MXPA98006324A - Mechanism of placement driver with adjustment of kickback for diffuser of variable pipes - Google Patents

Abstract

The present invention relates to a centrifugal compressor having a cover and an impeller rotatably mounted therein for carrying a working fluid from an inlet to the inlet of a radially disposed annular fissure ring diffuser, the diffuser includes a ring internal, the inner ring has a plurality of first sections of flow guide channel formed therein, one outer ring, the outer ring has a plurality of second flow guide channels formed therein, each second channel section of The flow guide has a first complementary flow guide channel section, the compressor includes a driving positioning mechanism for rotating the inner ring circumferentially within the outer ring between a first fully open position wherein the first and second channel sections of complementary flow are aligned to allow a maximum flow of fluid through the sections of complementary channels, and a second partially closed position wherein the first and second complementary flow guide channels are misaligned to restrict fluid flow through the complementary channel sections, the driving positioning mechanism is characterized in that it comprises: a driving means rotary mounted fixedly to the cover, having a driving end and a differential pinion mounted thereon; a rack gear fixedly mounted to the inner ring extending radially outwardly from the inner ring and adapted to be engaged in gear arrangement with the differential pinion, and a travel limiter for limiting the travel of the inner ring between a first position corresponding to the fully open position and a second position corresponding to the partially closed position, the travel limiter comprises a pair of brake blocks arranged in The outer ring for coupling d e) rigidly engaging the rack gear in a first position corresponding to the fully open position and for rigidly coupling the rack gear in a second position corresponding to the partially closed position.

Description

IMPULSER PLACEMENT MECHANISM WITH COUPLING ADJUSTMENT FOR VARIABLE PIPE DIFFUSER DESCRIPTION OF THE INVENTION The present invention relates to compressors centrifugal in general and in particular to a driving mechanism and a kickback adjustment mechanism for a variable pipe diffuser for a centrifugal compressor. * One of the main problems that arise with the use of centrifugal steam compressors for applications where the compressor load varies over a wide range, is the stabilization of the flow through the compressor. The inlet passages, the impeller and the compressor diffuser must be sized to provide the flow velocity maximum desired volumetric. When there is a low volumetric flow rate through such a compressor, the flow becomes unstable. As the volumetric flow rate decreases from a stable range, a slightly unstable flow range enters. In this range, there seems to be a partial inversion of the flow in the passage of the diffuser, creating noises and decreasing the efficiency of the compressor. Below this range, -the compressor enters what is known as an inrush, where there are periodic complete flow reversals in the diffuser passage, destroying the efficiency of the machine and endangering the integrity of the elements of the machine. Because a wide range of volumetric flow rates is desirable in many compressor applications, numerous modifications have been suggested to improve flow stability at low volumetric flow rates. Many schemes have been devised to maintain high efficiencies of the machine over a wide range of operation. In U.S. Patent No. 4,070,123, the entire configuration of the driving wheel varies in response to load changes in an effort to match the stress of the machine with the changing load demands. Adjustable diffuser flow restrictors are also described in U.S. Patent No. 3,362,625 which serves to regulate flow within the diffuser in an effort to improve stability at low volumetric flow rates. A common technique - to maintain high operating efficiency over a wide range of flow in a centrifugal machine is through the use of the variable-width diffuser along with fixed diffuser guide fins. U.S. Patent Nos. 2,996,996 and 4,378,194, issued to a common assignee, disclose diffusers with variable width fins where the diffuser fins are fixed securely, bolting them to one of the walls of the diffuser. The fins adapt to pass through the openings