EP1321679B1 - Drive position mechanism with backlash adjustment for variable pipe diffuser - Google Patents
Drive position mechanism with backlash adjustment for variable pipe diffuser Download PDFInfo
- Publication number
- EP1321679B1 EP1321679B1 EP03004927A EP03004927A EP1321679B1 EP 1321679 B1 EP1321679 B1 EP 1321679B1 EP 03004927 A EP03004927 A EP 03004927A EP 03004927 A EP03004927 A EP 03004927A EP 1321679 B1 EP1321679 B1 EP 1321679B1
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- EP
- European Patent Office
- Prior art keywords
- compressor
- diffuser
- backlash
- housing
- adjustment mechanism
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired - Lifetime
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Classifications
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D27/00—Control, e.g. regulation, of pumps, pumping installations or pumping systems specially adapted for elastic fluids
- F04D27/02—Surge control
- F04D27/0246—Surge control by varying geometry within the pumps, e.g. by adjusting vanes
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D29/00—Details, component parts, or accessories
- F04D29/40—Casings; Connections of working fluid
- F04D29/42—Casings; Connections of working fluid for radial or helico-centrifugal pumps
- F04D29/44—Fluid-guiding means, e.g. diffusers
- F04D29/46—Fluid-guiding means, e.g. diffusers adjustable
- F04D29/462—Fluid-guiding means, e.g. diffusers adjustable especially adapted for elastic fluid pumps
- F04D29/464—Fluid-guiding means, e.g. diffusers adjustable especially adapted for elastic fluid pumps adjusting flow cross-section, otherwise than by using adjustable stator blades
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D27/00—Control, e.g. regulation, of pumps, pumping installations or pumping systems specially adapted for elastic fluids
- F04D27/002—Control, e.g. regulation, of pumps, pumping installations or pumping systems specially adapted for elastic fluids by varying geometry within the pumps, e.g. by adjusting vanes
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- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
- F05D—INDEXING SCHEME FOR ASPECTS RELATING TO NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES, GAS-TURBINES OR JET-PROPULSION PLANTS
- F05D2250/00—Geometry
- F05D2250/50—Inlet or outlet
- F05D2250/52—Outlet
Definitions
- the present invention relates to centrifugal compressors in general and in particular to a backlash adjustment mechanism for a variable pipe diffuser for a centrifugal compressor.
- DE-A-3529281 discloses a device for changing the flow direction of an airflow entering the suction-side opening of a compressor.
- US2,382,913 describes a compressor having a diffuser with a plurality of diffuser vanes with a fixed portion and a moveable portion.
- the entire impeller wheel configuration is varied in response to load changes in an effort to match the machine performance with the changing load demands.
- Adjustable diffuser flow restrictors are also described in U.S. Pat. No. 3,362,625 which serve to regulate the flow within the diffuser in an effort to improve stability at low volumetric flow rates.
- variable width diffuser in conjunction with fixed diffuser guide vanes.
- variable width vaned diffusers wherein the diffuser vanes are securely affixed, as by bolting to one of the diffuser walls.
- the vanes are adapted to pass through openings formed in the other wall thus permitting the geometry of the diffuser to be changed in response to changing load conditions.
- variable geometry pipe diffuser (which may also be termed a split-ring pipe diffuser) splits the diffuser into a first, inner ring and a second outer ring.
- the inner and outer rings have complementary inlet flow channel sections formed therein. 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 outer ring are rotatable respective one another. The rings are rotated to improve efficiency for varying pressure levels between a fully open position and a partially closed position. In the partially closed position the misalignment of the exit pipes of the diffuser causes an increase in noise. Rotation of the rings past an optimum design point results in excessive noise and efficiency degradation.
- split ring diffusers Another problem with split ring diffusers is premature part wear. Lubricants are generally not used within the gas flow regions of centrifugal compressors to preclude contamination of the gases. The dynamic loads imposed upon the split ring diffuser by the gas flow exiting the impeller cause wear in the components of the diffuser to be accelerated by the absence of lubricating oil.
- the drive system for accurately positioning the rings relative to one another must, among other things, be rigid to avoid any fretting of components. Because of circumferential loading on the rings there is a propensity for the inner ring to oscillate relative to the outer ring which could cause compressor instability, part wear and could adversely affect efficiency. This causes several problems that need to be overcome.
- a drive system is needed that is capable of preventing the relative movement between the inner and outer rings.
- a bearing concept is also needed which would allow for the relative rotation of the two rings and also be capable of withstanding the circumferential and thrust loads while maintaining tight geometric tolerances between the rings. There is also a need to provide a positioning system that includes positive minimum and maximum stops to avoid unnecessary noise and efficiency degradation as well as simple field retrofit. In addition, there is a need for the drive and bearing systems have a long operating life and be easy to install and adjust properly.
- compressor comprising a backlash adjustment mechanism for use in a variable geometry pipe diffuser for a centrifugal compressor, as claimed in claim 1.
- a rack gear is attached to the inner ring of a variable pipe diffuser in a preferred embodiment of the present invention.
- a pinion gear is attached to a rotation drive means mounted to the casing of a centrifugal compressor in meshing arrangement with the rack gear.
- the rotational drive means is operated to rotate the inner ring relative to the outer ring between a position where the diffuser pipes are fully open and a position where the diffuser pipes are partially closed.
- the drive means is capable of aligning the pipes of the diffuser to a plurality of positions between the fully open and partially closed positions.
- a travel limiter is also provided to positively limit the travel of the inner ring at the fully open and the partially closed positions.
