FR2482209A1 - Rotary gas-compressor assembly - has pair of cooperating lobed rotors of dissimilar asymmetric configurations and incorporating balancing plugs - Google Patents

Abstract

The compressor has a pair of co-operating rotors (20,22) with parallel rotation axes and dissimilar asymmetric lobed configurations. The rotor (22) has a cross-sectional profile that is greater in area than that of the other rotor and comprises a sector (11) of radius (R), which is somewhat greater than the radius (r) of the outer edge of a discharge port (44). Balancing plugs (100) may be lodged in the rotors. Strips of, e.g. a lead-tinalloy may be attached to the rotor (22) and the inner peripheral wall of the casing to facilitate 'lapping-in' of the rotors. Sealing strips may of the same material may also be attached to the casing end-walls.

Description

 The present invention relates to the field of rotary compressors. More particularly the invention relates to rotary compressors in which two rotors are mounted in adjacent holes of a body, each rotor comprising a single projecting lobe or tooth which engages in a corresponding recess provided in the other rotor when the rotors rotate in opposite directions.

The invention relates in particular to rotary compressors of the type which can operate without the introduction of lubricant and / or coolant into the compressor chamber, as is the case otherwise for compressors supplied with oil or water.

 General type double rotor rotary compressors which are the subject of the present invention have been known for many years. Generally, compressors include two rotors rotating in opposite directions and mounted in cylindrical bores (ie cylinders) of a body or casing. Each rotor has a projecting tooth or lobe (and some designs have more than one) which engages in a corresponding recess in the other rotor. The end parts of the rotor are intended to apply sealingly against the interior walls of the cylinders, and parts of the hubs of the rotor are intended to apply sealingly one uSe against the other so as to rotate the relative to each other, so that a gas can be compressed internally when the rotors rotate. There are provided gas discharge inlet ports which are sealed from each other during the rotation cycle so that internal compression can occur.

It has been proposed to provide a machine of this type with two identical, or essentially identical, rotors which rotate in chambers or cylinders of equal dimensions and profiles, that is to say that the profiles of each of the rotors and of its tooth or teeth. protruding are identical. We also know how to design rotors of different dimensions or profiles and in this case the rotors can be mounted in different holes. We have also proposed to mount rotors of different profiles in cylinders having holes of identical size but some arrangements of this type has been affected by the major drawback that continuous fluid communication between the chambers is greatly reduced or even eliminated and that most or all of the internal compression must occur only in one of the chambers, which decreases the efficiency of the machine.

Examples of known embodiments include the following patents: UK 1,304,394, UK 900,881, UK 752,437, us. 3,535,060, US 3,472,445, US. 2,097,037,
UK 661 749, UK 341 324, US 4076 469, US 4033708, US. 3,941,521, US. Re. 29,627, US. 3,723,031 and
US 2,058,817.

 A persistent problem with the design and construction of these rotary compressors is to maximize the flow through the machine and maximize the discharge capacity while avoiding high internal discharge speed which causes large losses, especially when using axial inlet and discharge ports. Another problem concerns the dynamic balancing of each rotor, especially when using rotors of different profiles. In these machines, there are also still difficulties concerning the seal, in particular the seal between the rotors, the seal between the end parts of the rotors and the cylindrical walls or chambers, and the seal between the ends of the rotors and the end covers or plates which close the cylinders or bores.

The present invention aims to remedy the aforementioned drawbacks of known embodiments and to provide an improved rotary compressor of the double rotor type.

 Taking into consideration that all designs of rotary compressors are limited with respect to the speed and capacity that a model of given dimensions is capable of achieving efficiently, a main feature of the present invention is to increase the maximum the exposed area of the discharge port (in a configuration where the shape of the rotor is used to synchronize the opening of the discharge port). This maximum increase in the exposed area of the discharge port is achieved by a compressor design in which the discharge rotor (i.e. the rotor which covers and uncovers the axial discharge port) has larger dimensions and a profile different from the inlet rotor (i.e. the rotor which covers and uncovers the axial and / or radial inlet ports) .The profile of the discharge rotor has a section having a minimum radial dimension ( R) which determines the outer limit of the discharge port and which is made as large as possible while taking into account that (1) excessively high internal flow velocities and discharge velocities must be avoided and (2) it is necessary to maintain sufficient dimensions for the inlet or suction rotor to allow it to be balanced effectively.

This maximum increase in the exposed area of the discharge orifice is made possible, while taking account of the two considerations mentioned above, by using rotors of different dimensions and profiles but which nevertheless rotate in cylinders of equal dimensions. . The external profile of the discharge rotor is formed by a series of segments, some of which are arcs of a circle while other segments correspond to non-circular curves or profiles. For the discharge rotor of radius R, a segment corresponding to a large circular arc is chosen to maximize the radial dimension of the axial discharge orifice, since the maximum radial dimension of the discharge orifice must be slightly less than the radial dimension R of the discharge rotor. Another circular segment is chosen for the discharge rotor having a radius approximately equal to the radius of the cylinder in which the rotor must rotate in order to be able to establish the necessary seal between the rotor and the wall of the cylinder. Other profiled segments of the discharge rotor are determined by design considerations, for example as a function of the balance of the rotor or as a result of the joining or joining of profile segments or as a function of corresponding surfaces of the input rotor . Likewise, the profile of the input rotor is determined to some extent by similar design configurations as well as sealing considerations with respect to the discharge rotor and the cylinder wall, or by function discharge rotor.

 Another main characteristic of the present invention relates to the sealing of the rotor. Predetermined patterns or patterns of a material which can undergo crushing and / or abrasion are placed on the faces of the casings at each end of the cylinder and / or on the end faces of the rotor in order to ensure the sealing of the gap of leakage existing between them. The sealing between the tips of the rotors and the inner cylindrical wall is also established by using a similar material, which can undergo crushing and / or abrasion, on the tips of the rotors or else on the walls of the cylinders in order to form both a wear surface and a labyrinth seal. It is also possible to position, on one or both rotors, a solid coating or thin strips of a material capable of undergoing crushing and / or abrasion, at the places where the rotors must come into contact with each other at sealing purposes.

