WO2008025412A1

WO2008025412A1 - Axial piston compressor

Info

Publication number: WO2008025412A1
Application number: PCT/EP2007/006397
Authority: WO
Inventors: Otfried Schwarzkopf
Original assignee: Valeo Compressor Europe Gmbh
Priority date: 2006-08-30
Filing date: 2007-07-18
Publication date: 2008-03-06
Also published as: DE102006040490A1

Abstract

An axial piston compressor, particularly a compressor for air conditioning systems of motor vehicles, having at least one piston and in particular an annular captive C-washer (2), which can be adjusted with regard to the inclination thereof to a drive shaft (1), is rotationally driven by the drive shaft (1), and which is connected in an articulated manner to at least one supporting element (5) that rotates with the drive shaft (1), wherein the piston(s) each has/have a joint assembly against which the captive C-washer (2) is in sliding engagement, wherein the U/V ratio between the distance U of the drive shaft center axis from the divided circle radius of the shaft center axis (axes) and the distance V of the drive shaft center axis from the center of the pivoted joint, which results from the linkage of the supporting element (5) to the captive C-washer (2), is always smaller than 1 regardless of the size of the angle of deflection of the captive C-washer (2).

Description

, Axial piston "

Description

The invention relates to an axial piston compressor according to the preamble of claim 1.

From DE 103 15 477 A1, which goes back to the Applicant, an axial piston compressor with a swivel disk is known, the same being of annular design. The swash plate or the swivel ring has a sufficient thickness or wall thickness, so that a driver head or support element can be placed in a recess in the swivel ring. The driver has the task to transmit the torque from a drive shaft to the swivel ring. Another function of the driver is to support forces that occur in the compressor due to the gas compression. In principle, it is also conceivable that the driver only fulfills one of these two functions, that is, for example, serves only for gas power support.

The control of the above-described axial piston compressor is essentially borrowed by the equilibrium of forces on the swashplate. In general, the following moments influence the tilting behavior of the swashplate in the center of tilting movement (the direction of the moment is indicated in parentheses, with (-) decreasing, ie in the direction of the minimum lift, and (+) pending, ie in the direction of the maximum lift, means):

Moment due to the gas forces in the cylinder chambers (+) Moment due to the gas forces from an engine room (-) Moment due to a return spring (-) and, if necessary, moment due to a positioning spring (+) Moment due to the rotating masses (-), eg swashplate (including moment due to the center of gravity, this proportion can be (+)) Moment due to the translationally moving masses (+), eg pistons, sliding blocks and / or an oscillating swash plate.

By varying the pressure in the engine room can be in such a compressor according to the prior art, the piston stroke and thus regulate the refrigerant mass flow. If the swivel ring is to be adjusted to a smaller deflection angle under otherwise identical operating conditions (for example, at the same speed), the pressure in the engine room is increased. If the swivel ring is to be adjusted to a larger deflection angle under otherwise identical operating conditions, the pressure in the engine compartment is reduced.

With regard to the prior art, reference may be made to the following publications: DE 101 24 034 A1, DE 101 24 033 A1, DE 101 24 034 A1, DE 197 49 727 A1, DE 100 101 32 A1 and DE 102 27 608 A1 , From the above-mentioned documents further engines of the swash plate type are known. All of these engines have as a common feature the position of the support element or the support elements exactly in the center of the sliding blocks of the piston. The advantage of such a design is that during compaction, a dead space independent of the deflection angle is created, i. The distance between piston end faces and valve plates is constant for all possible deflection angles. It should be noted at this point that a dead space in such a construction is unavoidable. An enlargement of the dead space as a function of the swash plate deflection angle can be avoided by the measures explained above.

It should also be mentioned at this point the DE 190 10 132, in which a driver assembly, which consists of two drivers, is proposed.

Furthermore, axial piston compressors are known from the documents DE 10 2005 018 102 A1 and DE 10 2004 041 645 A1, which are based on the applicant.

If one compares the constructions of the two aforementioned publications with the constructions explained above, then one essential difference lies in the fact that at least one further degree of freedom or a bearing is introduced in the force transmission between the shaft and the thrust washer. This means that the driver consists of a support element (as a driving head) and a power transmission element. However, in the compressors according to the two aforementioned publications, the dead space can not be kept exactly constant. There is a small variation of the dead space as a function of the deflection angle of the swash plate.

The support element is in the designs for a given deflection angle depending on the design of the compressor outside, within or on the pitch circle diameter of the piston center axes (which correspond to the pitch diameter of the sliding blocks of the piston usually) and shifts when tilting the swash plate from small to large tilt angles of outside to inside (theoretically, a variant of "inside out" would be conceivable).