formed in the other wall, thus allowing the geometry of the diffusers to change in response to changing load conditions. The fixed mounting of the blades of the diffuser 5 to one of the walls of the diffuser presents a number of problems in particular with respect to the manufacture, maintenance and operation of the machine. Little space is left to secure the fins in the assembly. Any misalignment of the fins will cause the fin to lock or rub against the opposite wall as it returns to 'stand. Similarly, if one or more fins in the series have to be replaced in assembly, the entire machine usually has to be disassembled in order to perform the replacement. The efficiency of a compressor could be improved in greatly varying the geometry of the diffuser output. In the commonly assigned U.S. Patent Application Serial No. 08/658801, a variable geometry pipe diffuser is disclosed. This application is incorporated herein by reference. A diffuser Variable geometry pipe, divided (which may also be termed a split ring pipe diffuser) of the diffuser in a first inner ring and a second outer ring. The inner and outer rings have complementary inlet channel sections formed in the same. That is, each inlet flow channel section * of the inner ring has a complementary inlet flow channel section formed in the outer ring. The inner ring and the outer ring are rotatable with respect to each other. The rings are rotated to improve efficiency to vary the pressure levels between a fully open position and a partially closed position. In the partially closed position, the misalignment of the outlet pipes of the diffuser causes an increase in noise. The rotation of the rings above, result in excessive noise and efficiency degradation. The geometric tolerances within a centrifugal compressor are small. At the same time, the loads inside the compressor are large and dynamic in nature. In a split ring pipe diffuser, the problem of maintaining the tolerances in the phase 'of the dynamic loads is very «. expensive. Both axial loads (thrust) and circumferential loads on the pair of rings must be handled. Diffuser rings must be able to rotate one respect of the other and at the same time controlling the adjustment on its relative position, it must be maintained in order to ensure the proper alignment of the flow channels and the ultimate efficiency of the compressor. The cost of maintaining the required tolerances in a split ring diffuser is usually very high.

# Another problem with split ring diffusers is the premature wear of the parts. In general, lubricants are not used within the gas flow regions of the centrifugal compressors to avoid contamination of the gases. The dynamic loads imposed on the split ring diffuser by the flow of gas leaving the impeller, cause wear on the components of the diffuser that will: accelerate due to the absence of lubricating oil. The drive system to position accurately the rings one with respect to the other must, among other things, be rigid to avoid any chafing of the components. Due to the circumferential load on the rings, there is a tendency for the inner ring to oscillate with respect to the outer ring that can cause instability of the compressor, wear of the parts and can adversely affect the efficiency. This causes various problems that need to be overcome. A drive system is needed that is able to prevent relative movement between the inner and outer rings. You also need a bearing concept that It allows the relative rotation of the two rings and is also capable of supporting the thrust and circumferential loads while maintaining the geometric tolerances adjusted between the rings. There is also a need to provide a placement system that includes minimum and maximum caps positives to avoid unnecessary noise and degradation * of efficiency as well as simple field feedback. In addition, there is a need for the bearing and bearing systems to have a long operating life and to be easy to install and adjust properly. In accordance with its main and widely established aspects, the present invention relates to a variable geometry pipe diffuser for a centrifugal compressor. More specifically, the present invention relates to an impeller positioning mechanism and a associated kickback adjustment mechanism for use in a variable geometry pipe diffuser for a centrifugal compressor. A rack gear is attached to the inner ring of a variable pipe diffuser of the present invention.