Since the rotors have different dimensions and profiles, the difference between the profiles means that a relative speed is established between the rotors in their zone of minimum approximation. This relative speed causes removal of the coating material by abrasion so that optimal sealing surfaces are formed by wear in the sealing strips of the rotor profiles. This sealing establishment process cannot be applied to rotors of identical profiles since it does not establish the necessary relative speed between the rotors, at the point of minimum approach, over most of the profile.

 Consequently, the object of the invention is to provide a rotary compressor of the double rotor type.

 The invention also aims to provide an improved rotary compressor of the double rotor type in which the yields and the capacity for a given dimension are optimized by maximizing the exposed area of the discharge orifice while meeting the requirements of to avoid high losses resulting from an increase in internal speeds and to ensure a good balance of the rotors.

The invention further aims to provide an improved rotary compressor of the type b double rotor, which comprises rotors of different dimensions and profiles.
Another object of the invention is to provide an improved rotary compressor of the double rotor type, in which the sealing is improved using materials which can undergo crushing and / or abrasion and which are subject to wear in progress. operating rotors to improve sealing.

Other advantages and characteristics of the invention will be highlighted in the following description, given by way of nonlimiting examples, with reference to the appended drawings in which FIG. 1 (a) is a schematic representation of a rotary compressor group, FIGS. 1 (b) and 1 (c) are views in side elevation and in rear elevation of a casing of a rotary compressor according to the invention, FIG. 2 is a sectional view taken along line 2-2 of FIG. l (a), which shows the rotary compressor according to the invention, FIG. 3 is a detailed view of the discharge rotor of the rotary compressor of FIG. 2 fig 4 is a detail view of the intake rotor of the rotary compressor of fig 2, fig. 5 to 8 show the rotary compressor whose rotors occupy different positions during an operating cycle, FIG. 9 (a) gives details of a sealing structure provided on the face of the end casings, FIG. 9 (b) shows a variant of the sealing structure of FIG. 9 (a), and figs. ji0 and 11 show other details of the sealing structure.

In fig. l (a) and l (c), a two-stage rotary compressor group is shown. The system shown schematically in FIG. l (a) comprises a first stage rotary compressor 10 and a second stage rotary compressor 12. The rotary compressors 10 and 12 are of identical design, although they may differ in specific details regarding dimensions and profiles , or the axial length of the rotors. Consequently, only the essential elements of the rotary compressor 10 will be described hereinafter, since similar elements are involved in the rotary compressor 12, unless otherwise advised. The rotary compressor 10 comprises an external body 14 in which are placed two chambers or cylinders 16, 18, preferably identical and intersecting. An inlet or suction rotor 20 is mounted. rotatably in the cylinder 16 while a discharge rotor 22 is rotatably mounted in the cylinder 18. Extreme housings 24 and 26 provide sealing of the opposite ends of the cylinders. The rotors 20 and 22 are mounted respectively on shafts 28 and 30 which are in turn driven in rotation at equal speeds by gears 32 and 34. The gear 32 drives the gear 34 while the gear 32 is itself - even driven by a drive pinion 36 itself actuated by a motor 38.

 Air enters the compressor 10 both in an axial direction (i.e. in a direction parallel to the axes of rotation of the rotors) and in a radial direction (i.e. a direction oriented radially and perpendicularly to the axes of rotation of the rotors 20 and 22). The axial and radial air inlets have been shown diagrammatically at 40 and 42 in FIG. 2 and they are indicated in more detail in FIGS. 1 (b), 1 (c) and 2. The air which has entered the rotary compressor 10 is compressed inside it and then it is discharged , at a first stage outlet pressure and via two axial discharge orifices 44 in the end housings 24 and 26. The compressed air which leaves the first stage is channeled to a pulsation damper 46 (which basically works to dampen discharge pulses) then in an interstage cooler 48 (where the temperature of the discharged air is reduced) then in a water separator 50 (where the excess of normally existing moisture can be removed ). The compressed air leaving the first stage is then introduced, via an axial inlet 40 (a) (cf. FIG. 1 (c)) into the second stage compressor 12 where it is subjected to a second stage of internal compression to then be discharged into a second assembly comprising a pulse damper 46 ', a cooler 48' and a water separator 50 ', the air then being discharged for its final use. As indicated by the relative axial lengths of the compressors 10 and 12, the axial length of the rotors 8 of the second stage compressor 12 is less than that of the compressor 10, due to the reduced volume of compressed air which is introduced into the compressor 12.

 Considering fig. 1, it can be seen that the elements of the two-stage compressor unit have been shown only with a schematic configuration by way of illustration, no indication being given with regard to the details of configurations and dimensional tolerances of the rotors. The compressors also include rotor bearings, cooling jackets and other conventional parts, the details of which have not been shown in the figure. It also goes without saying that components such as gears, motors, pulse dampers, coolers and water separators can be conventional elements well known in this field. In reality, the elements of the compressor unit are arranged and distributed in the floors as shown in Figures 1 (b) and 1 (c).

Referring to Figures 1 (b) and 1 (c), it can be seen that the two stages of the compressor unit are placed one above the other, that is to say that the first stage compressor 10 is placed above the second stage compressor 12, unlike conventional multi-stage groups where the stages are arranged side by side. A single drive pinion 36 drives the two driving gears 32 of each compressor stage and all the gears are housed in a common casing 37 so that these gears are all placed at one end of the group and that the two compressor stages are arranged one above the other on one side of the gear housing 37. The pulsation dampers 46, '46 ', the coolers 48, 48' and the water separators 50; 50 'have been deleted from FIGS. 1 (b) and l (c) to clarify the drawing, but it goes without saying that they are incorporated into the group as indicated in FIG. 11 (a),
The arrangement of the compressor stages which has been shown in FIGS. 1 (b) and l (c) is important because it is possible to obtain a very compact structure occupying little space on the ground and making it possible to reduce the overall dimensions of the compressor unit by comparison with the case where compressors of similar capacity are placed side by side. It is possible to reduce the overall width of the machine (i.e. the dimension measured from left to right by about 25% by placing the stages l 'one on top of the other and not side by side. You can also get additional reductions with regard to the overall envelope (i.e. the envelope generally square or rectangular which usually surrounds the compressor unit), since the space in front of the gearbox can be used to position exhaust pipes or other accessory components.