The advantage of the constructions going back to the applicant is that the joints are optimized with respect to the surface pressures occurring in the bearing positions (pin or tongue and groove joints). In all the above-mentioned engine designs of the swash plate design as well as in here not further discussed compressors for the refrigerant Rl 34a occurs as a problem that in the region of small deflection angle a maximum in the control characteristic can occur, i. that from this point for the further reduction of the deflection angle of the engine room pressure would have to be lowered. One of the main causes of this Hegt im Momentengleichgewicht on the swash plate. As a further (inevitable) cause of the maximum, the dead space is known.

In terms of moment equilibrium, the gas forces are essentially decisive. A schematic of the gas forces acting on the swashplate is shown in FIG. The forces on the swash plate are assigned to the compression process. It is assumed that a compressor with seven axial piston, in which the tilting joint about which the swivel mechanism is pivoted, is arranged on the underside of a disc. For example, a swash plate can be arranged on this disk. Such a compressor is known from US 5,894,782 A. At a

Compressor of Schwenkringbauart this tilting joint would be arranged in the intended distance to the drive shaft center axis within a cavity of the pivot ring.

The forces acting as a result of p _d - p _c (p _d : pressure on the high pressure side, p _c : pressure in the engine compartment) from one side of the swash plate and the forces p _c - p _s (p _s :

Pressure on the suction side) from the other side of the swash plate. The tilting joint of Swivel disk is well away from the drive shaft centerline, approximately in the area of the piston, which is in the top dead center.

If the same is reduced on the basis of a large swash plate deflection angle, the engine room pressure must be increased. As a result, the pressure difference p _d - p _c becomes smaller and the pressure difference p _c - p _s larger.

As already mentioned, the tilting characteristic can have a maximum, and from this point the engine room pressure would have to be lowered for a further reduction of the deflection angle. The position of the maximum is influenced by the engine geometry, in particular by the position of the tilting joint. The maximum is the more pronounced and shifts to larger deflection angles, the greater the ratio between the pitch circle radius of the piston center axis (center of the piston sliding blocks) and the distance of the tilting joint to the drive shaft center axis. The tilting joint is comparable with the driver head or support element.

The maximum is the smaller and shifts to smaller angles of tilt or disappears completely, the smaller the ratio between pitch radius of the piston center axis and the distance of the center of the tilting joint to the drive shaft center axis. For both cases, the tilting joint is comparable with the driver head or support element.

Comparing engines of the conventional type, as they are e.g. For Rl 34a, with newer type swashplate engines, it can be said that the tendency for such maxima is generally more pronounced in conventional engines.

This is because the tilting joint in these engines with increasing deflection angle of a larger distance from the drive shaft moves to a smaller distance from the drive shaft back. In this case, due to the axially considerable size (the tilting joint is arranged below the swash plate) in the conventional swashplate engines a significantly greater way available and travels than in the engines, which have an annular swash plate.

In the case of swivel ring engines with a constant dead space, the ratio between pitch circle radius of the piston center axis and the distance of the tilting joint from the drive shaft center axis is 1 for all tilt angles. In the swivel ring engines with variable position of the support element, of course, a ratio of 1 is also desirable, but not feasible. There is a slight variation in the situation, which, depending on the operating parameters, is around 1 mm to 1.5 mm. For a given deflection angle, the above-mentioned ratio of 1 is adjustable.

Swash plate compressors with tilting joints disposed below the swashplate (for example, according to US 5,894,782 A) cause a variation of the tilting joint position partly by several millimeters. The ratio becomes greater than 1 and varies considerably. Especially at small tilt angles, the ratio is often significantly greater than 1.

It should be noted at this point that a separation of the functions of the transfer of torque and the support of the gas forces is conceivable. The torque may e.g. be transmitted via a further connection (drive pin) between the drive shaft and the swash plate. It is known in this context that the swash plate is pivotally mounted on a guide bush, which in turn is guided displaceably on the drive shaft in the axial direction. Such a construction has the advantage that it is more stable overall and / or can be made more compact.

Another advantage is that in addition to the tilting moment in the one plane, a torque in a further plane, which is arranged perpendicular to the first plane, can be better supported. In this connection, reference is again made to Fig. 15, in which the piston forces are shown schematically on the swash plate and can be easily illustrated by the above explanations.

As mentioned above, the center of the supporting joint (s), which may be approximately circular or spherical, cylindrical or barrel-shaped in section of the plane, coincides with the center of the piston joint. The center of the piston joint is defined by two sliding blocks, which together with the swashplate describe a circle. The center of this circle is in the parlance of the present application, the center of the piston joint. As a rule, the center of this piston joint lies on the central axis of the piston. The centers of the various pistons (usually between 5 and 9 pistons) describe a circle, which is referred to in the parlance of the present application as Kolbenteilkreisdurchmesser. The advantage offered by the fact that the center of the support joint is in the center of the piston is particularly energetic in nature, since the compressor due to a minimized dead space regardless of the deflection angle of the swash plate can provide a high and consistent efficiency (the dead space does not vary in Dependency of the deflection angle).

