Scroll type machine

(54) MACHINE, SUCH AS COMPRESSOR, FROM THE SPIRAL DISPLACEMENT TYPE (57) A scroll-type machine is described, particularly in an art form as a compressor compressor, wherein the non-orbiting scroll member 36 is axially movable and in particular can be prestressed by pressure against the orbiting scroll member 34 is to strengthen the seal. The spiral jackets 35, 37 and their opposite sealing surfaces 104, 117 are profiled to increase performance, and for directional supply of the suction medium, a guide wall 200 is provided. To prevent rotation of the circular spiral member 34, a cross-claw clutch is provided with a non-circular ring 38, wherein an enlarged thrust bearing surface and a smaller machine size is made possible.

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AT 401 090

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AT 401 090 B

The invention relates to a machine, such as compressor, the Spiralverdrängertyp with two spiral components, each having a face plate with a sealing surface and a spiral shell arranged on this, the center axis is generally perpendicular to the associated sealing surface, and of which one spiral member relative to the other is supported in a circularly movable manner on a stationary bearing body, and the other spiral component is supported axially resiliently movable, wherein the spiral components with their spiral coats intermesh to form during operation, when the spiral member moves in a circle moving between the spiral coats moving medium chambers, each facing away from the associated sealing surface edge (front side) of the spiral shell of one of the spiral components in sealing engagement with the Sealing surface of the other spiral component is.

Such machines are used in particular for compressing gaseous media, although they may also be designed, for example, as expander, pump, etc. However, for purposes of illustration, the embodiments described below will be concerned with a hermetic refrigerant compressor.

In general, a machine of the type in question has two spiral members of similar shape, which engage with one another, with one spiral member rotated 180 "from the other. In operation, one spiral member (the orbiting scroll) orbits relative to the other (the "fixed" or non-orbiting scroll) to make moving line contacts between the flanks of the shrouds, forming moving, closed, moon-shaped media chambers. The spirals are normally formed as involutes of a circle, and ideally there is no relative rotation between the spiral members during operation, ie the movement is a purely curvilinear displacement (ie no rotation of any line in the body). The medium chambers guide the medium to be compressed or displaced from a first zone in the machine where a medium inlet is provided to a second zone where a medium outlet is provided. The volume of a sealed chamber changes as it moves from the first zone to the second zone. At any given time, there are at least two closed chambers, and if there are several pairs of closed chambers at a given time, each pair will have different volumes. In a compressor, the second zone is at a higher pressure than the first zone, and it is centrally located in the machine while the first zone is located on the outer circumference of the machine.

Two types of contacts define the media chambers formed between the scroll members: axially extending tangential line contacts between the spiral side surfaces or flanks of the spiral jackets caused by radial forces (flank sealing) and axial forces between the planar edges (the tips or flats) End faces) of each shell and the opposite face plate effected surface contacts (tip seal). For high efficiency, a good seal must be achieved for both types of contacts, the invention being primarily concerned with the tip seal.

The construction of such a machine has therefore been known for a long time and certain advantages are attributed to it. For example, such machines have high isentropic and volumetric efficiency, and are relatively small and lightweight at a particular capacity. They are quieter and vibration-free than many compressors because there are no large reciprocating parts (such as Piston connecting rods, etc.) are in use, and since the entire medium flow takes place in one direction, with simultaneous compression in several opposing chambers, there are fewer vibrations due to the pressure. Also, such machines are reliable and stable because relatively few moving parts are used, the speed of movement between the scroll members is relatively low, and there is relative insensitivity to contaminants of the medium.

One of the difficulties in designing such a machine is the technique of achieving a tip seal under all operating conditions, and also at speeds in a variable speed machine. Conventionally, this has been attempted by firstly applying very precise and very expensive machining techniques; secondly, by providing the edges or end faces with spiral seals, which unfortunately are difficult to install and often unreliable, or thirdly, an axial return force is applied by axially biasing the orbiting scroll member toward the non-orbiting scroll member using compressed working fluid. The latter technique has some advantages, but presents some problems: thus, in addition to providing a restoring force to balance the axial separation force, it is also necessary to compensate for the tilting motion on the scroll member due to the radial forces caused by the pressure, as well the inertial loads resulting from the orbiting motion, both being speed dependent. Accordingly, the axial balance force must be relatively high, and it will be optimal only at a single speed.

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In order to generate an axial preload between the two spiral components of a spiral displacement machine, it is known from US Pat. No. 3,874,827 to press a spiral component against the associated other spiral component in the axial direction by means of an end-face spring washer (or an end-face elastic seal ring). As a result, however, the above-mentioned problems arise. A comparable construction is described in DE 2 428 228 A, wherein there are provided for pressing the two spiral components against each other in the axial direction on the one hand, for example, a spring ring and on the other hand, a pressure chamber contained in a hub extension of a spiral member. Thus, even with this known machine, the described problems can occur as a result of undesired tilting movements, sealing problems, etc.

DE-1 935 621 A shows a machine construction comparable to the constructions according to the first two cited documents, namely with a disk spring acting on an end face of the spiral components or a pressure equalizing chamber.

Finally, a displacement machine from DE-2 831 179 A1 is known, but the arrangement is only offset to a lateral elasticity of the one spiral component in order to allow a lateral evasive movement in order to compensate for inaccurate parallel guides, inaccurate contour production, etc., to a certain extent ,

The object of the invention is to provide a spiral displacement machine with a spiral component attachment, by which the problems occurring in the known machines with regard to sealing, etc. are eliminated.

This object is achieved in that an axially resilient support which is supported in a fixed position with respect to the bearing body is connected to the other spiral member to its axially movable mounting at a location generally in the middle between the planes of the two sealing surfaces. By this measure, inter alia, that undesirable tilting or tumbling movements of the axially movable Spiraibauteils, eg due to mechanical resonances, suppressed or prevented. In this case, an axial Vorsoannung the other non-rotating Soiralbauteils is advantageously made possible by pressure, with inertial load problems do not occur, and the required extent of such bias by pressure is limited to a minimum extent, which is just sufficient to the axial separation forces to compensate, whereby the required return forces are reduced considerably and advantageously. Although a compression bias of the non-orbiting scroll member has already been proposed as mentioned above (see US-A-3,874,827), it should be noted that the known machine has the same drawback with respect to the tilting movements as that in which the orbiting scroll member is biased , In addition, the inventive design provides improved control of the non-axial movements of the non-orbiting scroll member.

A structurally advantageous embodiment of the invention provides that the other spiral member is held by the holder against rotational movement and radial movement relative to the axis of the circular motion of the one spiral member. By this design, no additional means for preventing the rotational and radial movement of the other spiral member are required.

