial direction are feasible, especially in the region of the mounting of the drive shaft, as a result of which the oil collecting on the walls of the bore  32  can escape into the bearings and accordingly provide the latter with optimum lubrication. Finally, it can also be seen from  FIG. 11  that an end-stop disc  35  is fitted to the drive shaft  1  in order to limit the maximum angle of tilt, or angle of deflection, of the tilt ring  2 . 
     At this point the materials used for manufacture of the tilt ring  2  and of the pistons should be discussed briefly. In the above-described embodiments, the tilt ring  2  is made of steel and provided with a coating which minimises wear and friction between the sliding blocks of the pistons and the tilt ring  2 . Alternatively, the tilt ring  2  can also be made of brass or bronze. The mentioned materials ensure that the requirements brought about by the constructional form are met. The tilt rings  2  that are used are in fact rings whose height is much greater than in the prior art. The height is desirable on the one hand in order that the gas force support means, which is composed of the supporting element  5  and the force transmission element  6 , can be mounted therein; on the other hand, the height is advantageous in order to provide the component with a sufficient inertia of mass. This is necessary in order to be able to generate a tilting moment—due to the centrifugal effect on rotation of the tilt ring  2 —which is sufficiently large to be able, to the desired extent, to compensate or over-compensate the contrarily acting tilting moments due to the mass forces of the pistons. 
     For tilt rings  2  of such a kind, the materials steel, brass or bronze are, as mentioned, especially suitable because, by virtue of the height of the tilt ring  2 , these materials ensure sufficient strength and rigidity to be able to avoid deformation. In the case of tilt rings according to the prior art, this is frequently not ensured. Furthermore, the density of bronze or brass may, depending on the alloy, be somewhat greater than the density of steel or that of grey cast iron (a tilt ring  2  according to the invention can of course also be made of grey cast iron). The density increase, or higher density, of bronze or brass can be utilised in order to be able to compensate or over-compensate the piston masses even better. The height of the tilt ring  2  results in the fact that the pistons, which in the Application being discussed herein engage around the tilt ring  2  and are mounted on the latter by means of two sliding blocks, must have a large opening in order to engage around the tilt ring  2 . In the preferred embodiment, in which the tilt ring  2  is made of brass, the pistons are made of an aluminium alloy. Because brass has similar thermal expansion to aluminium, such a combination of materials ensures reduced wear and an increased service life for a compressor according to the invention, because the play of the sliding blocks in the pistons will (when heat is generated in use) be only insubstantially increased, or not at all increased, compared to the state on assembly. This results in a low degree of noise generation and prevents the sliding blocks from being able to drop out as a result of excessive play. If the tilt ring  2  is made of steel, pistons which are likewise made of steel offer the same advantages. Alternatively, however, other combinations of materials are also feasible (especially with a view to reducing the weight of a compressor according to the invention). 
     A mechanism as described hereinbefore, that is to say a tilt plate mechanism which comprises a tilt ring, is suitable especially for a compressor in which R744 (CO 2 ) is used. It can of course also be used for coolants such as R134a, R152a etc. and also for coolants as are mentioned in U.S. Pat. No. 6,969,701 and WO 2006/012095 (e.g. azeotropic mixtures of tetrafluoropropene and trifluoroiodomethane). Tilting plates like those described hereinbefore, which are relatively tall in height, are frequently operated with sliding blocks providing for the articulated connection of the pistons where the spherical cap of the sliding blocks has a (very) small radius of curvature. This is due to the fact that relatively large spherical diameters are used for the sliding blocks (tilting plate height+2×height of sliding blocks=diameter). In order to effectively prevent the sliding blocks from being pushed out of the piston caps because of excessive tilting (given the low degree of curvature of the sliding blocks), a maximum angle of deflection, or tilt, of the tilt ring of from 15.5° to 17.5° is preferred. 
     Although the invention is described using embodiments having fixed combinations of features, it nevertheless also encompasses any further feasible advantageous combinations of those features, as are especially but not exhaustively mentioned in the subordinate claims. All features disclosed in the application documents are claimed as being important to the invention insofar as they are novel on their own or in combination compared with the prior art. 
     REFERENCE NUMERAL LIST 
     
         
           1  drive shaft 
           2  tilt ring 
           3  sliding sleeve 
           4  spring 
           5  supporting element 
           6  force transmission element 
           7  groove 
           8  sleeve-shaped part of the force transmission element  6   
           9  groove 
           10  recess in the sliding sleeve  3   
           11  bore 
           12  grooved nut 
           13  bore 
           14  circlip 
           15  drive pin 
           16  circlip 
           17 ,  18  flattened region 
           19  grub screw 
           20  blind hole 
           21  grub screw 
           22  pocket 
           23  disc spring 
           24 ,  25  end-stop (disc) 
           26  recess 
           27 ,  28  arrow 
           29  radially outer edge of the force transmission element ( 6 ) 
           30  recess in the force transmission element ( 6 ) 
           31  fluid connection 
           32 ,  33  bore 
           34  cylindrical recess 
           35  end-stop disc