It is also conceivable to provide only one support joint for receiving in the swash plate, wherein the support joint is preferably located where the maximum piston force occurs. This usually occurs when the valve of the respective cylinder chamber opens under pressure control. This happens depending on the operating point usually before the piston in question reaches its top dead center. From Fig. 15 it can be seen that this time is approximately between the piston position 3 and 4. This means that the maximum piston force can be optimally intercepted when the support joint is arranged approximately 30 ° to 45 ° before the 0 ° position (top dead center position). As a result, in particular a perpendicular to the tilting movement suppressing twisting torque is absorbed, on the other hand at the same time also the tilting moment is partially absorbed. A portion of the tilting moment is also required to safely deflect the swashplate at all operating points can.

With reference to DE 190 101 32 Al, it can be stated that the position of the head of the support element shown there has to be arranged on a larger diameter, as long as the center of the head is to be positioned approximately 45 ° to 30 ° before the top dead center position. By this measure, the center of the tilting movement still coincides with the center of the joints of the piston. Depending on the selection of geometrical parameters, the arrangement of the head of the support member on a larger diameter may cause the head of the support member to protrude outwardly from the bore and opening of the swashplate. In particular, when the swash plate is pivoted from a small deflection angle to a larger deflection angle, the head of the support element moves out of the swash plate, with the result that the forces can only be transmitted in an edge region of the swash plate. This has, for example, deformations on the edge of the swashplate result.

Another shortcoming lies in the compressors according to the prior art in the field of control behavior or the tilting or Auslenkcharakteristik the swash plate. The deflection angle depends on the pressures acting on the swashplate (indirectly via the pistons). The tilting characteristic (representation of the engine room pressure p _c as a controlled variable over the swashplate deflection angle) is particularly strongly influenced by the pressure p _D on the high pressure side, which is undesirable.

Starting from the above-described prior art, it is an object of the present invention to provide a compressor which has an improved control behavior and in which in particular control ranges are avoided in which a maximum occurs, i. where more than one swashplate tilt angle is assigned to an engine room pressure. This task is intended for a simultaneous good Kraftbzw. Torque can be achieved between the support member and the swash plate, with deformations due to the transfer force to be avoided, while the dead space characteristic has an acceptable profile.

This object is achieved by a compressor with the features of claim 1, wherein preferred developments and embodiments are described in the subclaims.

An essential point of the invention is therefore that the ratio U / V between the distance U of the drive shaft center axis from the pitch radius of the piston center axis (n) and the distance V of the drive shaft center axis from the center of the tilting joint, which Linkage of the support member results on the swash plate, regardless of the size of the deflection angle of the swash plate is always less than 1. This ensures optimal power transmission between the support element or head of the support element and the swash plate with a good control behavior and an acceptable dead space characteristic.

Preferably, the ratio U / V decreases steadily with increasingly larger deflection angle of the swash plate accordingly. This avoids maxima in the control characteristic.

In a further preferred embodiment of a compressor according to the invention, the ratio U / V is about 0.93 to 0.97 for small deflection angles of the swash plate (about 0 ° to 2 °) and about for large tilt angles (about 15 ° to 20 °) 0.7 to 0.75. The parameter set described above provides optimum conditions for the operation of an axial piston compressor.

Optionally, the ratio U / V * between the distance U as defined above and the distance V * between the radially outer end of the support element and the Drive shaft center axis for all deflection angle of the swash plate be greater than 1. This ensures that the radially outwardly directed end of the support element, ie in particular optionally a head of the support element, is mounted for all deflection angles in a corresponding recess in the swashplate and can not protrude beyond the outer edge thereof. Thus, in addition to a continuously constant transfer of torque ensures that the compressor, in particular the swash plate, undergoes no damage or deformation.

The drive shaft central axis of the compressor according to the invention and the center axis of the support element can form an angle that is less than 90 °. Preferably, the angle is 80 ° to 85 °. By this inclined with respect to the drive shaft position of the support member, a compact stable design can be ensured.

The alignment of a recess in the swashplate, in which the radially outer portion, in particular head of the support element engages, may deviate from a radial direction with respect to the drive shaft center axis.

In a further preferred embodiment, the central axis of the support element and the center axis of the recess in the swash plate, in which engages the radially outer portion, in particular head of the support member, and in which it is mounted, form an angle which is smaller than 180 °. Preferably, the angle defined above is in a range of 165 ° to 175 °. By such a construction, a further reduction of the dead space is possible.

Alternatively, the recess in the swashplate, in which the radially outer, in particular cylinder, barrel, drop or spherical portion of the support element (in particular head) is mounted, be arranged in the radial direction relative to the drive shaft central axis. Such an arrangement is easy to manufacture and thus inexpensive to manufacture.