Again considering Figures 1 (b) and 1 (c), it should be noted that the first stage inlet orifices 40, 42 are placed in a large stabilizing chamber 43 which can be part of the inlet silencer, and that the second stage inlet 40 (a) is part of a large stabilizing chamber 45. The positions of the various entrances in the large stabilizing chambers 43 and 45 make it possible to obtain large volumes of entry into the passageways intake to minimize areas of high speed air flow and thereby reduce pressure drop.

In fig. 2, there is shown a sectional view taken along line 2-2 of FIG. the). As shown in fig. 2, the inlet or suction rotor 20 rotates clockwise while the discharge rotor 22 rotates counterclockwise.

In the positions occupied by the rotors in FIG. 2, the inlet port and the discharge port 44 are shown in a partially uncovered condition, while the covered portions of each port are indicated by dashed lines. It is obvious that the discharge port 44 is covered and uncovered (i.e. closed and open) by action of the discharge rotor 22 during its rotation, while the suction port 40 is covered and discovered by; the combined action of the inlet rotor 20 and the discharge rotor 22. Also the communication between the radial inlet orifice 42 and the interior of the body 14 is interrupted intermittently by the suction rotor 20 when it goes past the inlet 42 to start each compression cycle.

 The rotor 20 can be considered to be formed from a hub 52 and a protruding lobe 54. Similarly, the rotor 22 can be considered to be formed from a hub 56 and a protruding lobe 58.

 The particular shapes and profiles of the rotors 20 and 22 are of great importance in the present invention. The shape and profile of the discharge rotor 22 are imposed in part by the imperative of increasing the cross-section of the discharge orifice 44 as much as possible. This maximum increase in the cross-section of the discharge orifice is obtained mainly by a design which gives the maximum radial value "r" to the external dimension of the discharge orifice. The dimension "r" is in turn determined by defining the maximum dimension "R" which can be given to the hub of the discharge rotor, while observing the requirements of balancing and elimination of excessive flow velocities. definition of the considerations involved in the design of the discharge rotor 22, it is useful to take into account the fact that the rotors 20 and 22 are designed to rotate in a synchronized relationship in order to fulfill the functions of air intake, compression and discharge. The rotors rotate at equal angular speeds and the tips of each of lobes 54 and 58 meet twice at each revolution in the junction zones of the cylindrical chambers 16 and 18.

For determining the outline or profile of the discharge rotor 22, it is advantageous to divide the rotor into a series of quadrants of 90o. Quadrant I covers an arc of 900 clockwise from the tip of lobe 58. Quadrant II covers the next arc of 90, quadrant III covers the third arc of 900 and the quadrant IV covers the fourth arc of 900. To avoid confusion in the drawing of fig. 2, only the quadrants I and II of the discharge rotor 22 have been indicated. The four quadrants in question are indicated in detail in FIG. 3, which is a detailed and enlarged view of the profile of the discharge rotor 22. Likewise, as shown in FIG. 4, the input or intake rotor 20 can also be considered to be formed by four quadrants, starting from the tip of the lobe 54, these quadrants progressing in the counterclockwise direction and being numbered in correspondence with the quadrants of the discharge rotor. The relationship between the different quadrants of the two rotors will be explained below.

 We will first describe the quadrant II of the discharge rotor 22 (with particular reference to FIG. 3), since its size has a critical influence on the design of the rotor, and also on several other dimensions. The part of the discharge rotor 22 which is included in quadrant II constitutes the area of the discharge rotor which covers the radial end of the discharge orifice 44 at the start of the compression phase of an operating cycle of the rotor ( when the ends or points of lobes 54 and 58 are located in the upper part of the body cavity 14, as shown in Fig. 5, and move away from each other). with the imperative consisting of a maximum increase in the radial dimension of the discharge orifice, the profile of the quadrant II of the rotor 22 is an arc of a circle.

The radial dimension R which defines the arc of a circle of quadrant II is selected so as to satisfy as much as possible the requirement of avoiding narrowed passages producing high pressure drops between the rotor and the cylinder as well as imperative of giving the suction rotor a shape allowing it to be balanced. Another limitation imposed on the value of the radius R is that there must remain in the cylinder holes an appropriate suction volume. The radial dimension i'r "of the discharge orifice is slightly less than the dimension R of the rotor It should be noted that, when it is contemplated in accordance with the present invention to increase the dimensions r and R to the maximum, we can take as reference values the values that said dimensions would have in an identical body compressor , where the inlet and discharge rotors would have essentially the same dimensions and profiles.

The surface of quadrant III of rotor 22 has been given, in the embodiment shown, the shape
of an ellipse but it goes without saying that we can adopt other contours. The formulas concerning the ellipse are then determined by the volume of material which it is desirable to have in this part of the rotor hub to ensure correct dynamic balancing, and the desired spacing between the largest dimension is also involved. (at point p1) of this segment of the rotor and the cylinder wall, said spacing determining the speed of flow in the compressor.

 The quadrant I of the discharge rotor 22 is formed of two parts, the first of which corresponds to an arc of a circle of 12 which starts from the point of the lobe 58 in the direction of the hands of a watch while the other part corresponds to an arc of 780 and to a shape of ellipse modified so as to obtain slopes equal to the junction between the elliptical part and the arc of circle of 12 ". The arc of circle of 12" has a radius R1 , which is essentially equal to the internal radius of the cylindrical part of the body 14. The angle of 120 corresponding to this arc is chosen so as to be equal to the arc length of the radial inlet orifice 42. we determine the relation existing between the modified ellipse and the circular segment of radius R at the junction of the quadrants I and In so that the ellipse is placed at a distance R from the center of the rotor and is perpendicular to a radius R at the junction with the curve of radius R of quadrant II, in order to obtain equal slopes and a uniform profile at the junction point of quadrants I and II.

 The quadrant IV of the discharge rotor 22 is defined, that is to say determined, by the surfaces of the intake rotor 20 and a more detailed explanation will be given thereafter after description of the determination of the shape of the rotor. admission 20.