Furthermore, it is advantageous if the holder is connected to a plurality of, in a common plane in the middle between the two sealing surfaces points spaced from each other with the other spiral member, so that a uniform force transmission between the bearing body and the axially movable scroll member on the Holder is enabled.

It is also advantageous if the holder contains at least one leaf spring which is stretchable within its elastic limits in normal axial deviations of the axially movable other spiral member. By this elastic connecting member is achieved that the axially movable scroll member is returned to its original position even at low movements.

In order to prevent deviations or movements in the circumferential direction or in the radial direction in the case of the axially movable spiral component, it is furthermore advantageous if the holder contains counter-bearing surfaces in sliding engagement with the bearing body or spiral component.

A particularly advantageous and simple design of these abutting abutment surfaces is characterized in that one of the abutment surfaces is formed by a pin and the other by a pin slidably receiving bore: it is further favorable when the pin is adjustably mounted; Preferably also have pin and bore circular cross-section.

So that the occurring during start-up of the machine axial separation forces between the two spiral components can not be too large, is an advantageous development of the machine the invention is characterized by a stop for limiting the axial movement of the other spiral member away from a circularly movable scroll member to a predetermined maximum stroke. Advantageously, this stop is provided such that the maximum stroke is set sufficiently small, so that the

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Machine in start-up mode, at maximum displacement, can work. By this measure it is achieved that the compression function of the machine is guaranteed even at startup.

According to a preferred embodiment of the invention it is provided that the holder contains at least one in plan view generally U-shaped spring whose web portion is held in position relative to the bearing body in position and whose legs are connected near their ends with the other spiral member. Due to the U-shaped design of the spring, a mechanically favorable attachment of the same is made possible, on the one hand, with the bearing body and, on the other hand, with the axially movable spiral component.

On the other hand, it has also proven to be advantageous if the holder includes a spring ring which is held with its outer side relative to the bearing body and connected with its inner side with the other spiral member. The annular formation of this type of spring allows a relatively simple production. To influence the rigidity of this spring ring or To reduce it, it is further favorable here, if the spring ring has several openings in order to increase its flexibility. These openings can also be produced in a simple manner, for example stamped. Preferably, each aperture in plan view is elongated and disposed at an angle with respect to a generally radially extending axis. This makes it possible to reduce the rigidity of the spring ring particularly effective without the same occurs in an axial movement of the spiral member a noticeable rotational movement.

In view of the desired elastic properties and corrosion resistance, it is particularly advantageous if the spring consists of spring steel, wherein it preferably consists of flat spring steel.

An expedient, structurally simple manner of fastening a spring arrangement is characterized in that a spring belonging to the holder is fastened to a generally flat, transverse end face of a bearing body belonging to axially extending support column.

It is further advantageous if the end face of the supporting column lies substantially in a plane parallel to the planes of the sealing surfaces, wherein it is further preferably provided that the plane of the end face of the support column lies between the planes of the sealing surfaces. This ensures that any tilting or tumbling movements of the axially movable scroll member are reduced to a minimum.

On the other hand, it is also advantageous if the axially movable other spiral member is provided with a generally planar mounting surface on which the spring is secured with a protruding leg; Preferably, in this case, the mounting surface is approximately in the plane of the end face of the support column. Even with this arrangement, the tilting or tumbling movements can be minimized.

A particularly advantageous embodiment of the support column further consists in that the end face of the support column has a generally perpendicular to the leg of the spring edge to facilitate the bending of the spring with minimal stress.

In a particularly advantageous manner, a relatively soft seal is disposed between the end face of the support column and the spring, whereby the distribution of the clamping load is reduced to the spring. Preferably, the seal has an edge which coincides substantially with the edge of the end face of the support column, whereby also the deflection of the spring is facilitated. Also, it has proved to be useful if the seal consists of a relatively soft metal. This results in advantages in terms of mechanical properties and corrosion behavior of the seal.

In a further preferred fastening of the spring is provided that the spring is held by a stop on the end face in position, wherein the stop limits the axial movement of the other spiral member away from a circularly movable scroll member to a predetermined stroke This results in the saving a separate component as a stop.

It has been found that improvements can be made by slightly altering the configuration of the sealing surfaces and the edges of the scroll members. According to the invention, it is therefore preferably provided that the sealing surfaces are slightly concave, wherein optionally a part of each of the two sealing surfaces between the opposite flanks of the spiral jacket is formed axially stepped to define a slightly concave surface.

It can also be provided with advantage that the edges of the spiral sheaths are slightly concave.

By such profiling, inter alia, the advantage is achieved that a thermal increase near the center of the machine can be compensated, and the application of relatively rapid machining operations for the production is facilitated; this results in a compressor that achieves its maximum power in a much shorter interval time than conventional machines.

Another structurally advantageous embodiment is characterized in that the holder includes a plurality of resilient brackets which are between a machine housing and the axially movable

AT 401 090 Β other spiral component are connected; in this case, preferably, each console is L-shaped, wherein one leg is fixed to the housing and the other leg on the axially movable other spiral member; Advantageously, furthermore, each console is extensible within its elastic limits in a normal axial movement of the axially movable other spiral member. This arrangement is characterized by an extremely simple structure.

A further advantageous embodiment is characterized in that the holder includes a plurality of tubular elements, each retained with a flange relative to the bearing body and connected to another flange with the axially movable other spiral member; Preferably, here the flanges are arranged in a generally horizontal transverse plane, and / or the tube elements are arranged in the circumferential direction at intervals around the axially movable other spiral member, further preferably the tube elements each contain a tube member having a central axis, which is generally tangential to the axially movable other spiral member, and preferably the tube elements with their axes in pairs of 0 'or 180 * include different angles. In these embodiments, a very stable structure is achieved by the use of the tubular elements.

In the case of the embodiment of the holder using a leaf spring, it is further advantageous according to the invention, when the leaf spring is held centrally relative to the bearing body and fixed with its ends to the axially movable other spiral member, or if the leaf spring is mounted centrally on the axially movable other spiral member and with their ends is held relative to the bearing body. Preferably, the spring is elongate and in plan view, relatively straight, or the spring is elongate and arcuate in plan view. This arrangement is characterized by a relatively simple structure, with slight axial movements of the spiral member only cause stretching of the leaf springs within their elastic limits.