The recess in the swash plate may be formed as a bore, the same at its (in the radial direction) inside end may have a region in which a further, also cylindrical recess, which in particular approximately in the radial direction (ie at an angle of about 90 ° to the drive shaft center axis), be superimposed. A cylindrical recess of the swash plate is easy to manufacture, a further superimposed recess is used in particular in the case in which the support element is inclined against the drive shaft central axis, for an unhindered derten optimal carry the torque and / or ensures optimal support of the gas forces.

The wall sections which are defined by the recess in the swashplate and which extend on both sides of the swashplate in the axial direction, preferably have a different thickness in the axial direction. The different thickness is formed such that the material concentration of the swash plate on the higher load side is greater than on the lower load side. In the case of a radially extending recess in the swash plate, this means that it is not arranged concentrically with the (radially extending) center axis of the swash plate, but with a recess center which deviates from the central axis. This makes it possible to obtain on the more heavily loaded side of the swash plate a thicker wall thickness and thus a more stable construction. The wall thickness on the less loaded side is lower and thus leads to a weight-optimized design of a compressor according to the invention.

The object of the invention is further achieved by a compressor in which the center of the tilting joint is offset axially relative to the center plane of the swash plate, in particular to the side facing away from the piston. This feature is unique, as well as in combination with the features described above

Solution of the task according to the invention, which is simple and inexpensive to manufacture.

The invention will be described below with reference to further advantages and features by way of example and with reference to the accompanying drawings. The drawings show in:

Fig. Ia and b the swivel mechanism of a first preferred

From a guide shape of the compressor according to the invention at a deflection angle of 0 ° (Fig. Ia) and a mean deflection angle (Fig. Ib);

Fig. 2 shows the swivel mechanism of the first preferred

Execution form in exploded view;

3 shows the dead space characteristic of the first preferred embodiment; 4 is a schematic representation of the swash plate mechanism of a compressor according to the invention;

Fig. 5 and 6 control characteristics for the preferred embodiment of FIG. 1;

7 a representation of the moments as a result of translationally and rotationally moved masses as a function of the tilting angle of the swashplate;

8 is a qualitative representation of the control characteristic of a second preferred embodiment of a compressor according to the invention;

9a-9c show the second preferred embodiment in cross-section at a deflection angle of 0 ° (9a) and a maximum deflection angle (FIG. 9b) and the resulting dead space characteristic (FIG. 9c);

Fig. 10a-10c a third preferred embodiment of a compressor according to the invention in illustrations analogous to FIGS. 9a-c;

FIGS. 1 a and 1 b show a fourth preferred embodiment in a representation analogous to FIGS. 9 a and 9c;

12a-12c show a fifth preferred embodiment in illustrations analogous to FIGS. 9a-9c,

13a and 13b is an illustration of a compressor according to the prior

Technique illustrating the advantages of a compressor according to the invention;

14a and 14b different control characteristics of the preferred embodiments of the compressor according to the invention. The preferred embodiments of a compressor according to the invention include (not shown in the drawings) a housing, a cylinder block and a cylinder head. In the cylinder block pistons are mounted axially movable back and forth. The drive of the compressor via a pulley by means of a drive shaft. 1

In particular, from Fig. 2 it can be seen that the swashplate mechanism of the first preferred imple mentation form the drive shaft 1, a swivel ring 2, a sliding sleeve 3, on the drive shaft 1 axially against the action of an elastic element in the form of a ring or helical return spring 4 is mounted, and a support member 5 and a power transmission element 6, which together form a driver comprises. The support element 5 is articulated both radially and (in a direction perpendicular to the drive shaft axis) perpendicular to the power transmission element 6, which means that the support element 5 in a plane (and not only along an axis) is slidably mounted. The support member 5 is cylin the bolt - shaped and has a groove 7, by means of which the support member 5 with the

Power transmission element 6 is in operative engagement. For this purpose, the support element 5 facing the end or is the support member 5 facing end portion of the power transmission element 6 in the form of a flat steel. This means that the said end region of the force transmission element 6 has an approximately rectangular peripheral contour. This approximately rectangular shaped end portion is engaged with the groove 7 of the support member 5 in engagement. The advantage of the construction of the power transmission member 6 and the support member 5 and, in particular, the strength and rigidity (small deformation) is provided by the width of the bearing. In a central region, the strength of the force transmission element 6 increases while it is sleeve-shaped at its end facing the drive shaft 1. With the help of the sleeve-shaped part 8 of the force transmission element 6 selbiges is mounted or fixed to the drive shaft 1. It should be noted at this point that in the present preferred imple mentation form, the power transmission element 6 is integrally formed and also einstoffig with the sleeve-shaped part 8. Alternatively, it could, of course, the power transmission element 6 and the sleeve-shaped part 8 by two different components (possibly even of different materials) act. It should also 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.