Referring now to the inlet or inlet rotor 20 of Figs. 2 and 4, it should be noted that this inlet rotor 20 is smaller than the discharge rotor 22, which results directly from the fact that the discharge rotor 22 has been dimensionally increased to maximize the cross section of the discharge port, as previously described. Starting with quadrant II of the input rotor, which cooperates with quadrant Il of the discharge rotor, the profile of quadrant II of the input rotor is an arc of circle of radius R2 equal to the difference between the difference " S "existing between the axes of rotors 20 and 22 and the dimension
R of the rotor 22. The surfaces of the quadrants II of each rotor are in sealed contact with each other during each cycle of the compressor and these surfaces execute a relative sliding movement (without rolling component). The quadrant III of the input rotor 20 is a surface which is generated, that is to say completely determined, by the elliptical shape of the surface of the quadrant IIT of the discharge rotor 22. The surfaces of the quadrants III of each rotor are also in contact sealed with each other during each compressor cycle and these surfaces perform a relative sliding movement. (A sealing contact corresponds to an actual contact condition or else to a very small interval of the order of 0.025 to 0.05 mm which is obtained after wear-in of the seals subject to abrasion, as will be explained in in more detail below.).

 With regard to quadrant I of the input rotor 20, the first part 9, that is to say the part which extends anticlockwise from the start of the quadrant, over an arc of approximately 520, up to the point designated by P2, is arbitrarily long enough not to penetrate the corresponding arc of 520 of quadrant I of the discharge rotor. This area of the input rotor can be chosen somewhat arbitrarily since it does not have to establish leaktight contact with the corresponding part of the discharge rotor. Likewise, the remaining part corresponding to an arc of 380, of the quadrant I of the input rotor is a surface generated by the modified ellipse, which forms the remaining part corresponding to 330 of quadrant I of the discharge rotor.

E: ra now considering quadrant IV of the input rotor 20, we see that the part which starts from the tip of lobe 54 and which extends in the sendes of a watch, corresponds to an arc of 120 d ' a circle of radius R1, that is to say a radius equal to the interior radius of the cylindrical body 14 in which the rotor is placed (and equal to R1 of the quadrant I of the rotor 22). We chose an arc length of 120 corresponding to the arc length of 12- provided for the opening of the radial inlet 42.The rest of this fourth quadrant of the input rotor is defined, always in considering the clockwise direction, by two circular arcs, namely an arc whose center is located at a point P3 located on the radius R1 in position 120 and having a radius R3 extending over approximately 980 to 'to P4 on the rotor surface; the other arc has a radius R4, which is centered on the ordinate Y which defines the end of the quadrant
IV. It is obviously important that the arcs of a circle defined by the radii R1, R3 and R4 have slopes equal to the junction points of the arcs in order to maintain the continuity of the profile.

The quadrant IV of the discharge rotor 22 is a surface which is entirely determined by the shape of the quadrant IV of the inlet rotor 20. The first section of 780 of the surface profile of the quadrant IV of the discharge rotor is generated directly by the parts having arc lengths determined by the radii R3 and R4 of the surface profile of quadrant IV of the input rotor. Likewise, the last section of 12 of quadrant IV of the discharge rotor is a circular arc of a radius suitable for sliding on the circular arc of 120 of the first section of quadrant IV of the input rotor (this is ie the arc length from the tip of lobe 54 to the end of quadrant IV of the input rotor, determined by the arc of 120 of radius R1).

The entire profile of the internal tooth or lobe recess of the discharge rotor, that is to say of the hollowed part which extends from the tooth tip or lobe 58 to point P5 located at the end quadrant
IV, is a regulated surface which is determined by the path followed by the tip of the tooth 54 when it moves along this surface of the discharge rotor during rotation. The profile of the corresponding recess provided for the tooth or lobe of the intake rotor 20, and which extends from the tip of tooth 54 to point P6 located at the beginning of quadrant I, corresponds to a profile which is absolutely not critical from the point of view of correspondence Aside from the fact that this recess profile of the intake rotor must be arranged so as to avoid interference with the profile of quadrant I of the discharge rotor when it passes in front of the recess of the rotor. intake, said recess profile of the intake rotor is arbitrary and can be determined by taking into account considerations of balance, weight and resistance.

 As indicated above, the fact that excessive flow velocities and corresponding high pressure losses have to be avoided constitutes a limiting factor in order to maximize the radius R of the quadrant II of the discharge rotor, which causes a limitation. of the radius r of the discharge orifice. The radius R of quadrant II of the discharge rotor determines the beginning of the elliptical part of the profile of quadrant III, which is desirable for balancing purposes. This elliptical part is oriented perpendicular to a radius R at the end of the quadrant He and she is limited, in its protrusion outwards from the center of the rotor, by the radial interval which exists between the point P1 of the surface of the discharge rotor in quadrant III and the interior surface of the chamber 18 during the compression and discharge phases of the rotor operating cycle e the point P1 defining on the surface of the discharge rotor, in quadrant III, a point which corresponds to the greatest radial spacing with respect to the center of rotation of the rotor ).

The radial interval existing between the point P1 and the interior surface of the chamber 18 defines a neck which determines the speed of at least part of the gas in the compressed state when it flows from the space existing between the rotors towards the discharge orifice during the compression phase and / or the discharge phase.

The smaller the radial interval between point P1 and the chamber wall, the greater the gas flow speed and the higher the pressure losses. Consequently, a compromise must be chosen between the desired enlargement of the discharge port and the increase in the pressure losses which occur when the discharge port is increased, and therefore the dimensions of the discharge rotor.

Another compromise which must be taken into account is imposed by balancing conditions. When the radius R of quadrant II of the discharge rotor is increased, the radius R2 of associated quadrant II of the intake rotor must be correspondingly reduced, and it -Also reduce the base or root of lobe 54 in the vicinity of the junction point of rays R3 and R4 in quadrant ~ IV of the intake rotor. As the radius R2 and the root of the lobe 54 of the intake rotor decrease in size, it becomes increasingly difficult, and ultimately impossible, to balance the intake rotor and the mechanical strength of the intake rotor. in a transverse direction at the base or root of the lobe becomes insufficient. Consequently, the enlargement of the discharge rotor is also limited by considerations of maintaining the dimensions of the intake rotor at values large enough for correct balancing. of this rotor and also by considerations of mechanical resistance.