Of particular advantage, especially with regard to the control of movements during operation, it is then when the holder contains a plurality of balls, which are each arranged in two opposing, axially extending grooves, wherein a groove is fixed relative to the bearing body and the other groove is fixed relative to the axially movable other spiral member; In this case, preferably, a groove is arranged in a ring which surrounds the axially movable other spiral member, and which is biased to pre-load the balls in the grooves. Because of the low rolling resistance of the balls a possible axial movement of the spiral member is opposed to a low resistance, so that the return to the starting position under leadership can be done quickly and with low energy consumption.

This particular advantage is also achieved in a similar manner when the support comprises a plurality of rollers each disposed in two opposed, axially disposed grooves, one of which is fixed relative to the bearing body and the other is fixed relative to the axially movable other spiral component. further preferably, the one groove is disposed in a ring surrounding the other scroll member, the ring being biased to load the rollers in the grooves.

It has also been found to be favorable with regard to the prevention of tilting and tumbling movements, when the holder contains at least two axially extending guide surfaces that are fixed relative to the bearing body, and when with respect to the axially movable other scroll member fixed counter bearing surfaces with these guide surfaces in engagement stand, wherein a biasing means presses the abutment surfaces in abutment with the guide surfaces; Preferably, the guide surfaces are flat and turned radially inwards; Further, the guide surfaces may be disposed at locations at a distance corresponding to a central angle of 90 'and, preferably, the biasing means exerts a force in the direction of a line bisecting the angle between the guide surfaces. This embodiment is simple in construction and robust.

In another embodiment with a holder using leaf springs is preferably provided that a plurality of leaf springs are circumferentially spaced around the axially movable other spiral member around, or that the⁷ Holder includes a pair of leaf springs, which are arranged on opposite sides of the axially movable other spiral member. Due to the restoring force of the leaf springs, axial movements can be brought to an end more quickly.

To tilt or To prevent tumbling movements of the axially movable scroll member relative to the circular scroll member particularly effective, is a particularly advantageous embodiment of the machine according to the invention, in which the axially movable scroll member is biased by fluid pressure against the circular scroll member, characterized by a mounted in a fixed position relative to the bearing body cylinder chamber . in which a piston connected to the axially movable other scroll member is slidably received in a direction substantially parallel to the axes, and the pressurized fluid is deliverable. Due to the arrangement of a piston on the axially movable scroll member is on the one hand an axial guide and on the other hand to a circle5

AT 401 090 B, the spiral component directed toward force or bias achieved. Advantageously, this machine further extends generally transversely through the side walls of the piston or the cylinder chamber to guide pumped medium at outlet pressure from the machine, and an elastomeric ring seal between the piston and the cylinder chamber is provided on axially opposite sides of the channel. This arrangement causes a fluidically favorable discharge of the compressed medium into the outlet chamber. At the same time it is prevented by the seal that medium between the piston and Zyiinderwand can flow back.

A favorable development is further characterized by a further, with respect to the bearing body in a fixed position mounted cylinder chamber in which a further connected to the axially movable other spiral member piston is slidably disposed in a direction substantially parallel to the axis, and the pressurized Medium is fed. The arrangement of two cylinder chambers, the possibility is given to apply their pistons either with different pressures. Preferably, a cylinder chamber with outlet pressure and the other cylinder chamber with a pressure between the suction pressure and the outlet pressure act bar. In particular, it is provided that the first-mentioned cylinder chamber is that cylinder chamber which can be acted upon with outlet pressure. Alternatively, however, both cylinder chambers can be acted upon with a pressure between discharge pressure and intake pressure. Thus, here is the opportunity to achieve optimal axial balance for different modes.

An advantageous embodiment of the machine in which the two spiral members are biased against each other by fluid pressure is further characterized by two pressurized fluid chambers whose pressures together bias the two spiral members toward each other generally parallel to the axis of circular motion of the one spiral member to strengthen the sealing effect. Preferably, here one of the fluid chambers can be acted upon with outlet pressure; and / or one of the fluid chambers with a pressure between the suction pressure and the outlet pressure can be acted upon.

In a further embodiment of this machine is preferably provided that at least one of the fluid chambers, preferably both fluid chambers, (each) with respect to the bearing body fixed cylinder chamber in which a connected to one of the spiral components piston is arranged generally axially displaceable; In this case, the two pistons can be connected to the other, axially movable spiral member.

On the other hand, preferably an axially directed surface of one of the Spiraibauteile beaufschiagbar with the two pressures; In this case, each fluid chamber is preferably defined in part by an exposed surface of the spiral member; Preferably, an elastomeric ring seal is arranged between the fluid chambers. Also in this machine, it is possible to apply the pistons in an advantageous manner with different pressures.

For this purpose, according to the invention can also be provided that in one of the spiral members, a channel for guiding fluid from one of the media chambers is provided at a pressure between suction pressure and outlet pressure in one of the cylinder chambers, wherein preferably further comprises a channel for guiding fluid at outlet pressure in the other cylinder chamber, preferably in the same spiral member as the former channel, is provided. The arrangement and cross-section of the channels can be chosen to achieve the desired pressures.

It is particularly advantageous if the cylinder chambers and the pistons are generally concentric with each other and the cylinder chambers are formed by a stepped cylinder wall with two different inner diameters, wherein the further piston is formed by an annular shoulder on the first-mentioned piston, which enclosed by the smaller diameter part of the cylinder wall is, whereas the other piston is surrounded by the larger diameter part of the cylinder wall. This design of the cylinder and piston is characterized by easy manufacturability and high reliability.

In the machine of the type according to the invention, with an engine, a crankshaft driven by this engine about a substantially vertical axis and a lubricant source, it is particularly advantageous if in the circular scroll member a circular cylindrical axial bore is mounted, in which a Mitnehmerbüchse is mounted in turn has another cylindrical axial bore in which a crank pin of the crankshaft is received, whereby the scroll member is set in a circular motion with rotating crankshaft, and when in the Kurbeiwelle an oil supply passage is provided, the lubricating oil from the oil source to the top of the crank pin leads, from where the lubricating oil is pressed at circulating Kurbeiwelle by the centrifugal force outwards, wherein in the Top of the drive bushing is formed a recess for collecting lubricating oil to supply this to the axial bores for lubrication. By these measures, in connection with the inventive central, resilient support for the other spiral component favorable drive training is achieved, in particular to avoid the unwanted tilting moments, which further adequately a supply of

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Lubricating oil is ensured for the drive connection between the crankshaft and the rotating spiral component.

In order to dissipate excess oil effectively, it is also provided with advantage that the Mitnehmerbüchse outside has a flat surface which defines a gap between it and the first-mentioned axial bore for the oil flow, which is in communication with the recess; while the flat surface may further extend axially from the bottom to the top of the driving bush; Further, preferably, the further axial bore has a non-circular cross section, whereby an oil flow clearance is defined between the driving bush and the crank pin, which is in communication with the recess; the further axial bore is preferably generally oval in shape and the crankpin is generally circular. For easy coupling of the crank pin with the drive bushing, it is also advantageous if the further axial bore and the crank pin each have a flat surface which are in drive connection with each other.