Since the inner diameter of the spring 4 is greater than the outer diameter of the sleeve-shaped part 8 of the force transmission element 6, the sleeve-shaped part 8 can be placed under the spring 4 in the assembled state of the swashplate mechanism. be pushed. This means that the sleeve-shaped part 8 is placed over the drive shaft 1 and fixed radially on the drive shaft 1 by the spring 4. On the side facing away from the spring 4 of the power transmission element 6, the sliding sleeve 3, which has a recess 10 corresponding to the force transmission element 6, is then slipped over the drive shaft 1. The sliding sleeve 3 also has two recesses in the form of holes 11. Axially, the force transmission element 6 and the sliding sleeve 3 are secured by a groove nut 12 (see Fig. 1) on the drive shaft 1. For a better starting behavior of the compressor, a plate spring 13 is further arranged on the drive shaft 1, which ensures that the compressor does not start at a minimum deflection angle of the pivot ring 2. This acts almost as a return spring on a hub, where the compressor can start. Furthermore, 2 stops in the form of stop plates 14, 15 are arranged on the drive shaft, which limit the deflection angle of the pivot ring. The stop disc 14 serves as a stop for a minimum deflection angle, while the stop plate 15 serves as a stop for a maximum deflection angle of the pivot ring 2.

The support element 5 is mounted in a cylindrical recess in the form of a bore 16 in the pivot ring 2. The bore 16 extends perpendicular to the drive shaft axis. The backup of the support member 5 in the pivot ring 2 by means of two snap rings 16 a.

It should be noted at this point that the power transmission element 6, which in the present preferred embodiment is rotatably connected to the drive shaft 1, in other embodiments, may also be rotatable with the same in operative engagement. Furthermore, it should be noted at this point that the drive shaft 1 is not broken through the sleeve-shaped formation or the sleeve-shaped part 8 of the force transmission element 6 and thus has corresponding stability. The clear width of the bore of the pivot ring 2 is at least slightly larger than the corresponding extent of the force transmission element. 6

In the present preferred embodiment, the mechanism of support element 5 and force transmission element 6 is not intended to transmit the torque from the shaft to the swash plate in the form of the swivel ring 2. The bearings between the support member 5 and the power transmission element 6, between the power transmission element 6 and drive shaft 1 and between the support member 5 and pivot ring 2 are not designed to transmit torque. It therefore eliminates a kind of driving function for the support member 5 and the power transmission element 6. That is off Because of the hysteresis deliberately chosen so that the tilting of the swivel ring 2 and the torque transmission are functionally decoupled from each other. The mechanism of power transmission element 6 and support member 5 essentially receives the piston forces. The torque in turn is transmitted from the drive shaft 1 to the swivel ring 2 by means of a tilting joint (realized by drive bolt 17) provided on the drive shaft central axis. The torque between the sliding sleeve 3 and the pivot ring 2 transmitting drive pin 17 are locked or secured to the pivot ring with snap rings 18. The swivel ring 2 has flattenings 19, which correspond to flattenings 20 on the sliding sleeve 3. In principle, it is also conceivable in other embodiments for the sliding sleeve 3 to be omitted and for the torque transmission to take place in any desired form between the drive shaft and the swivel ring 2 (eg via flattenings on the drive shaft 1 and the swivel ring 2).

By decoupling the torque transmission and the gas power support can be achieved that in addition to the ability to dimension the support member 5 and the power transmission element 6 correspondingly small, an optimized surface pressure, in particular between the power transmission element 6 and support member 5 and between support member 5 and pivot ring 2 can be achieved. As a result, and by the invention-specific design of support element 5 and power transmission element 6 or by the invention-specific articulation between the power transmission element 6 and support member 5, a compact design of the compressor can be achieved.

In FIGS. 1 a and b, the first preferred embodiment of the compressor according to the invention is shown in the assembled state for an angle of minimum deflection (FIG. 1 a) and an angle of maximum deflection (FIG. 1 b). In Fig. Ia with V, the position of the joint formed by the support member 5 and the power transmission element 6 is indicated, while U represents the position of the piston. Due to the configuration of support element 5 and force transmission element 6 according to the invention, U is always less than V, i. U / V is always less than 1.

In general, given all swivel ring engines already mentioned, it is possible to choose the ratio U / V less than 1. Particularly advantageous is an arrangement of the support element according to the condition V> U for the swivel ring drive with variable position of the support element (additional degree of freedom). For small tilt angles, the offset then brings about an advantageous control characteristic (cf., for this purpose, FIGS. 5 and 6). In the range of large tilt angles, the offset decreases. In Fig. 3, the dead space characteristic for the first preferred embodiment of the compressor according to the invention is shown. In the range of small deflection angle of the dead space, influenced by the deflection angle of the swash plate is minimized. In the range of large tilt angles, the dead space, influenced by the deflection angle of the swash plate, is also minimized, while in the range of average deflection angle, the clearance caused by the deflection angle of the swash plate is maximum, but the resulting 0.08 mm represents an acceptable value. Furthermore, the parameters with which the calculations of the dead space were carried out are indicated in FIG. For an explanation of the parameters, reference is made to FIG. 4 (a representation of the kinematics of the first preferred embodiment), according to which the mentioned parameters are represented as follows:

The kinematics of the compressor takes into account the position of the sliding blocks of the pistons through the center at C and the position of the support element 5 at B. The distance between C and B is a snapshot, which depends on the deflection angle. For large deflection angles results (see Fig. 4a), a position of the pivot ring 2, in which the center B of the support member 5 in the pivot ring further within the circular cylinder b lies on which the piston center axes lie, as at a mean deflection angle or a small deflection angle. More specifically, the terms in Fig. 4 mean the following:

A joint (center) of the swivel ring 2 on the drive shaft guide; B joint (center) of the support element 5 in the pivot ring 2; C center of the piston joint (sliding block) for the piston, which is in the upper

Dead center is located;

D intersection of the axis of the support member 5 with the drive shaft axis; E intersection of the drive shaft center axis with its perpendicular connection to the center C; F intersection of the drive shaft center axis with its vertical connection to the center B;

G intersection of the axis b with the axis e; a drive shaft center axis; b Central axis of the piston and the cylinder for the piston (cylinder), which is in the top dead center (usually five, six or seven

Piston used); c center line of the swivel ring 2, on which (preferably) the centers B and C are located; d support element or power transmission element center axis; e perpendicular connection from the drive shaft center axis to the center B; f perpendicular connection from the drive shaft center axis to the center C; α tilt angle of the support or force transmission element (constant, structurally selected) ß tilt angle of the swivel ring 2.

If the swivel ring 2 is pivoted more with respect to a starting position, the center of the joint B moves toward the drive shaft center axis a; This shortens the distance B-D. The degree of freedom necessary for the variation of the route length results from the articulation of the support element 5 to the force transmission element 6. The distance B-D is the hypotenuse of the triangle BDF. Catheters D-F and F-B of the right triangle also shorten. For the dead space, which is normally to be minimized, the distance D-F is of great importance.

Significant is still the (right-angled) triangle BCG. If the center of the joint B moves toward the wave axis A, then the triangle BCG increases (opposite effect as in the case of the triangle BDF, which decreases in size). In addition to the route D-F and the route C-G for the dead space of importance. Concerning the two above-mentioned distances occur when pivoting the pivot ring 2 two opposing effects, which can be used well for mutual compensation, if the parameters are selected accordingly. It is advantageous, the support member 5 and the power transmission dement 6 at an angle α (which is not equal to 0 °) to order at all to provide an effect (triangle BDF), which counteracts the effect due to the triangle BCG. The effect of BCG can only be influenced, but never avoided (as with BDF), because the bearings C and B move relative to each other depending on the tilt angle. The resulting dead space for each construction is proportional to the course of the "CG + DF" curve.

FIGS. 5 and 6 show the control characteristic of the first preferred embodiment of the compressor according to the invention for various parameters. The above-mentioned figures contain the calculation results representing the influence of the offset (U less V compared to a compressor in which U is greater than V). The rule process is comparatively more stable. The effect can be increased by increasing the offset, but then there are other disadvantages such as eg a further enlargement of the dead space. The calculation was based on the parameters already known from FIG. 3. With the given sizes, the tilt characteristics of the engine can be completely calculated. It should be mentioned at this point that simplifications have been made in the calculation, such as neglecting the friction or a frequently occurring hysteresis behavior. However, the calculation is primarily concerned only with the qualitative representation of the effect, while the aforementioned parameters only have a small influence on the quantitative behavior.

In Fig. 7, the moments are shown, which influence the tilting behavior with respect to the mass forces occurring in a compressor of the type according to the invention. The mass forces due to the piston and sliding blocks (in Fig. 7 referred to as M _k ) cause a Aufregeln the swash plate, ie a control behavior from a small to a larger deflection out. On the other hand, the mass inertia of the swivel disc has the effect that the mass forces of the swash plate have a decreasing effect from a larger deflection angle to a smaller deflection angle (the moments due to the rotationally moved masses are designated M _sw in FIG. 7). These influences occur when the speed is changed at a certain operating point and a corresponding engine room pressure. By the moments, a change in the swash plate inclination or the deflection angle of the swash plate can be effected (with appropriate design).

In modern compressors, for example for use with a refrigerant R744, an attempt is made to interpret the moments as shown in FIG. Changes in the speed then cause due to the almost constant sum of the moments no (significant) change in the deflection angle. In Fig. 8 is based on the Fig. 7, the tilting characteristic of a swash plate compressor of the second preferred embodiment shown. The tilt characteristic is usually represented as pressure in the engine room p _c via the deflection angle of the swash plate. As in FIG. 7, only a qualitative representation is given in FIG. 8, which is valid for all preferred embodiments of a compressor according to the invention.