In fig. 5 to 8, we have shown different positions of the rotors and different modes involved in a compressor operating cycle. In fig. 5, the rotors have just closed the admission orifices of the chambers 16 and 18, between the outer peripheries of the rotors and the interior walls of the chambers. In the position shown in fig. 5, chambers 16 and 18 are filled with air or gas at an inlet pressure and the compression cycle will begin. Now considering fig. 6, the rotors continuing to rotate in the directions indicated, the volume between the front surfaces of the rotors is constantly reduced so that the volume of gas in this space is compressed. The space behind the rotors obviously expands and is filled with gas, at inlet pressure, for the next cycle. The discharge port is sealed with parts of quadrants III and IV of the rotor 22. The parts in an arc placed at the points of each of the teeth of the rotors 54 and 58 are applied in a sealed manner against the interior walls of the chambers so as to hermetically seal the compression volume existing between the front surfaces of the rotors with respect to the inlet pressure prevailing behind the rotors, a seal being also established between the corresponding surfaces existing between the rotors, in the zone where they meet, in order to hermetically insulate in the same way the reduction zone of volume and pressure increase over the inlet pressure.

 In the position shown in fig. 6, the compression cycle is almost, but not quite finished, and the discharge port 44 begins to open as the discharge rotor 22 continues to rotate counterclockwise. When the discharge opening begins to be discovered by the rear edge of the rotor in quadrant IV, the gas has not yet been fully compressed to the discharge or discharge pressure. As a result, gas momentarily flows back to the compressor from the pulsation damper as compression continues in the machine. With rotor 22 continuing to rotate counterclockwise, the gas continues to be compressed and a larger section of the discharge port is cleared. When the maximum discharge pressure is reached, gas then leaves the discharge orifice in the direction of the pulse damper. This process of premature opening of the discharge port (that is to say before compression is completed) makes it possible to obtain a larger opening section of the discharge port at the start of the discharge process, with reduction corresponding to the total losses at the landfill.

 Now considering fig. 7, it can be seen that the two rotors have arrived by rotation in the position where they are in the zone of intersection of the chambers 16 and 18 at the lower part of the body In the position indicated in FIG. 7, the rotors are in sealed contact along parts of the quadrants IV of each rotor in order to sealingly isolate the inlet from the discharge, the gas or compressed air being in the gap existing between the rotors being discharged by via the discharge orifice 14 as the closed volume existing between the rotors continues to reduce in volume during the rotation of the rotors.

Fig. 8 represents the rotors which have further progressed during their rotation; we see that the rotors continue to remain in sealed contact along the surfaces of quadrants IV as the volume delimited between the rotors continues to decrease and as the compressed air is discharged through the discharge orifice The rotors continuing to rotating, all trapped and compressed air in the decreasing volume between the rotors is discharged through the discharge port and essentially no communication is established between the inlet port and the discharge port. As the rotors continue to rotate beyond the position shown in fig. 8, the discharge orifice is completely closed by the quadrants I and II of the rotor 22. The rotors continuing their rotation to return to the position indicated in fig. 5, the inlet or suction port remains open while the discharge port is closed. When the rotors have returned to the position indicated in fig. 5, the suction port is again closed and isolated from the compression volume and another compression and discharge cycle is initiated.

 It is important to note that, during the compression cycle, the segments of arcs of a circle which are at the tip of each of the lobes 54 and 58 press against the interior walls of their respective chambers, and the corresponding parts of the rotor profiles located in corresponding quadrants are in sealed contact between the axes of rotation of the two rotors, so that the inlets are always sealed from the discharge port in order to prevent backflow of flow. This tightness between the corresponding points of the rotor profiles is established at the point, or near the point where the rotor profiles intersect an imaginary line connecting the axes of the rotors as they rotate.

To obtain optimal compressor performance, it is very important to have an effective sealing system. To this end, there is provided, in accordance with the present invention, a sealing system which comprises seals placed between the rotors rotating in opposite directions, seals placed between the rotors and the walls of the cylinders, seals for the axial shafts rotors and seals placed between the end faces of the rotors and the faces of the end casings.

Figures 3 and 5 show the sealing system provided between the two rotors and also between the rotors and the cylinder walls. Sealing between the two rotors is ensured by bands 70 formed of a material capable of abrasion and placed on the surface of the discharge rotor 22. These bands of material capable of abrasion, which could alternatively be arranged on the suction rotor 20 are spaced from each other on the outer surface of the rotor profile and they extend from the start of quadrant II over all of quadrants II and III and about 60C of the quadrant IV. These bands extend over the entire axial depth of the rotor profile (ie over the entire length of the rotor between the end housings). These strips of material capable of abrasion preferably consist of a solder (for example a solder containing 50% lead and 50% tin) or else they can be formed of teflon, a coating of graphite and The important characteristic is that these strips of material are relatively soft and suitable for abrasion so that they can be easily lapped or worn out by the sliding action occurring between the rotor 22 and the rotor 20 at the point of minimum approximation between the two rotors (that is to say at the place where the rotors meet for sealing on or near the line drawn between the axes of the rotors).

The strips of material capable of abrasion extend over quadrants II and III and about 60 of quadrant IV, since this corresponds to the contact zones between the rotor 22 and the associated parts of the rotor 20 where the difference in pressure existing between the compressed gas and the suction side makes it necessary to establish a tight seal between the two rotors during an operating cycle. Due to the differences in dimensions and profiles existing between rotors 20 and 22, the rotors always exert a sliding action one on the other, instead of a rolling action, at the point of contact and this relative sliding movement is necessary and extremely effective to promote the "lapping" of the bands capable of abrasion to establish a correct seal between the rotors.