A preferred embodiment with respect to the said recess for collecting lubricating oil is characterized in that the recess is a groove in the upper surface of the Mitnehmerbüchse extending between the further axial bore and the outer surface of the sleeve. It is also advantageous if the angular position of the recess lags slightly relative to that of the oil feed channel in the direction of rotation of the crankshaft, whereby a reliable discharge of excess oil can take place.

For supplying oil to the lubricating system, it is further advantageous if in the lower part of the crankshaft, an oil pump is arranged, which is mounted in an oil sump forming the oil source and supplies lubricating oil from this Öisumpf to Ölzuführkanal with rotating crankshaft. The operation of the oil pump can be based on the centrifugal force generated during the rotation in a simple manner.

In order to reliably prevent the circular spiral component from rotating relative to the bearing body and the axially movable spiral member, it is also particularly advantageous if the bearing body has a part which is generally circular around the machine axis and a securing member is secured against rotation relative to the bearing body to secure the circular spiral member Cross-claw clutch (Oldham coupling) is provided wherein on the bearing body generally diametrically oriented first bearing surfaces and the circular spiral member generally formed second bearing surfaces, at right angles to the first bearing surfaces are defined and a transversely arranged ring substantially surrounds the circular bearing body part, wherein the inner peripheral surface of the ring deviates from the circular shape Has shape and the ring on opposite sides comprises circular arcs of the same radius, the centers of curvature of which are separated by a predetermined distance, and wherein substantially straight parts connect the arcs, which ring further comprises on one side a first pair of wedges in linear sliding engagement with the former bearing surfaces and on the opposite side a second pair of trowels in linear sliding engagement with the first having second-mentioned bearing surfaces. The use of this specially trained Oldham coupling allows ideal storage of the circular spiral member relative to the bearing body, without causing the spiral member rotates about its own axis. Another advantage is also to be seen in the fact that the Oldham ring allows the use of a larger thrust bearing or an outer housing with reduced diameter at a pressure bearing of predetermined size.

It is of particular advantage in this embodiment further when the radius is equal to the radius of the circular bearing body part plus a predetermined minimum clearance, and further when the circular bearing body part defines a flat transverse thrust bearing surface which slidably supports the orbiting scroll member. Preferably, the predetermined distance is in a direction generally parallel to the diameter on which the former bearing surfaces are aligned, and / or the predetermined distance is equal to twice the orbital radius of the orbiting scroll member. Also, it is beneficial if the former bearing surfaces are formed by a pair of radial slots in the bearing body, which on diametrically opposite sides, with respect to the axis, are arranged, and / or when the second bearing surfaces are formed by a pair of radial slots in the circular scroll member, which are provided with respect to the axis diametrically opposite sides.

A further preferred embodiment of the machine according to the invention is characterized in that for formation as a hermetic compressor, a sealed housing is provided with a medium inlet opening in a wall, wherein at a distance from this inlet opening a compressor medium inlet is provided that attached to the housing in overlap with the inlet opening a baffle is that defines an opening below the inlet opening, which serves as a drain for entrained in the inlet medium oil, which separates on impact with the baffle, and that extending from the baffle up an axially extending passage defining member extends to receive inlet medium supplied at the opposite end of the component in the compressor inlet is directed. At this

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Training, particularly suitable for a refrigerant compressor, prevents mixing of suction medium with oil dispersed within the compressor housing, with the baffle functioning as an oil separator to remove any oil already carried, and preventing the transfer of engine heat to the intake medium, thereby preventing the transfer of oil Overall efficiency is significantly improved. The passage can be limited in part by the plastic part and partly by the housing; Preferably, the lower opening is defined between the baffle and the housing, and it is further favorable if a flexible tab is formed on the component, which is pressed against an abutment surface within the housing in order to press the component in position.

The invention will be further elucidated below with reference to embodiments shown in the drawing, in which: Fig. 1 is a vertical section through a scroll type compressor, with parts broken away, showing a section generally taken along the line 1-1 in Fig. 3; where certain parts are shown slightly twisted; Figure 2 is a similar sectional view taken along the line 2-2 in Figure 3, wherein also some parts are shown slightly rotated; Fig. 3 is a plan view of the compressor of Figs. 1 and 2 with a portion of the top removed; Fig. 4 is a plan view similar to Fig. 3, but with the entire upper assembly of the compressor removed; Figures 5, 6 and 7 are partial plan views similar to the right-hand portion of Figure 4 with successive parts removed to more clearly show the details of the construction; 8 shows a partial section along the line 8-8 in Figure 4; 9 shows a partial section along the line 9-9 in Figure 4; Fig. 10 is a sectional view taken along the line 10-10 in Fig.1; 11A and 11B in development spiral-shaped vertical sectional views substantially according to the lines 11A-11A and 11B-11B in Figure 10, wherein the profile is graphically shortened and greatly exaggerated. FIG. 12 is a sectional view, taken generally along line 1212 in FIG. 10; FIG. Fig. 13 is a plan view of an Oldham ring in a changed new form; Fig. 14 is a side view of the Oldham ring of Fig. 13; Figure 15 is a partial sectional view taken substantially along the line 15-15 in Figure 10, illustrating some lubrication channels. Fig. 16 is a sectional view taken substantially along the line 16-16 in Fig. 15; Figure 17 is a horizontal section substantially along the line 1717 in Figure 2; Figure 18 on an enlarged scale a vertical partial section to illustrate another embodiment; 19 shows a view similar to FIG. 18 of a further embodiment; Fig. 20 is a somewhat diagrammatic horizontal sectional view illustrating another technique for mounting the non-orbiting scroll member for limited axial compliance; Fig. 21 is a sectional view taken substantially along the line 21-21 in Fig. 20; Fig. 22 is a sectional view similar to Fig. 20, but illustrating another technique for axially compliant attachment of the noncirculating scroll member; Fig. 23 is a view similar to Fig. 20, but illustrating another technique for mounting the non-orbiting scroll member; Fig. 24 is a sectional view taken substantially along the line 24-24 in Fig. 23; Fig. 25 is a view similar to Fig. 20 of another technique for mounting the non-orbiting scroll member; Fig. 26 is a sectional view taken substantially along the line 26-26 in Fig. 25; Fig. 27 is a view similar to Fig. 20 of yet another technique for mounting the noncirculating scroll member; Fig. 28 is a sectional view taken substantially along line 28-28 in Fig. 27; Fig. 29 is a view similar to Fig. 20 of another technique for mounting the non-orbiting scroll member; Fig. 30 is a sectional view taken substantially along line 30-30 in Fig. 29; Figs. 31 and 32 are views similar to Fig. 20, illustrating two additional similar techniques for a limited axial compliant mounting of the non-orbiting scroll member; and Fig. 33 is a view similar to Fig. 20, schematically illustrating yet another technique for limiting axially compliant attachment of the noncirculating scroll member.