By way of example, FIG. 8 shows two operating points which can be characterized essentially by the pressure p _s on the suction side of the compressor and a pressure p _d on the high-pressure side thereof. For each operating point are different

Speeds shown. As a result of the moment compensation, which is explained with reference to FIG. 7, the points converge at maximum deflection angle of the swash plate (cf. See also Fig. 7, which has the complete torque compensation for about a maximum deflection angle). Within one operating point, the control range is still influenced by a spring system, which defines a more or less steep slope of the curves of a family of curves in the diagram. In the case of a comparatively strong return spring, the control range is comparatively greater or the slope of the curves greater than in the case of a comparatively weak return spring. The control range in this example is about 5 bar within one operating point. If there were no compensation of the moments, the curves for the speeds from 600 to about 9000 rpm would clearly fan out and the control range would be larger. Furthermore, in this case there would be curve areas with unstable behavior (maxima, plateaus or the like). It should be noted at this point that the assumption has been made for the calculation that p _d + p _s are constant for all speeds. In particular, p _s (the pressure applied to the piston) would, however, lower due to pressure losses at higher speeds.

The representation of another operating point also given in FIG. 8 makes it clear that the present suction pressure p _s and the present high pressure p _{d have} a considerable influence on the entire control range. The operating point lying closer to the X axis in the diagram or the group of curves closer to the X axis in the diagram represents an operating state with a low suction pressure p _s and a lower high pressure p _d (70 bar / 30 bar). The operating point further away from the X axis or the family of curves further away from it stands for an operating state with a high suction and a high pressure (150 bar / 30 bar). Any other conceivable operating states are in between. Due to the described independence of the speed with respect to the tilt angle of the control range can be limited.

For a further reduction of the control range, it is advantageous if the support element gets an offset, in such a way that it is within the pitch circle diameter of the Kolbenteilkreisdurchmessers or the centers of the sliding blocks. For clarification, reference is made in this connection to FIGS. 9a to 12c. The influence of such an offset is again shown in FIG. 14 in the context of a diagram of FIG

Control characteristic shown. At a suction pressure of 30 bar, a difference at high pressures p _d of 70 bar and 150 bar is essentially no longer present. Further conceivable high pressures p _d are essentially between the stated pressures of 70 bar and 150 bar.

A variation of the suction pressure p _s to 50 bar shows that the dependence of the control range of the high pressure p _{d is} not substantially avoided as at 30 bar, but only limited. It can also be seen from FIG. 14 that the offset can not become arbitrarily large since discontinuities in the control characteristic occur or are comparatively intensified in the region of smaller deflection angles. The influence can be partially compensated by using a spring in addition to a return spring. It should be noted that the calculations of FIG. 14 are performed without the consideration of such a spring.

In addition to the advantageousness of the offset, measures are necessary to ensure that the dead space within the cylinder is substantially constant. If the head of the support element or the support element itself is displaced in the direction of the drive shaft center axis on a compressor according to the prior art, a variation of the dead space of 7/10 mm would result for an offset of 2 mm. This is too much for most technical implementations of a compressor according to the invention.

In order to avoid such disadvantages, reference is made to FIGS. 10a to 13c, in which proposals for further preferred embodiments of a compressor according to the invention and the associated dead space characteristics are shown. In FIGS. 9a and 9b, respectively, the swivel disk mechanism of a second preferred embodiment is shown, in which the recess in the swivel ring 2 extends in the axial direction. FIG. 9 a shows a representation at a minimal deflection angle, while FIG. 9 b shows a representation at a maximum deflection angle of the swash plate or swivel ring 2. In Fig. 9c, the associated dead space characteristic, i. So the dead space in millimeters above the tilt angle or the deflection angle of the pivot ring, shown.

In the illustrated in FIGS. 10a and 10b in turn at minimum and maximum deflection angle of the swivel ring 2 third preferred embodiment of a compressor according to the invention, the bore or recess 16 in the swivel ring 2, in which the support member 5 is mounted, at an angle with respect to the power transmission element 6 and the support member 5 attached. The angle is about 9 °. An offset of 2 mm causes a dead space of less than 2/10 mm. This is a value that does not pose a (major energetic) problem when operating a compressor.

In Fig. Ha a further preferred embodiment is shown, in which the influence of an offset (in the present case of 1 mm) is shown perpendicular to the previously explained offset. The influence is in a range of 1/100 mm and is therefore in the Operation fully acceptable. Such a solution with different wall thicknesses is advantageous because on the side of the swash plate on which higher forces are effective, a thicker wall thickness can be provided than on the less loaded side. This is all the more advantageous as the weight is not increased by the different wall thicknesses.