A differential pinion is fixed to a rotary drive means mounted to the housing of a centrifugal compressor in the rack gear arrangement. The rotary drive means is operated to rotate the inner ring relative to the outer ring between a position in where the pipes of the diffuser are completely open and a position where the pipes of the diffuser are partially closed. In addition, the driving means are capable of aligning the pipes of the diffuser, with respect to a plurality of positions between the positions completely open and partially closed. A * travel indicator is also provided to positively limit the travel of the inner ring in fully open and partially closed placements. The kickback adjustment mechanism has a concentric housing around a first centerline and a hole therethrough, concentric around a second center line. The differential pinion is mounted to a driving shaft that passes through the hole. The housing is mounted rotatably to the roof. The housing of the present invention is rotatably operable around the second center line to adjust the kickback between the rack gear and the differential pinion. BRIEF DESCRIPTION OF THE DRAWINGS In the drawings, where similar numbers 15 are used to indicate the same elements through the views; Figure 1 is a cross-sectional side view of the compressor according to the invention having a variable pipe diffuser according to the present invention; Figure 2 is a perspective view of a variable pipe diffuser according to the invention; Figures 3 and 4 'are cross sectional front views of a variable pipe diffuser according to the invention in a first fully open position and a second partially closed position, respectively; Figure 5 is a top view of a compressor having a variable diffuser of the present invention; Figure 6 is a cross-sectional view of a ring support mechanism of the present invention taken substantially along line 6-6 of Figure 5; Figure 7 is a cross-sectional view of a ring support mechanism of the present invention taken substantially along line 7-7 in Figure 6; Figure 8 is a cross-sectional view of a roller assembly of the present invention; Figure 9 is a cross-sectional view of an axle of the present invention; Figure 10 is a cross-sectional view of a positioning drive of the present invention of the area 10 in detail in Figure 1; Figure 11 is a top view of a positioning drive of the present invention; Figure 12 is a perspective view of a rack gear of the present invention; Figure 13 is a performance diagram for a variable pipe diffuser according to the present invention; Figure 14 is a performance diagram for a compressor having only inlet guide fins; Figure 15 is a performance diagram for a compressor according to the present invention having a variable pipe diffuser and inlet guide fins; and Figure 16 is a cross-sectional view of a compressor having an axial restriction mechanism in accordance with the present invention. Referring now to Figure 1, the invention is shown installed in a centrifugal compressor 10 as part of an HVAC system (not shown) having an impeller 12 for accelerating the refrigerant vapor at a high speed, a diffuser 14 for decelerating the coolant at a speed • low while converting the kinetic energy 'to pressure energy, and a full discharge in the form of a collector 16 to collect the discharge vapor for subsequent flow to a condenser. The power to the impeller 12 is provided by an electric motor (not shown) which is hermetically sealed at the other end of the compressor and which operates to rotate a high-speed arrow 19. With reference now to the manner in which the Refrigerant flow in the compressor 10, the coolant # enters the inlet opening 29 of the suction housing 31, passes through the paddle ring assembly 32 and the guide vanes 33, then enters the compression suction area 23 which leads to the defined compression area 5 on its inner side by the impeller 12 and on its outer side by the housing 34. After compression, the refrigerant then flows to the diffuser 14, the manifold 16 and the discharge line (not shown). A diffuser 14 of variable geometry tubing according to the present invention includes a first inner ring 40 and a second outer ring 42, a ring support mechanism 35, and a positioning drive mechanism 121. With reference to Figures 3 and 4, the inner and outer rings have sections 44 and 46 of flow channel complementary formed in them. That is, each section 44 of the flow channel of the inner ring 40 has a complementary channel section 46 formed in the outer ring 42. The inner ring 40 and the outer ring 42 are rotatable with respect to each other. In a preferred embodiment, the ring internal 40 rotates circumferentially within a fixed outer ring 42. When one ring is rotated with respect to the other, the alignment between each pair of complementary inflow channels of the inner and outer rings changes as seen with reference to Figures 3 and 4.