Although the invention can be basically applied to a variety of types of scroll type machines, it will be explained below with reference to, for example, a hermetic compressor, specifically to a compressor used for compressing refrigerant for air conditioners and refrigeration systems.

1-3, the machine comprises three main units, namely a central assembly 10 within a circular cylindrical steel housing 12, a top assembly 14 and a bottom assembly 16 welded to the bottom and top of the housing 12, respectively, to terminate and seal. Housing 12 houses the major components of the machine, including generally an electric motor 18 having a stator 20 (with conventional windings 22 and guard 23) press-fitted within housing 12, a rotor 24 (with conventional tabs 26). mounted on a crankshaft 28 in a heat shrink fit, a compressor body 30, preferably connected to the housing 12 at a plurality of circumferentially spaced locations; as welded at 32, for example, and carries a circular scroll member 34 having a spiral jacket 35, with a desired standard flank profile and a tip or end face 33, an upper crankshaft bearing 39 in a conventional two-piece bearing construction, a non-circular, axially compliant

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Spiral member 36 having a spiral shell 37 with a desired standard edge profile (preferably equal to that of the spiral shell 35) which meshes with the spiral shell 35 in the usual manner, and further has a tip or end face 31, wherein in the spiral member 36 an outlet opening 41st is provided and between the scroll member 34 and the compressor body 30, an Oldham ring 38 is mounted to prevent rotation of the scroll member 34; Further, a suction inlet fitting 40 is soldered or welded to the housing 12 and a straightening suction assembly 42 is provided to direct the suction gas to the compressor inlet; a lower bearing support bracket 44 welded to the housing 12 at each end, such as at 46, carries a lower crankshaft bearing 48 in which the lower end of the crankshaft 28 is journaled. The lower end of the compressor forms a sump which is filled with lubricating oil 49.

The bottom assembly 16 includes a simple steel compact 50 having a plurality of feet 52 and perforated mounting flanges 54. The pressing piece 50 is welded to the housing 12, such as at 56, to cover and seal the lower end thereof.

The shell assembly 14 is an outlet muffler having a lower pressed steel closure member 58 which is welded to the top of the housing 12, such as at 60, to cover and seal more tightly. The closure member 58 has an upstanding peripheral flange 62 from which a perforated retaining tab 64 (Figure 3) projects, and in its central region defines an axially disposed circular cylindrical chamber 66 having a plurality of apertures 68 in the wall. To increase its rigidity, the closure member 58 is provided with a plurality of protruding or ribbed portions 70. An annular gas outlet chamber 72 is defined above the closure member 58 by an annular muffler member 74, which at its outer periphery with the flange 62, approximately at 76, and at its inner periphery with the outer wall of the cylindrical Kamme? 66, about 78, is welded. Compressed gas coming from the outlet port 41 passes through the openings 68 into the chamber 72, from where it is normally discharged via an outlet fitting 80 soldered or brazed into the wall of the muffler 74. A conventional internal pressure safety valve assembly 82 may be mounted in a suitable opening in the closure member 58 to vent outlet gas into the interior of the housing 12 at positive pressure.

The crankshaft 28 rotatably driven by the motor 18 has at its lower end a smaller-diameter bearing portion 84, with which it is mounted in the bearing 48, being supported by the thrust washer 85 with the shoulder bounding the bearing portion 84 (see FIGS and 17). At the lower end of the bearing 48, an oil inlet channel 86 and a mud discharge channel 88 are provided. The support bracket 44 is formed in the shape shown and provided with upstanding side flanges 90 to increase their strength and rigidity. The bearing 48 is lubricated by immersion in the oil 49 and oil is pumped to the remainder of the compressor by a conventional crankshaft centrifugal pump having a central oil passage 92 and an associated, outwardly inclined oil supply passage extending to the top of the crankshaft 94 has. A transverse channel 96 extends from the oil supply channel 94 to a circumferential groove 98 in bearing 39 to lubricate it. A lower counterweight 97 and an upper counterweight 100 are secured to the crankshaft 28 in any suitable manner, such as by fitting onto protrusions in the lugs 26 in a conventional manner (not shown). These counterweights are of conventional design.

The orbiting scroll member 34 includes an end plate 102 having generally planar, parallel upper and lower surfaces 104 and 106, the lower surface 106 slidingly resting on a planar circular thrust bearing surface 108 on the body 30. The thrust bearing surface 108 is lubricated via an annular groove 110 which receives oil from the channel 94 in the crankshaft 28 via the channel 96 and the groove 98, the latter communicating with a further groove 112 in the bearing 39, which oil the intersecting channels 114 and 116 in the body 30 (see also Fig. 15). The tips 31 of the spiral mantle 37 are in tight engagement with the surface 104, and the tips 33 of the spiral mantle 35 are in tight engagement with a generally planar and parallel surface 117 on the spiral member 36.

From the scroll member 34 is in a piece down from a hub 118 from which has an axial bore 120 in which a circular cylindrical relief drive sleeve 122 is mounted in the axial bore 124, an eccentric crank pin 126 is arranged as a drive connection, said crank pin 126 in a Piece is formed at the upper end of the crankshaft 28. The drive is radially yielding, wherein the crank pin 126 drives the sleeve 122 via a flat surface 128 on the pin 126, which is in sliding engagement with a flat bearing insert 130 which is disposed in the wall of the bore 124 (see also Fig. 16). , Rotation of the crankshaft 28 causes the sleeve 126 to rotate about the crankshaft axis, which in turn causes the scroll member 34 to move in a circular orbit. The angle of the planar drive surface is chosen such that the drive introduces a slight centrifugal force component on the orbiting scroll to the flank seal

AT 401 090 Β reinforce. The bore 124 is cylindrical, but is also somewhat oval in cross-section to allow limited relative sliding between the journal and bearing, which in turn allows for automatic separation and therefore relief of intermeshing spiral flanks when liquids or solids are drawn into the compressor.