Now considering fig. 5, we will describe the sealing system established between the ends of the rotor tips and the cylinder walls. Sealing strips 72, formed of a material capable of abrasion like the strips 70, are distributed on the inner wall of each cylinder from the low position, corresponding on a clock to the six o'clock position "(as indicated on the Fig. 5), up to the intersection zone of the cylinders near the inlet 42. The bands 72 also extend over the entire length or axial depth of the cylinders between the end housings. When the rotors 20 and 22, the ends of the lobes 54 and 58 adjacent to the cylinder walls rub against the strips 72 so as to produce their wear or running-in and to form the sealing surfaces between the rotor tips and the sealing strips.
wear "is produced by the friction action between the circular arc of 120 provided at the end of each rotor tip and the cylinder walls. This wear is produced by the rotor tips which in fact have a cutting action on the strips 72. This: cutting effect is very effectively achieved by the pointed end of the lobe 58, because its direction of rotation causes the tip to sink into the material of the strips 72. However, the desired cutting action is not carried out as effectively on the tip of the lobe 54 because the pointed edge of the lobe 54 is driven by friction on the strips 72, said pointed end being placed behind, considering the direction of rotation. a cutting notch 73 (also visible in FIG. 4) is formed in the profiled surface of the rotor 20 at the end of the circular arc of 120. This cutting notch extends over the entire axial length of the ' rotor profile and it serves as a cutting edge to form the sealing surfaces in the strips 72 when the rotor 20 passes in front of said strips.

 Alternatively, sealing strips could be placed on the circular arc of 120 provided at the end of each rotor tip and these strips could establish a tight seal against the walls of cylinders (which would not have strips of sealing).

 Now considering figs. 9 (a) and 10, we will describe the sealing system which is provided between the end faces of the rotors and the inner faces of the end casings 24 and 26. The sealing system established between each of the opposite faces of the rotor and its adjacent extreme casing is the same so that it will suffice to describe one, since the same structure is used in both cases. Fig. 9 shows the end casing 24. This end casing has two holes 76 and 78 intended to receive the shafts 28 and 30. Around each opening for the passage of the shaft, an annular recess is provided in the end casing. sealing 80 is positioned in each of the recesses, the seals 80 being each fixed in place by four countersunk head screws 81. The upper surface of each seal 80 comprises a set or pattern of concentric rings of sealing material 82.

The concentric rings 82 are made of a material capable of abrasion, such as brazing, etc., like the sealing strips 70 and 72. The concentric rings 82 delimit recesses 84 between them so that it there is in fact a series of steps or labyrinth seals which surround the tree passage openings so as to stop the air and minimize leaks. A set or pattern of sealing strips 85, also formed from the same type of material capable of abrasion, protrude radially outward from each of the groups of concentric rings around the tree passage openings, and the discharge port.

is also surrounded by a set of sealing strips.

The set of sealing bands surrounding the discharge orifice 44 comprises a first band 86 which starts from the concentric rings placed in the vicinity of the shaft seal, along the lateral and circumferential parts of the discharge orifice. The discharge port seal also includes a second sealing strip 88, which is slightly spaced outward from the first strip 86 and which follows the pattern of this first strip; the second strip 88 has a series of V-shaped sealing parts or saw teeth 90 which protrude radially outward from the body of the strip 88, around the outer periphery of the discharge orifice 44. The end faces of the rotors 20 and 22 are pressed against all the sealing strips shown in FIG. 9 so as to ensure that the sealing surfaces are run in and to establish an effective seal between the end faces of the rotor and the sealing strips of the end casings.

The sealing strips 85, 86, 88 and 90 are all applied against the surface of the end casing 24.

 In fig. 9 (b), a modification of the sealing system of FIG. 9 (a), where the concentric rings 82 are interrupted by radial bars 83 in order to form a series of separate pockets or cells which further contribute to trapping the leakage air. In FIG. 9 (b), the fixing screws 81 have been eliminated.

 In fig. 10, an important detail has been highlighted of the structure with concentric sealing joints which is provided around the shaft passage openings 76, 78. The concentric circular sealing strips 82 have the shape of trapezoidal volumes comprising sides inclined outwards by an angle of approximately 15. (ie which make an angle of 750 with the face of the corresponding extreme casing). Associated concentric grooves 94 are formed in each of the faces of the rotor. These grooves 94 are concentric circles surrounding the rotor shafts and they are positioned so as to correspond to the protruding bands 82 provided on the end casing. The concentric grooves 94 have sides inclined towards the outside by an angle of approximately 50. Consequently, when a strip 82 coincides with a groove 94, this strip penetrates only partially in the groove and a tight contact is made between the sides, of different inclinations, of the strips and grooves. The concentric grooved structure which is provided around the rotor shaft exerts an action of commat; lapping on the sealing rings 82 and / or the bars 83 provided on the extreme casing in order to form an effective sealing device, which in fact constitutes a series of labyrinth seals. If the rotor is moved away from the extreme rotor, the leakage gap thus created between the inclined faces of the seal is always less than the degree of axial displacement of the rotor, which makes it possible to maintain a good seal in a wide range of operating conditions.

 In all cases, the sealing system is formed by a relatively soft material, capable of abrasion or crushing, such as a solder, etc.

This material is easily worn or lapped by the action of the rotors in order to form effective seals which are profiled in correspondence with the profiles of the different
surfaces of rotor elements, including irregularities.

After wear, the height of the sealing strips can be
between 0.025 and 0.05 mm. This sealing system
used to manufacture the various elements of the compressor,
such as rotors, cylinders and extreme housings,
with wider tolerances and further these elements
can operate in a wider range of temperatures
without any mechanical disturbance between the rotors or leakage occurring
which could produce a reduction in performancies and a dangerous accumulation of heat and
expansion likely to cause the machine to seize.

The sealing strips 70 provided on the rotor 22 established
feels an effective seal between the movable rotors while
avoiding direct contact between them. Although the sealing strips can have a height as low as 0.025 to 0.05 mm after wear or lapping, they are not used
only to establish a seal between the rotors in
their minimum point of approximation but they reduce
also the leakage interval between the rotors on the hand
and across from the minimum point of approach, in view
amplify the sealing effect.

Due to the existence of this very elaborate sealing structure, the compressor can, during
break periods, operate under conditions
which produce a heating of the rotors beyond the
normal operating temperature. It follows that
causes crushing and abrasion of the seals
cheeness during these break periods, which leads to
the establishment of nominal tolerances in normal operation.