The rings 40 and 42 are adjustable between a first fully open position, as illustrated in Figure '3, where the complementary channel sections are aligned and a maximum amount of fluid passes through the inner and outer rings 40 and 42 and a second partially closed position, as illustrated in Figure 4, wherein the complementary channels are misaligned and flow is restricted through the channel sections 44 and 46. In Figure 5, a ring support mechanism 35 according to one embodiment of the present invention is shown. The embodiment shown illustrates the use of three such separate equidistant mechanisms circumferentially around the diffuser. With reference to Figures 6-7 ,. The ring support mechanism of the present invention includes an internal bearing groove 41 and a recess 43 disposed in the inner ring 40, a roller assembly 54, a roller shaft assembly 36 and an outer bearing groove 45 disposed in the outer ring. The roller assembly as shown in Figure 8 includes a roller 55 having an outer bearing surface 56, and a pair of thrust bearing surfaces 57. The axle assembly as shown in Figures 6-7 includes a shaft 37 and a bolt shaft 39. As seen in Figure 9, shaft 37 includes a hexagonal head 38 and an axle body 37, a center line of the axle body 48, an axle hole 49 and a center line of the hole of the shaft 50. In addition, the shaft 39 includes a pair of concentric shoulders 73, 74, with respect to the center line of the hole of the shaft 50. Another problem with the split ring diffusers is the premature wear of the parts. In general, no lubricant is used within the gas flow regions of the centrifugal compressors to avoid contamination of the gases. The dynamic loads imposed on the split ring diffuser by the flow of gas leaving the impeller causes wear on the components of the diffuser that will accelerate due to the absence of lubricating oil. Due to the non-availability of lubricating oils in most compressors, it is usually necessary to take measures to minimize friction and wear due to chafing. Accordingly, in certain embodiments of the present invention and as described hereinafter, the interfaces of the components are hard coated, the parts are fabricated from ultra high molecular weight plastic materials, the ring assemblies are pre-load and eliminates the kickback from the gears of the positioning drive system. With reference now to the manner in which the inner ring is assembled and its movement. The outer ring 42 is fixed with respect to the suction housing and three sets of ring support mechanisms 35 * are installed in the outer ring by placing the ring assembly 54 into the bearing groove of the outer ring 45, passing the shaft through the mounting hole 58 and the roller assembly and then installing the shaft bolt 39 through the shaft 5 and loosely threading the shaft bolt 39 into the threaded holes in outer ring 59. The inner ring '40 is installed inside the outer ring with the recesses 43 of the # inner ring aligned circumferentially with bearing groove 45 and roller assemblies 35 and after by rotating the inner ring in the clockwise direction as shown in Figure 7 to place the roller assemblies inside the bearing slot 41. With the inner ring installed inside the outer ring, the supporting mechanism ring are employed to properly center and position the inner ring by rotating the shaft by using a wrench placed on the hexagonal head 38. The center line 38 of the axle body 48, on which the roller 55 is mounted, is sent from the line central 50 of the shaft hole, which is concentric with the shoulders 73, 74, by 0.053 cm (.021 inches). The rotation of the hexagonal head 38 causes the roller assembly to rotate around the shoulders within the outer ring and cause the roller assembly to move radially relative to the outer ring. Once the ring internally is appropriately centered within the outer ring, the hexagonal head is further rotated to pre-load the outer bearing surface 56 of the roller assemblies against the inner ring. The pin of the shaft 39 is then tightened. The conditioned preload is preferred since it prevents the inner ring from moving due to the tangential and circumferential loads. In one embodiment of the present invention, the roller 55 and the inner ring # 40 are aluminum and both the outer bearing surface 56 and the internal bearing groove 41 are hardened to avoid wear. The roller assemblies restrict the movement of the rings to the axial direction due to the thrust loads by placing the thrust bearing surfaces 57 within the hardened internal bearing groove 41 and the outer bearing groove 45 relatively. soft. • The thrust bearing surface 57 of the roller assembly must allow rotation of the inner and outer rings, and at the same time support the thrust loads produced by the compressor. In a preferred embodiment, the thrust bearing surface 57 is made of plastic from ultra high molecular weight having a low coefficient of friction of 0.16 and a hardness of 64 on the Shore D scale. The plastic thrust bearing surfaces prevent contact between the hardened roller and the soft outer bearing groove and used to carry the thrust loads of the compressor and to adjust the axial tolerances of the * inner ring. An additional feature of the ring support mechanisms is that with the rings assembled as described above, it is possible to previously assemble the inner and outer rings and transport them to the compressor to finally assemble them. Another embodiment of the present invention for limiting and preventing axial movement of the inner ring # with respect to the outer ring is shown in Figure 16. An axial restriction system 90 is shown comprising a threaded shaft 91, a threaded mounting hole 92, a bearing pad 93, a counter nut 94, a hexagonal head 95 and a recess 96. During the assembly of the diffuser, the axial restriction mechanism 90 is installed so that the bearing pad 93 is placed in the recess 96. The bearing pad placed inside the recess leaves space for the cover 34 to be mounted - # external ring 42 without accidental contact of the bearing pads with the inner ring. Once the housing 34 is installed, the threaded arrow 91 is made rotate to contact the bearing pad with the inner ring. With the bearing pad properly positioned, the mechanism is releasably clamped against the counter nut 94. In a preferred embodiment, the bearing pad is made of a plastic