The described radially compliant circular motion drive is lubricated using an improved oil delivery system. Oil is pumped through channel 92 to the top of channel 94 from where it is forced radially outward by the centrifugal force, as indicated by dashed line 125 in FIG. The oil is collected in a recess in the form of a radial groove 131 which is disposed in the top of the sleeve 122 along the path 125. From here it flows downwardly into the clearance between pin 126 and bore 124 and between bore 120 and a flat surface 133 on sleeve 122 which is aligned with groove 131 (Figure 16). Excess oil then flows via a channel 135 in the body 30 to the oil sump 49.

Rotation of the scroll member 34 relative to the body 30 and the scroll member 36 is connected by an Oldham coupler having a ring 38 (see FIG. 13 and 14) having two downwardly projecting integral diametrically opposed wedges 134 slidably received in diametrally opposed radial slots 136 in the body 30, and further offset by 90 ° therefrom two upwardly projecting, diametrically opposed, one-piece ones Has keys 138 which are slidably disposed in diametrically opposed radial slots 140 in the scroll member 34 (one of which is shown in FIG. 1 shown).

The ring 38 has a particular shape that allows the use of a maximum size thrust bearing for a given overall engine size (as viewed in cross section) or a minimum engine size for a given thrust bearing size. This is due to the fact that Oldham's ring 38 moves in a straight line relative to the compressor body and therefore is configured with a generally oval or raceway-like configuration with minimum internal dimensions to expose the peripheral edge of the thrust bearing. The inner peripheral wall of the ring 38, with a shape for control, has a side 142 with a radius R, starting from a center x, and an opposite side 144 having the same radius R, starting from a center y (FIG. 13). the intermediate wall portions being substantially straight, as illustrated at 146 and 148 in FIG. The center points x and y are provided at a distance from each other equal to twice the orbital radius of the spiral member 34, and they lie on a line passing through the centers of the keys 134 and radial slots 136; the radius R is equal to the radius of the thrust bearing surface 108 plus a predetermined minimum clearance. Except for the shape of the ring 38, the Oldham coupling operates in a conventional manner.

There will now be described the particular suspension system over which the upper non-orbiting scroll member is axially movably mounted while at the same time being retained against any radial movement and rotational movement to allow axial thrust biasing to the face or tip seal with this suspension. This suspension is best seen in FIGS. 4 to 7, 9 and 12. FIG. 4 shows the top of the compressor with removed top assembly 14, and Figs. 5 to 7 show similar partial top views, with each additional parts are removed. On each side of the compressor body 30 there is provided a pair of axially projecting support columnar holders 150 having flat top surfaces lying in a common transverse plane. The scroll member 36 has a peripheral flange 152 with a transverse planar top having a recess 154 for receiving the holders 150 (Figures 6 and 7). The retainers 150 have axially extending threaded holes 156, and the flange 152 has corresponding holes 158 equidistant from the holes 156.

On the top of the holders 150 is disposed a flat soft metal gasket 160 having the shape shown in Fig. 6, and on top of the gasket 160 is a flat spring spring 162 made of spring steel and having the shape shown in Fig. 5; above that, a headband 164 is provided, and all the aforesaid parts are held together by threaded bolts 166 which are threaded into the holes 156. The outer ends of the spring 162 are secured to the flange 152 by threaded bolts 168 disposed in the holes 158. The opposite side of the spiral member 36 is held in an identical manner. As can thus be seen, the spiral member 36 can move slightly in the axial direction by bending and stretching (within the elastic limits) of the springs 162, but it can not twist or move in the radial direction.

The maximum axial movement of the spiral members 34, 36 in the separation direction is limited by a mechanical stop, namely the abutment of the flange 152 (see part 170 in Figures 6, 7 and 12) against the under surface of the spring 162, which is a retaining clip 164 is supported, and in the opposite direction by investing the spiral jacket Simseiten on the end plate of the respective opposite

AT 401 090 Β

Spiral component. These mechanical stops cause the compressor in operation to a compression even in the rare situation in which the axial separation force is greater than the axial return force, as in the case of tarnishing. The maximum end-face clearance allowed by the stopper may be relatively small, on the order of about 0.1 mm for a spiral with a diameter of 7.5 to 10 cm and a shell height of 2.5 to 5 cm.

Prior to final assembly, the scroll member 36 is aligned relative to the body 30 by means of a retainer (not shown) having pins which are insertable into positioning holes 172 on the body 30 and positioning holes 174 on the flange 152. The support column-like holders 150 and the seal 160 are provided with substantially aligned edges 176 which are generally perpendicular to the extending portion of the spring 162 to reduce the stresses. The seal 160 also aids in distributing the clamp load to the spring 162. As shown, the spring 162 is in its unloaded condition when the scroll member 36 is in its maximum peak clearance state (ie on headband 164) to facilitate manufacture with assembly. However, since the stress in the spring 162 is so low over the full range of axial motion, the initially unloaded axial position of the spring 162 should not be critical.

It is essential, however, that the transverse plane, in which the spring 162 is arranged, as well as the surfaces of the body 30 and the non-orbiting scroll member 36 to which it is attached, are arranged substantially in an imaginary transverse plane through which Center of the interlocking spiral coats runs, ie approximately in the middle between the surfaces 104 and 117th This allows the support provided for the axial compliant scroll member 36 to be minimized by the compressed medium acting in the radial direction, ie, the tilting moment caused by the pressure of the compressed gas acting radially against the flanks of the spiral jackets , If this tilting moment is not compensated accordingly, this could lead to a lifting of the spiral member 36 from the seat. This technique of balancing this force is superior to the use of conventional thrust bias because it reduces the possibility of over-biasing the scroll member toward one another and also makes the face sealing bias substantially independent of compressor speed. Due to the fact that the axial separation force does not act exactly in the center of the crankshaft, a small tilting movement may remain, but this is comparatively insignificant compared to the separation and return forces normally encountered. There is therefore a particular advantage in the axial preload of the non-orbiting scroll member 36 as compared to that of the orbiting scroll member, in the latter case there being a need to compensate for tilting forces due to radial separation forces and those due to inertial forces, which is a function of speed This can lead to excessive compensation forces, especially at low speeds.

The described axially compliant attachment of the scroll member 36 allows the use of a very simple pressure biasing arrangement to reinforce the tip seal. This is accomplished by using pumped medium at outlet pressure, or an intermediate pressure, or pressure reflecting a combination of both. In its simpler and presently preferred form, axial bias is achieved in a tip sealing or merging direction using the outlet pressure. As best seen in FIGS. 1 1-3, the top of the scroll member 36 is provided with a cylindrical wall 178 which surrounds the outlet port 41 and defines a piston slidably disposed in a cylinder chamber 66 with an elastomeric seal 180 provided to reinforce the seal , In this way, the scroll member 36 is biased in the return direction by compressed medium at discharge pressure acting on the area at the top of the scroll member 36 defined by the piston 178 (reduced about the area of ​​the outlet opening 41).