During normal running periods, the performance of the compressor can therefore be better predetermined since the seals were brought to their final profile during the pause period. Also, penetration of sealant into the normal running fluid stream does not occur. As indicated above, the heights of the sealing strips can reach values as low as 0.025 to 0.05 mm after running in. The total height and width of the sealing joints can vary from one position to another at the inside the machine and the dimensions of the seals must be controlled so that the number of seals that have been subjected to crushing or abrasion during periods of pause does not result in disturbing mechanical or thermal stresses on the rotors, bearings and other compressor components.

Considering again Figures 2, 3 and 4, we see that there is shown a balancing system of the rotors. Holes are drilled and tapped in each rotor (parallel to its axis). These holes are then plugged using threaded pads 100 which are fixed by pointing or by other means. The balancing holes are drilled completely through the rotor in the axial direction and the pads are inserted into each opposite end face of the rotor, the outer surface of each pad coming flush with the face of the rotor and intervals being between the opposing pads. Three buffers 100 have been shown in one face of the rotor 22 and four buffers 100 in one face of the rotor 20.

As shown in Figures 2, 3 and 4, the pads can have different dimensions and they are distributed around the axis of each of the rotors in an appropriate manner, as follows. which may be necessary to ensure general balancing of each particular rotor. Preferably, two buffers provided on each face of each rotor are located on the mutually perpendicular axes (as indicated for buffers 100 (a) and 100 (b) on rotor 20 and buffers 100 (c) and 100 (d) on rotor 22). it is then possible to carry out a precise balancing of the rotors by changing the weight of one or the other, or else of the two mutually perpendicular buffers, by operating either by replacing one buffer with another to increase the weight, either by removing material from the inside of a tampon.
In this way, the centrifugal force can be varied positively or negatively, depending on what may be necessary for the precise balancing of each rotor. The arrangement of the buffers on mutually perpendicular axes simplifies the calculations necessary for determining the weights of the buffers with a view to achieving precise balancing. By having precise balancing buffers on mutually perpendicular axes, a change to a buffer has no influence on the other axis. Consequently, precise adjustment can be made by simply adjusting the buffers placed on the mutually perpendicular axes.

 Of course, the invention is not limited to the embodiments described and shown above, from which other modes and other embodiments can be provided, without going out.

of the scope of the invention.

CLAIMS 1. Rotary compressor, of the type comprising profiled rotors rotating in intersecting cylindrical holes, characterized in that it comprises a first cylinder (16), a second cylinder (18), an input rotor (20) placed in the first cylinder (16), comprising a hub (52) and a lobe (54) and rotating on a first shaft (28), a discharge rotor (22) placed in the second cylinder, comprising a hub (56) and a lobe (58) and rotating on a second shaft (30), an input means (42) provided in at least said first cylinder (16) for ensuring the admission of the gas to be compressed, a discharge means provided in the second cylinder (18) for discharging the gas after compression, said discharge means comprising an orifice (44) formed in at least one end plate of the second cylinder (18) and comprising an outer circumferential arc of a maximum radius "r" which is determined by the profile of a segment of said discharge rotor (22) and an inner circumferential arc ur of minimum radius slightly greater than the radius of the second shaft, in that said discharge rotor (22) comprises four quadrants (I, 11, III, IV> starting from the point of the lobe (58) of said discharge rotor (22) and extending successively around said discharge rotor (22), the second quadrant (II) being an arc of a radius "R", which is slightly greater than said radius of the discharge orifice (44), said radius gtR9 'being selected to increase. at most the radius' r "of the discharge orifice while avoiding excessively high flow velocities in said second cylinder (18), in that the first (I) and the third (III) quadrant each have a dimension "RE ;;" at their junction with the second quadrant (IÌ) and each have contoured contours of a different dimension of the radius "R" on at least the major parts of said first and third quadrants (I, III), and in that said input rotor tel6) and said discharge rotor (22) have different dimensions and profiles, the discharge rotor (22) having dimensions greater than those of the input rotor (20).

2. Rotary compressor according to claim 1, characterized in that said first quadrant (I) of the discharge rotor (22) starts from the tip of said lobe (58) of said discharge rotor (22) and comprises a first part having the shape an arc of a nominal radius R, approximately equal to the radius of the second cylinder, as well as a second part having the shape of an ellipse, said ellipse being modified if necessary in order to adapt the slope of said part in the shape of an arc of a circle with said ellipse, in that said ellipse has the dimension "R" at the end of said first quadrant (I) and is perpendicular to a radius "R" at the end of said first quadrant, in that the third quadrant (III) of said discharge rotor (22) is an ellipse which is determined at least in part by factors of dynamic equilibrium and flow velocities, and in that the fourth quadrant (IV) of said discharge rotor (22) is generated by corresponding surfaces of said r input otor (20).

3. Rotary compressor according to one of reven dicati.ons 1 and 2, characterized in that said input rotor (20) comprises four quadrants (I, II, III, IV) which are related to the four quadrants of discharge rotor (22), in that the first quadrant (I) of the input rotor (20) has a first part which has an arbitrarily chosen length so as not to interfere with the corresponding part of the first quadrant of the discharge rotor and a second part which is generated by a part of the ellipse of the second discharge rotor part (22), in that the second quadrant (II) of the input rotor (20) is an arc of radius "R2" equal to the difference between the distance separating the axes of said first and second rotors and the dimension '2R "of said second quadrant (II) of said discharge rotor (22), in that the third quadrant (III) of the rotor d input is generated by the ellipse of the third quadrant of the discharge rotor, and in that the fourth quadrant (IV) of the rotor input (20) is formed by three arcs of radius 'lu41', "R3" and "R1 ,,, the arc of radius" R1 "starting from the tip of the lobe (54) of the input rotor (20) and having a radius approximately equal to the radius of the first cylinder (16) while the arc of radius' vu3 "has its center placed on a radius of the arc of radius 1 and that the arc of radius "R4, 'is centered on a line which defines the junction between the third and fourth quadrants.