material of ultra high molecular weight. One embodiment of the present invention includes six such axial restriction mechanisms equally spaced in a circle around the inner ring. A positioning drive mechanism 121 for rotating the inner ring 40 circumferentially within the outer ring 42 is described with reference to Figure 10. The outer ring 42 has the rack gear 123 fixedly attached thereto, which extends radially outwardly from the outer ring 42. In relation to gear with the rack gear 123 is the differential pinion 124 which is driven by the differential pinion shaft 126 by the actuator 128. The actuator 128 is selected and controlled to effect the movement of the inner ring 40 relative to the outer ring 42 between a first fully open position and a second partially closed position and any number of positions • # intermediate. The shaft 126 is housed in a containment housing 130 that hermetically seals the shaft 126 from the interior of the compressor 132 and which prevents leakage out of the flow from the compressor 10 through the containment housing 130. The tangential and circumferential load on the rings by the cooling flow within the diffuser causes the inner ring to rattle from front to back within the outer ring. Excessive movement or rattle of the inner ring causes the • gear of the rack 123 and the differential pinion 124 to rub together and also cause other parts to wear out. The pre-loading of the inner ring through the roller assemblies as discussed hereinabove prevents movement 5 of the inner ring as well. the rattle under normal operating conditions. In cases of abnormal conditions, such as operation in an inrush, a * secondary mechanism to prevent movement of the inner ring. The present invention provides a system of impeller assembly to prevent reverse movement and rattle of the inner ring preventing kickback between the toothed section and the differential pinion by adjusting the relative center positions of the differential pinion and the rack gear using the containment housing 130 of the axis. The external surface 125 of the shaft housing is concentric around the center line 127 and the * W hole of housing 129 is concentric about the center line of housing bore 131. In one embodiment of the present invention, the centerline of the housing housing 127 and the centerline of the housing bore 129 is 1524 cm (.060 inches). With reference to Figure 11 the flat parts for allowing a nut 135 and the adjustment slots 134 of the position drive are shown. After the installation and is placed in the suction housing 31, the kickback ¥ between the rack gear 123 and the differential pinion 124 is removed by rotating the positioning drive, by placing a wrench (not shown) through the flat portions for allowing to put a nut 135. Once a minimum backlash is achieved, the positioning drive is fixed in place by adjusting the cover screws 133. Once the kickback is eliminated, the tendency of the inner ring to move is discharged directly by the actuator through the system. gear. The flow of fluid through the diffuser 14 in a second partially closed position with respect to the flow velocity of the fully open position, is determined by the ratio of the minimum cross-sectional area of a flow channel of a diffuser, in a partially closed position with respect to the minimum cross-sectional area of a flow channel (defined by the complementary channel sections 44 and 46) in a fully open position. This area of minimal flow channel, known as "throat area", will usually be determined by the smaller diameter of the flow passage 52 of the inner ring channel 44 when the diffuser 14 is in a fully open position, and will be controlled by the width 53 at the interface between the inner and outer rings 40 and 42 when the diffuser 14 is in a second partially closed position. For example, if a diffuser channel has a minimum area (throat area) of 1/8 sq. in a partially closed second position, in a minimum area (throat area) of 1/4 sq. in. in a fully open position, then the volumetric flow rate of the fluid through a diffuser in the partially closed position will be approximately 50% of the flow rate # as in the fully open position. The flow velocity of the fluid through the compressor 10 when the diffuser 14 is in a partially closed second position, will generally be between about 10% and 100% of the fluid flow rate through the compressor 10 when the diffuser is in the first fully open position. 15 In a second partially closed position (Figure 4), at least about 10% of the flow volume as in the fully open position, should flow through the diffuser 14 in order to avoid excessive thermodynamic heating, excessive noise and a degradation in compressor efficiency. For this purpose, the amount of relative movement between the two ring sections should be limited to a quantity of rotation necessary to effect a second partially closed position. In other words, the rings should not be adjustable for completely shut off a fluid flow between them. The allowable degree of rotation between the two rings is determined by the desired flow between the rings in a fully closed position, and the number and volume of the inlet flow channel sections 44, 46, in the 5 sections of ring 40 and 42 in relation to the volume of the ring sections 40 and 42. Continuing with reference to Figure 4, R-2 defines '^ -r the radius of the tip of the impeller, R3 defines the radius of the inner ring 40 and R4 defines the outer radius of the ring external. When making the thickness, defined by the quantity T = R3-R2 of the inner ring 40 not greater than what is necessary to block a desired portion (for example 