In Fig. 12, a combination of the advantages of the embodiments of Figs. 10 and 11 is realized in which the two offsets are combined. As can be seen from Fig. 12c, also in this preferred embodiment, the dead space characteristic is highly satisfactory. It should be noted at this point that the constructions according to FIGS. 9 to 12 can be seen as examples and a further optimization of the dead space characteristic is conceivable. It can be shown on the basis of the examples explained above that the variation of the dead space can be largely compensated. As a result, an offset can be provided which improves the control behavior.

In addition, the hole or opening in the pivot ring 2 can be configured so that deformation and high voltages in the range of low residual wall thicknesses of the same can be avoided.

Furthermore, the problem resulting from FIG. 13 can be solved, namely that the head of the supporting element can protrude out of the swiveling disk in the case of different operating states and thus the forces can only be transmitted in peripheral areas. By the proposed inventive offset this behavior can be prevented. Problems occur here especially when the support element the

Point support in which takes place in about the valve opening. As already mentioned in connection with Fig. 15, this point is temporally and spatially (in the direction of rotation) before the top dead center of the piston.

Although the invention will be described with reference to embodiments with fixed feature combinations, but it also includes the conceivable further advantageous combinations of these features, as they are given in particular, but not exhaustive, by the dependent claims. All features disclosed in the application documents are claimed to be essential to the invention insofar as they are novel individually or in combination with respect to the prior art. reference numeral

1 drive shaft

2 swivel ring

3 sliding sleeve

4 spring

5 support element

6 power transmission element

7 groove

8 sleeve-shaped part of the power transmission element. 6

9 groove

10 recess in the sliding sleeve 3rd

11 bore

12 groove nut

13 plate spring

14, 15 stop disc

16 hole

17 drive bolts

18 snap ring

19, 20 flattening

21 snap ring

Claims

"Axial piston" claims

1. Axial piston compressor, in particular compressors for motor vehicle air conditioning systems with at least one piston and with an inclination to a drive shaft (1) adjustable, rotationally driven by the drive shaft (1), in particular annular swash plate (2) with at least one with the drive shaft (1) pivotally arranged support member (5) is hingedly connected, wherein the piston or each having a hinge assembly or on which the swash plate (2) is in sliding engagement, characterized in that the ratio U / V between the distance U of the drive shaft center axis of the pitch circle radius of the piston center axis (n) and the distance V of the drive shaft center axis from the center of the tilting joint, which results from the articulation of the support member (5) to the swash plate (2), regardless of the size of the deflection angle of the swash plate (2 ) is always less than 1.

2. A compressor according to claim 1, characterized in that the ratio U / V with increasingly larger deflection angle of the swash plate

(2) accordingly decreases steadily.

3. Compressor according to claim 1 or 2, characterized in that the ratio U / V for small deflection angles of about 0 ° to 2 ° is about 0.93 to 0.97 and for large deflection angles of about 15 ° to 20 ° is about 0.7 to 0.75.

4. Compressor according to one of the preceding claims, characterized in that the ratio U / V * between the distance U of the drive shaft center axis of the pitch radius of the piston center axis (s) and the distance V * of the radially outer end of the support element (5) from the drive shaft Central axis for all deflection angle of the swash plate (2) is less than 1.

5. A compressor according to any one of the preceding claims, characterized in that the drive shaft central axis and the central axis of the support element (5) form an angle which is smaller than 90 °, wherein the angle is in particular 80 ° to 85 °.

6. Compressor according to one of the preceding claims, characterized in that the alignment of the recess (16) in the swash plate (2), in which a radially outer portion, in particular head of the support member (5) is mounted, from the radial direction with respect to the Drive shaft center axis deviates.

7. A compressor according to any one of the preceding claims, characterized in that the central axis of the support element (5) and the central axis of a recess (16) in the swash plate (2), in which a radially outer portion, in particular cylinder, barrel, ball -, or drop-like head of the support member (5) is mounted, form an angle which is <180 °, in particular in the range of 165 ° to 175 °.

8. A compressor according to any one of claims 1 to 6, characterized in that a recess (16) in the swash plate (2), in which a radially outer portion, in particular head of the support element (5) is articulated, in the radial

Direction is arranged relative to the drive shaft central axis.

9. A compressor according to any one of claims 6 to 8, characterized in that the recess (16) in the swash plate (2) is formed as a bore, which has at its radially inner end an area by her another, in particular cylindrical recess is superimposed, which in particular in the radial direction, ie extends at an angle of about 90 ° to the drive shaft center axis.

10. A compressor according to any one of claims 6 to 9, characterized in that by the recess in the swash plate (2) defined on both sides of the swash plate (2) extending in the axial direction wall portions have a different thickness in the axial direction such that the material concentration of the swash plate (2) is greater on the higher load side than on the lower load side.

11. A compressor, in particular according to one of the preceding claims, characterized in that the center of the tilting joint is offset axially relative to the center plane of the swash plate (2), in particular for the side facing away from the piston.