4. Rotary compressor according to claim 3, characterized in that said arcs of the fourth quadrant (IV) of said input rotor (20) have slopes equal to the junctions between the arcs.

5. Rotary compressor according to any one of claims 1 to 4, characterized in that it comprises sealing means placed between said rotors (20, 22) and
es sealing means placed between said rotors (20, 22) and said cylinders (16, 18).

6. Rotary compressor according to claim 5, characterized in that -said sealing means placed between the rotors (20, 22) comprise strips (70) of an abrasion-resistant material which are arranged on at least a part of the profiles of each rotor, said strips (70) being worn or lapped during operation in order to establish the seal between the rotors.

7. Rotary compressor according to any one of claims 1 to 6, characterized in that it comprises a plurality of balancing holes which are formed in opposite surfaces of said rotor as well as balancing buffers (100) placed in said balancing holes.

8. Rotary compressor according to claim 7, characterized in that at least two of said balancing holes are placed on mutually perpendicular axes and in that the buffers corresponding to the two aforementioned balancing holes are selectively modifiable in view to ensure precise balancing of said rotor.

9. Rotary compressor according to claim 8, characterized in that at least one balancing hole is provided, comprising buffers, in addition to the two aforementioned balancing holes.

10. Rotary compressor according to claim 8, characterized in that one of the mutually perpendicular axes starts from the tip of said lobe and intersects the axis of rotation of said rotor.

11. Rotary compressor according to claim 10, characterized in that the two aforementioned balancing holes are formed in the hub of said rotor.

12. Rotary compressor, of the type comprising rotors (20, 22) rotating in opposite directions in intersecting cylinders (16, 18), characterized in that the rotors each comprise a hub and at least one lobe and in that a sealing structure is provided comprising strips (70) formed of a material suitable for abrasion and placed on at least part of the profile of at least one rotor, said strips (70) made of material suitable for abrasion being worn or lapped during operation to establish a seal between 1 rotors, as well as means for establishing a seal between the rotors and the cylinders (16, 18).

13. Rotary compressor according to one of claims 5 or 12, characterized in that said sealing means provided between the rotors (20, 22) and the cylinders (16, 18) comprise bands (72) formed of a material suitable for abrasion and placed on the walls of said cylinders (16, 18), said strips being worn or lapped by the tips of said rotors in order to establish a seal between said rotors (20, 22) and said cylinders (16, 18).

14. Rotary compressor according to claim 13, characterized in that there is provided in one of the lobes (54) a cutting notch (73) intended to act on said strips (72) of material capable of abrasion being on the walls of the cylinders (16) which said lobe performs its rotational movement.

15. Rotary compressor according to one of claims 5 or 12, characterized in that said sealing means provided between said rotors and said cylinders comprise sets of sealing strips suitable for abrasion and placed on opposite end plates said cylinders (16, 18), said end plates being placed in positions immediately adjacent to end faces of said rotors, said set of abrasion-resistant bands which are provided on each end plate comprising an annular network of several bands of sealing (82) spaced apart from each other and surrounding a shaft passage opening formed in each end plate, and in that said set of strips of material capable of abrasion comprises stripe patterns (85) - formed of said material and radiating outwards from each annular network.

16. Rotary compressor according to claim 15, characterized in that it comprises a discharge orifice (44) formed in an end plate of a cylinder and in that said sealing structure comprises a sawtooth pattern (90 ) formed of a sealing material suitable for abrasion, said pattern being placed around at least part of said discharge orifice (44).

17. Rotary compressor according to claim 15, characterized in that it comprises a plurality of grooves (94) formed in opposite faces of the rotor so as to coincide with said bands (82) of the annular network of bands surrounding the openings shaft passage, said strips having sides.

inclined (92) by a first slope while said grooves have inclined sides (96) by a second.

slope.

18. Rotary compressor according to claim 12, characterized in that said sealing means provided between said rotors (20, 22) and said cylinders (16, 18) comprise bands formed of a material suitable for lssabra- sion and placed on the lobe in an area adjacent to the tip of the lobe.

19. compressor rotor comprising a hub (52, 56) arranged to be mounted for rotation on a shaft (28, 30), a lobe (54, 58) provided on said hub, said hub and said lobe having opposite surfaces, characterized by a plurality of balancing holes in opposite surfaces of said rotor and balancing pads (1002 placed in said balancing holes.

20. Multi-stage rotary compressor, characterized in that it comprises a first stage rotary compressor (10), a second stage rotary compressor (12), said first and second stage rotary compressors being arranged so that one stage is positioned above the other, and a common drive means (37) which is coupled. to each of said rotary compressors of the first and second stage to ensure their drive.

21. Rotary compressor i ~ .multiple stages according to claim 20, characterized in that the first stage compressor (10) comprises a first and a second rotating element which are driven by a first and a second synchronization gear in engagement with each other, in that the second stage compressor (12) has a first and a second rotating element which are driven by a first and a second synchronization gear engaged with each other and in that said common drive means (37) is a driving gear which is coupled to one of said synchronization gears of each stage.

22. Multi-stage rotary compressor according to claim 21, characterized in that said first and second gears of one stage are placed above the first and second corresponding synchronization gears of the other stage and in that said driving gear is positioned on one side of said synchronization gears.

Claims (3)

1. - Multi-stage rotary compressor1 characterized in that it comprises a first stage rotary compressor (10), a second stage rotary compressor (12), said first and second stage rotary compressors being arranged so that one stage is positioned above the other, and a common drive means (37) which is coupled to each of said rotary compressors of the first and second stage to ensure their drive. 2. - rotary multistage compressor according to claim 1, characterized in that the first stage compressor (103 comprises a first and a second it is rotated which are driven by a first and a second synchronization gears engaged one with the other, in that the second stage compressor (12) has first and second rotating elements which are driven by first and second synchronizing gears in engagement with each other and in that said means common drive (37) is a driving gear which is coupled to one of said synchronization gears of each stage. 3. - Multi-stage rotary compressor according to claim 2, characterized in that said first and second gears of one stage are placed above the first and second corresponding synchronization gears of the other stage and in that said driving gear is positioned on one side of said synchronization gears.