50% of the flow) of the flow through the channels 46 of the outer ring, the flow of fluid through the diffuser 14 can be controlled efficiently. The rotation of the inner ring with respect to the outer ring will reduce the throat area of the diffuser before any diffusion is effected, thus preventing acceleration of the flow after diffusion. In addition, the smaller the thickness of the inner ring, the smaller the will be the angles of flow rotation through the diffuser in the partially closed position. Both effects described above tend to improve the efficiency of the compressor under operating conditions of part of the load. With reference now to Figures 5 and 12, shows an embodiment of the present invention having a 'ß-r mechanism for providing positive positioning of the inner ring corresponding to a first fully open position and a second completely closed position. The cavity 137 is machined in the outer ring 42 to accommodate the rack gear 123. The rack gear 123 is precisely mounted to the inner ring 40 in the form of a tongue and groove wherein the rack gear is provided with a notch. circumferential 143 adapted to receive the tab section 139 of the inner ring 40.

To determine the fully open position, the inner ring is placed inside the outer ring and the rings are rotated relative to each other until the flow passages 52 are completely aligned with the external flow channels 46. With the rings in this position, and the ring support mechanism adjusted as described hereinabove, the rack gear is mounted to the inner ring with the face 145 of the gear in contact with the fully open stop 140 of the cavity 137. then install bolts (not shown) to through the gear mounting holes 142 and securely and tightly into the threaded holes in the inner ring 138. The rack gear and the cavity are sized to provide a predetermined amount of pipe diffuser closure. For example, in a modality of the present invention is dimensioned so that the difference between the angular width of the rack gear and the cavity provide a 10% open position. In this example, the required travel of the rack gear is 10 degrees, the angular width of the rack gear is 35 degrees and the angular width of the corresponding cavity is 45 degrees. With the rack gear positioned in this manner, a positive stop is created between the rack gear and the cavity to accurately and repeably place the rings at the points corresponding to a fully open position and to a partially closed position. The positive stops also allow the field retroajustment of the actuator 128 without the need to adjust the position of the inner and outer rings. The operation and use of the present invention can be understood with reference to Figure 5, which shows a performance diagram for a compressor having a variable pipe diffuser according to the invention integrated therewith. The performance diagram of Figure 5 includes a plurality of performance points 60, 62, 64, 66 and 68, each corresponding to a discrete placement between the inner and outer ring sections 40 and 42. Each performance point, per Example 60, is characterized by an inrush point, for example 70, which is the maximum available pressure point. Operating a compressor at a flow rate that is at or below the point of inrush will most likely result in an inrush condition, as described in the Background of the Invention section herein. For the purposes of illustrating the invention, diagram 60 may correspond, for example, to a first fully open position, diagram 62 may correspond to a partially closed position of intermediate 2 degrees, diagram 64 may correspond to a partially closed position of 4 intermediate degrees and diagram 68 may correspond to a partially closed position of 8 degrees maximum. It is noted that the adjustment ring sections 40 and 42 with respect to a closed position, have the effect of adjusting the point of inrush, for example, 70, 72 in a performance diagram for a compressor with respect to a lower flow rate. Therefore, an inrush condition can be avoided during periods of low flow demand by adjusting the diffuser rings 40 and 42 relative to a closed position. It is helpful to understand the invention for comparing the performance diagram of Figure 5, for a compressor having a diffuser variable with respect to the performance diagram shown in Figure 6 which corresponds to a compressor having only adjustable inlet guide vanes. In Figure 6, graphs 80, 82, 84 and 86 and 88 * correspond to the discrete placement of guide vanes 33 in the increasingly closed positions. It is noted that the closing guide flaps 33, like the closing of the diffuser ring sections 40 and 42, have the effect of decreasing the flow velocity of the point of inrush. Therefore, an inrush condition can often be avoided by adjusting the inlet guide vanes 33 relative to a closed position. However, it is observed, from the diagram of performance of Figure 6 that by adjusting the guide vanes 33 to a closed position, has the effect of decreasing the available head pressure 10 at the point of induction. Therefore, a low flow rate operation condition that requires a pressure Relatively high will not be able to be satisfied by adjusting only the guide fins 33. • * In contrast, it is observed from the performance diagram of Figure 5, that the pressure of the point of inrush, available from the compressor 10, remains essentially stable when the rings of the diffuser 40 and 42 are adjusted to a closed position. Therefore, an operation condition that requires a low flow rate and a high compressor pressure can be satisfied by adjusting the diffuser rings 40 and 42 to a position closed.