Since the axial separation force is a function of the discharge pressure of the engine (among others), it is possible to choose a piston area which results in an excellent tip seal under most operating conditions. Preferably, the area is chosen such that there is no substantial separation of the scroll member 34, 36 at any time in the cycle during normal operating conditions. Furthermore, in a situation of maximum pressure (maximum separation force), there would at best be minimal net axial compensation force and, of course, no substantial separation.

With respect to the tip seal, it has been found that substantial improvements can be achieved with a minimum start-up or inter-period time by slightly altering the configuration of the end plate surfaces 104 and 117 and the spiral shell end surfaces 31 and 33. Preferably, each of the end plate surfaces 104 and 117 is slightly concave, and when the spiral casing end faces 31 and 33 are similarly configured (ie the surface 31 is generally parallel to the surface 117 and the surface 33 is generally parallel to the surface 104), the result stagnates contrary to what is expected

AT 401090 Β would, since this leads to a certain initial axial clearance between the spiral components in the middle region of the machine, which is the area of ​​highest pressure. However, it has been found that, because the central area is also the hottest, there is a greater thermal increase in the axial direction in this area, which would otherwise lead to excessive, greatly reducing the efficiency of rubbing in the central area of ​​the compressor. By providing this initial extra clearance, the compressor reaches a state of maximum tip sealing when it reaches operating temperature.

Although a theoretically smooth concave surface may be better, it has been found that the surface can be formed to have a stepped helical shape which is much easier to machine or machine. As best seen in the greatly exaggerated illustration in FIGS. 11A and 11B, which also applies to FIG. 10 The surface 104, although generally planar, is actually formed by spiral stepped surfaces 182, 184, 186, and 188. The end face 33 is similarly formed with helical steps 190, 192, 194 and 196. The individual steps should be as small as possible, the total displacement from the plane being a function of the spiral jacket height and the thermal expansion coefficient of the material used. For example, in a three turn (spiral turns) cast iron scroll machine, the ratio of overhang or flight to total axial displacement may be in the range of 3000: 1 to 9000: 1, with a ratio of about 6000: 1 being preferred. Preferably, both spiral members 34, 36 have the same face plate and tip surface shapes, although it should be possible to provide the entire axial surface displacement only with a scroll member, if desired. It is not critical where the steps are arranged because they are so small (they can not be detected with the naked eye), and because they are so small, up to the surfaces in question can be considered as generally plane. This stepped surface is very different from designs according to previous, unpublished proposals in which relatively large steps (with a step seal between the mated scroll members) are provided to increase the pressure ratio of the machine.

In operation, a cold machine has a tip or face seal in the outer peripheral area at start-up, but an axial play in the center area. When the machine reaches operating temperature, the axial thermal expansion of the central spiral turns reduces the axial play until a good face seal is achieved, which is increased by the compression bias (as described above). If there is no such initial axial surface displacement, the thermal expansion in the center of the machine causes the outer spiral turns to axially separate, accompanied by a loss of the good face seal.

The present compressor is also provided with means for directing the intake gas entering the housing directly to the inlet of the compressor itself. This advantageously facilitates the separation of oil from the inlet suction medium and prevents the suction medium from picking up dispersed oil within the housing. It also prevents the Ansaugges from unnecessarily picking up heat from the engine, which would reduce the volume efficiency.

The straightening suction assembly 42 includes a lower baffle 200 formed of a sheet of metal and provided with circumferentially spaced vertical flanges 202 welded to the inner surface of the housing 12 (see Figures 1, 4, 8) and 10). The baffle 200 is disposed directly opposite the inlet of the suction fitting 40 and provided with an open bottom portion 204 so that oil entrained in the incoming suction gas will impact the baffle 200 and then drain into the compressor sump 49. The assembly further includes a plastics molding 206 having a downwardly extending, integral, arcuate channel portion 208 which extends into the space between the top of the baffle 200 and the wall of the housing 12, as best seen in FIG , The upper portion of the plastic molding 206 is generally tubular (diverging radially inwardly) to supply gas flowing up the channel portion 208 radially inwardly into the peripheral inlet of the intermeshing scroll members 34, 36. The member 208 is retained in the circumferential direction by a notch 210 in which one of the fastening bolts 168 is braced, and in the axial direction by an integrally formed tab 212 which is pressed against the underside of the closure member 58, as best seen in FIG is. The tab 212 resiliently biases the member 206 axially downwardly to the position shown. The radially outer extent of the directional intake port is defined by the inner wall surface of the housing 12.

Energy is supplied to the compressor motor in a conventional manner using a conventional terminal block which is protected by a suitable lid 214 (see also Figures 3 and 4).

AT 401 090 Β

There are several ways to achieve the compressive bias in the axial direction to strengthen the tip seal, as illustrated in Figs. 18 and 19, wherein the parts shown there have similar functions as those parts in the above-described first embodiment are provided, wherein the same reference numerals are used for the ensprechenden parts.

In the modified embodiment of Fig. 18, the axial bias is achieved through the use of compressed medium at an intermediate pressure lower than the outlet pressure. This is accomplished by providing a piston 300 at the top of the scroll member 36 which slides in a cylinder chamber 66 but has a closure member 302 which prevents the top of the piston from being exposed to the discharge pressure. Instead, outlet medium flows from the discharge port 41 into a radial passage 304 in the piston 300, followed by an annular groove 306 which is in direct communication with the ports 68 and the discharge chamber 72. Elastomeric seals 308 and 310 provide the required seal. Pressurized compressed media is tapped off the desired sealed pocket defined by the spiral turns via a channel 312 to the top of the plank 300 where it exerts an axial return force on the noncircular scroll member 36 to reinforce the face seal.

In the embodiment of Fig. 19, a combination of outlet and intermediate pressures is used for the axial face seal bias. To accomplish this, the closure member 58 is formed to form two separate coaxial spaced cylinder chambers 314 and 316, and the top of the scroll member 36 is provided with coaxial pistons 318 and 320, each in one of the chambers 314 and 316 are arranged. Exhausted pressurized medium is supplied to the top of the piston 316 in much the same manner as in the first embodiment, and intermediate pressure medium is supplied to the annular piston 318 via a passage 322 extending from an appropriately located pressure tap. If desired, the piston 320 may also be exposed to a second intermediate pressure rather than the outlet pressure. Since the areas of the pistons and the locations of the pressure tap can be changed, this embodiment provides the best way to achieve optimum axial balance for all desired operating conditions.