An operating condition that requires a low flow velocity and a high pressure ratio with respect to the total load operating pressure ratio (eg 90% total load) is common in the case where there is a large difference ( for example, about 10 ° C or more (50 ° F or more)) between ambient air temperature and indoor temperature, but with occasional light load in a building that is being cooled. In such a situation, a relatively high compressor pressure ratio (eg, above about 2.5) is required by the saturation pressures of the refrigerant corresponding to the condenser, and the evaporation temperatures, but only a reduced flow rate for example, 25% of the full load is needed to remove the heat generated inside the building. Figure 7 shows a performance diagram for a compressor having both adjustable guide vanes and a variable pipe diffuser according to the invention.

It is noted that the efficiency of a compressor can often be optimized by combining an adjustment of the guide vanes 33 with an adjustment of the diffuser rings 40 and 42. With reference to Figure 7, the dotted curves 111, 112, 113, 114 , 115 and 116 show the performance diagrams for a compressor having a variable diffuser in a fully open position for various fittings of the inlet guide fins 33, while the in-dotted non-dotted curves 101, 102, 103, 104 and 105 show the performance diagrams for a compressor having the diffuser rings partially closed (in this case, there is about 40% original flow rate in the closed position) in the placement of several guide fins. As is well known to those skilled in the art, a compressor operates at optimum efficiency when operating at the "angle" (for example 81 in Figure 6) of the graph of performance that characterizes the performance of the compressor. With reference to diagram 7, the operating condition that requires, for example, a pressure of about 0.7 maximum and a flow velocity of about "0.3 maximum, will be satisfied very efficiently by a compressor operating in accordance with the graph 104, made by adjusting the diffuser rings 40 and 42 to a closed position and adjusting the guide vanes 33 to a 10 degree position. Although the present invention has been explained with Referring to a number of specific embodiments, it will be understood that the spirit and scope of the present invention will be determined with reference to the appended claims.

Claims (15)

1. CLAIMS 1. A positioning drive to facilitate the rotary movement of a cylindrical member, the mechanism includes a stationary member, the mechanism of The driving arrangement is characterized in that it comprises: rotating driving means mounted in a fixed manner to the stationary member having a driving end and J? F a differential pinion mounted thereto; a rack gear mounted fixedly to the cylindrical member extending radially outwardly from the cylindrical member and adapted to engage in a gear arrangement with the differential pinion; and a travel limiter for limiting the travel of the cylindrical member between a first position and a second position.
2. 2. In a centrifugal compressor - having a cover and an impeller rotatably mounted thereon for carrying a working fluid from an inlet to the inlet of a radially disposed annular split ring diffuser, the The diffuser includes an inner ring, the inner ring has a plurality of first sections of flow guide channel formed therein, an outer ring, the outer ring has a plurality of second flow guide channels formed therein, - every second channel section of The flow guide has a first section of complementary flow guide channel; the compressor includes an impeller positioning mechanism for rotating the circumferential inner ring within the outer ring between a first fully open position wherein the first and second complementary flow channel sections are aligned to allow a maximum flow of fluid through the Complementary channel sections, and a second ~ ^ partially closed position where the first and second complementary flow guide channels are misaligned In order to restrict the flow of fluid through the complementary channel sections, the positioning drive mechanism is characterized in that it comprises: rotary drive means mounted in a fixed manner to the cover, having a driving end and a pinion 15 differential mounted to them; a rack gear mounted in a fixed manner to the inner ring extending radially outwardly from the inner ring and adapted to engage in the gear arrangement with the differential pinion; and a travel limiter for limiting the travel of the inner ring between a first position corresponding to the fully open position and a second position corresponding to the partially closed position.
3. 3. The positioning drive according to claim 2, characterized in that the rotating drive means is an electric motor.
4. 4. The positioning drive according to claim 2, characterized in that the rotary drive means is an actuator.
5. The positioning drive according to claim 2, characterized in that the inner ring is circumferentially rotatably adjustable to a plurality of positions between fully open and partially closed positions.
6. The positioning drive according to claim 2, characterized in that the trip limiter is composed of switches operatively connected in electrical form to the rotating drive means.
7. The positioning drive according to claim 2, characterized in that the trip limiter is composed of a pair of stop blocks arranged on the outer ring for rigidly coupling the rack gear in a first position corresponding to the fully open position and for rigidly coupling the rack gear in a second position corresponding to the partially closed position.
8. 8. The positioning drive according to claim 2, characterized in that the trip limiter is composed of a cavity disposed inside the outer ring and adapted to receive the rack gear, the cavity comprises: a first end placed for rigidly coupling the rack gear in a first position corresponding to the fully open position; and a second end positioned to engage the rack gear in a second position • corresponding to the partially closed position.
9. The positioning drive mechanism according to claim 2, characterized in that the rack gear having a lower surface includes a circumferential groove disposed in the lower surface and the inner ring includes a circumferential raised portion concentric with the inner ring and adapted to engage the notch of the rack gear and place the rack gear in circumferential alignment with the inner • ring.
10. 10. A kickback adjustment mechanism, in combination with a pivotable drive system having a stationary member, a movable member, a driving shaft having a differential pinion fixedly attached to one end, a rack gear fixedly attached to a movable member, the kickback adjustment mechanism is characterized in that it comprises: the fixed member having a mounting hole disposed therein; a housing having a cylindrical body concentrically positioned around a first central line having a hole axially disposed through the body positioned around a second central line; the driving shaft arranged rotatably concentric with the second center line within the hole of the housing; the movable member movably connected to the stationary member so that the differential pinion meshes with the rack gear to engage them in a pulsable manner; the housing mounted within the mounting hole and rotatably operable to effect an adjustment of the kickback between the differential pinion and the rack gear.
11. The kickback adjustment mechanism according to claim 1, characterized in that the housing includes an area adapted for coupling a wrench.
12. The kickback adjustment mechanism according to claim 1, characterized in that the housing includes clamping means for releasably securing the housing in a selected adjustment position.
13. 13. The kickback adjustment mechanism for use 5 in a centrifugal compressor having a cover, an annular radially disposed split ring diffuser, disposed within the cover, the diffuser includes an outer ring * fixedly attached to the cover, an inner ring rotatably mounted inside the outer ring, a driving shaft 10 having a differential pinion fixedly attached to one end, a rack gear fixedly attached to a movable member, the kickback adjustment mechanism is characterized in that it comprises: the cover having a mounting hole 15 arranged in it; a housing having a cylindrical body positioned concentrically about a first central line and having a hole axially disposed through the body positioned around a second central line; 20 the driving shaft is rotatably arranged concentrically with the second center line within the hole of the housing; the inner ring is movably connected to the cover so that the differential pinion meshes with the 25 rack gear to connect them impulsively; the housing mounted within the mounting hole and rotatably operable to effect a kickback adjustment between the pinion gear and rack gear. The kickback adjustment mechanism according to claim 4, characterized in that the housing includes an area adapted to engage a wrench. The kickback adjustment mechanism according to claim 4, characterized in that the housing includes fastening means for releasably securing the housing in a selected adjustment position.