The pressure taps may be selected to provide the desired pressure and, if desired, may be located to experience different pressures at various locations in the cycle so that a desired average pressure can be obtained. The pressure channels 312, 322 and the like. Are preferably relatively small in diameter, so that the flow (and thus the pumping loss) is minimal and pressure changes and thus force changes) are damped.

In Figs. 20-33, several applied suspension or support systems, ie brackets, are illustrated which serve to secure the non-orbiting scroll member 36 allowing limited axial movement while retaining this scroll member 36 against radial and circumferential movement. Each of these embodiments is such as to secure the non-orbiting scroll member 36 at its center, similar to the first embodiment, so as to compensate for tilting moments on the scroll member caused by radial medium compressive forces. In all embodiments, the upper surface of the flange 152 is in the same geometric position as in the first embodiment.

20 and 21, a support is obtained by means of a spring steel ring 400, which is anchored on its outer periphery by bolts 402 to a mounting ring 404 which is fixed to the inner wall of the housing 12; on its inner circumference, the spring steel ring 400 is anchored to the upper surface of the flange 152 on the non-orbiting scroll member 36 by means of fastening bolts 406. The ring 400 is provided with a plurality of angled elongate openings 408 disposed throughout its full extent to reduce rigidity and allow limited axial excursions of the noncirculating scroll member 36. Since the openings 408 are inclined with respect to the radial direction, a displacement of the inner circumference of the ring relative to the outer circumference does not require tensioning of the ring, but causes extremely slight twisting. However, this extremely limited rotational movement is so insignificant that it should cause no noticeable loss in efficiency.

In the embodiment of FIG. 22, the non-orbiting scroll member 36 is most simply mounted by means of a plurality of L-shaped brackets 410 welded to one leg on the inner wall of the housing 12 and the other leg to the upper surface of the flange 152 are fastened by means of a suitable fastening bolt 412. The brackets 410 are configured so that they can stretch slightly within their elastic limits to accommodate axial deviations of the non-orbiting scroll member 36.

AT 401 090 Β

In the embodiments of Figs. 23 and 24, the fastening means comprises a plurality (three shown) of tubular parts 414 provided with a radially inner flange 416 on the upper side of the flange 152 on the non-orbiting scroll member 36 by means of suitable bolts 418 and a radially outer flange 420 are connected by suitable bolts 422 with a bracket 424 which is welded to the inner wall of the housing 12. Radial deflecting movements of the non-orbiting scroll member 36 are prevented by the fact that a plurality of tubular parts are used, at least two of which are not directly opposed to each other.

In the embodiment of FIGS. 25 and 26, the non-orbiting scroll member 36 is limitedly axially movably supported by leaf springs 426 and 428 having their outer ends secured to a mounting ring 430 welded to the inner wall of the housing 12 by suitable fasteners or bolts 432 and attached to the top of the flange 152 at its center by suitable bolts 434. The leaf springs may be either straight, as in the case of the spring of the 426, or arcuate, as is the case with the spring 428. Slight axial deflecting movements of the scroll member 36 cause the leaf springs to stretch within their elastic limits.

In the embodiment according to FIG. 27 and 28, a movement of the non-orbiting scroll member 36 is prevented radially and circumferentially by a plurality of balls 436 (one shown), these balls sealingly seated in a cylindrical bore defined by a cylindrical surface 437 on the inner peripheral edge of a mounting ring 440, which is welded to the inner wall of the housing 12, and defined by a cylindrical surface 439, formed in the radially outer peripheral edge of a flange 442 on the non-orbiting scroll member 36, the balls 436 lying in a plane midway between the face plate surfaces of the scroll members, for the reasons already explained above.

The embodiment of Figures 29 and 30 is apparently identical to that of Figures 27 and 28, but differs therefrom in that, instead of the balls, a plurality of cylindrical rollers 444 are used, one of which is shown, and those in a tight fit are pressed within a rectangular slot which is defined by a surface 446 on the ring 440 and a surface 448 on the flange 442. Preferably, the mounting ring 440 is sufficiently resilient so that it can be stretched over the balls or rollers to bias the assembly and eliminate any undesirable backlash or backlash.

In the embodiment of FIG. 31, the non-orbiting scroll member 36 is provided with a centrally located flange 450 having an axially extending hole 452 extending therethrough. Within the hole 452, a pin 454 is slidably mounted, which is fastened at its lower end close to the body 30. As shown in FIG. 31 can be seen, axial deviations of the non-circular spiral member are possible, whereas deviations in the circumferential direction or in the radial direction are prevented. The embodiment of FIG. 32 is similar to that of FIG. 31 except that the pin 454 is adjustable. This is achieved by providing an enlarged hole 456 in a suitable flange on the body 30 and forming the pin 454 with a support flange 458 and a threaded end portion extending through the hole 456 and threading a threaded nut 460 Has. Once the pin 454 has been accurately positioned, the nut 460 is tightened to permanently anchor the parts in position.

In the embodiment of Figure 33, the inner wall of the housing 12 is provided with two lugs 462 and 464 having precisely machined radially inwardly facing flat surfaces 466 and 468 which are disposed at a right angle relative to one another. The flange 152 on the non-orbiting scroll member 36 is provided with two mating lugs, each having a radially outwardly directed flat surface 470 or 472 have, wherein these surfaces 470, 472 relative to each other at a right angle and abut the surfaces 466, 468. These lugs and surfaces are precisely machined to position the noncirculating scroll member 36 in the proper radial and angular position. To keep it in this position, while allowing limited axial movement, is an extremely stiff spring in the form of a Belleville washer or the like. 474 provided between a projection 476 on the inner surface of the housing 12 and a projection 478 which is attached to the outer periphery of the flange 152. The spring 484 exerts a strong biasing force on the non-orbiting scroll member to hold it in position against the surfaces 466 and 468. This force should be slightly greater than the maximum radial and rotational force that normally occurs to move the scroll member out of its seating position. The spring 474 is preferably positioned such that the biasing force it exerts has equal components toward each of the lugs 462 and 464 (ie, its diametral line of force represents an angle bisector with respect to the two lugs). As in the previous embodiments, the lugs and the point of application of the spring are provided substantially midway between the spiral member and the plate surfaces to compensate for tilting moments.

AT 401 090 B

In all embodiments according to FIGS. 20 to 33, an axial movement of the non-rotating

Spiralbauteiles 36 separating direction by any suitable means, such as a mechanical stop, as described in the first embodiment, are limited. Of course, movement in the opposite direction is limited by engagement of the spiral members.