- the backlash adjustment mechanism has a housing concentric about a first centerline and a bore therethrough concentric about a second centerline.
- the pinion gear is mounted to a drive shaft passing through the bore.
- the housing is rotatably mounted to the casing.
- the housing of the present invention is rotatably operable about the second centerline to adjust the backlash between the rack gear and pinion gear.
- a centrifugal compressor 10 as part of an HVAC system (not shown) having an impeller 12 for accelerating refrigerant vapor to a high velocity, a diffuser 14 for decelerating the refrigerant to a low velocity while converting kinetic energy to pressure energy, and a discharge plenum in the form of a collector 16 to collect the discharge vapor for subsequent flow to a condenser.
- Power to the impeller 12 is provided by an electric motor (not shown) which is hermetically sealed in the other end of the compressor and which operates to rotate a high speed shaft 19.
- the refrigerant enters the inlet opening 29 of the suction housing 31, passes through the blade ring assembly 32 and the guide vanes 33, and then enters the compression suction area 23 which leads to the compression area defined on its inner side by the impeller 12 and on its outer side by the housing 34. After compression, the refrigerant then flows into the diffuser 14, the collector 16 and the discharge line (not shown).
- a variable geometry pipe diffuser 14 which can incorporate a backlash adjustment mechanism 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.
- the inner and outer rings have complementary flow channel sections 44 and 46 formed therein. That is, each flow channel section 44 of the inner ring 40 has a complementary channel section 46 formed in outer ring 42.
- Inner ring 40 and outer ring 42 are rotatable with respect to one another. In a preferred embodiment, inner ring 40 rotates circumferentially within a stationary outer ring 42.
- Rings 40 and 42 are adjustable between a first fully open position, as illustrated in Fig. 3, wherein complementary channel sections are aligned and a maximum amount of fluid passes through inner and outer rings 40 and 42, and a second, partially closed position, as illustrated in Fig. 4, wherein complementary channels are misaligned and flow through the channel sections 44 and 46 is restricted.
- a ring support mechanism 35 is shown The mechanism shown illustrated the use of three such mechanisms spaced circumferentially equidistant about the diffuser.
- the ring support mechanism includes an inner bearing slot 41 and a cutout 43 disposed in inner ring 40, a roller assembly 54, a roller axle assembly 36 and an outer bearing slot 45 disposed in the outer ring.
- the roller assembly as shown in Fig. 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 Figs. 6 - 7 includes an axle 37 and an axle bolt 39.
- axle 37 includes a hex head 38 and an axle body 47, an axle body centerline 48, an axle bore 49 and an axle bore centerline 50.
- the axle 39 includes a pair of shoulders 73, 74 concentric with axle bore centerline 50.
- split ring diffusers Another problem with split ring diffusers is premature part wear.
- Lubricants are generally not used within the gas flow regions of centrifugal compressors to preclude contamination of the gases.
- the dynamic loads imposed upon the split ring diffuser by the gas flow exiting the impeller cause wear in the components of the diffuser to be accelerated by the absence of lubricating oil. Due to the non-availability of lubricating oils in most compressors it is usually necessary to take steps to minimize friction and fretting wear. Accordingly, in certain embodiments of the present invention and as described herein below, component interfaces are hard coated, parts are manufactured from ultra high molecular weight plastic materials, the ring assemblies are preloaded and backlash is eliminated from the gears of the positioning drive system.
- the outer ring 42 is stationary with respect to the suction housing and three sets of ring support mechanisms 35 are installed into the outer ring by positioning the roller assembly 54 within the bearing slot 45 of the outer ring, passing the axle through the mounting hole 58 and the roller assembly and then installing the axle bolt 39 through the axle and loosely threading the axle bolt 39 into threaded holes 59 in the outer ring.
- the inner ring 40 is installed inside of the outer ring with the cutouts 43 of the inner ring circumferentially aligned with the bearing slot 45 and the roller assemblies 35 and then rotating the inner ring clockwise as shown in Fig. 7 to position the roller assemblies within the bearing slot 41.
- the ring support mechanisms are employed to properly center and position the inner ring by rotating the axle through the use of a wrench placed on the hex head 38.
- the axle body centreline 48, on which the roller 55 is mounted is offset from axle bore centerline 50, which is concentric with the shoulders 73, 74, by 0.5mm (.021 inches).
- the rotation of hex head 38 causes the roller assembly to rotate about the shoulders within the outer ring and causes the roller assembly to be radially displaced relative to the outer ring.
- the hex head is further rotated to preload the outer bearing surface 56 of the roller assemblies against the inner ring.
- the axle bolt 39 is then tightened.
- the preload conditioned is preferred because it prevents the inner ring from movement due to tangential and circumferential loads.
- the roller 55 and the inner ring 40 are aluminum and both the outer bearing surface 56 and the inner bearing slot 41 are hardened to prevent wear.
- the roller assemblies restrain movement of the rings in the axial direction due to thrust loads by positioning the thrust bearing surfaces 57 within the hardened inner bearing slot 41 and the relatively soft outer bearing slot 45.
- the thrust bearing surface 57 of the roller assembly must allow for the rotation of the inner and outer rings and at the same time withstand the thrust loads produced by the compressor.
- the thrust bearing surface 57 is manufactured from ultra high molecular weight plastic 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 slot and are utilized to carry the thrust loads of the compressor and to adjust axial tolerances of the inner ring.
- FIG. 16 A mechanism for limiting and precluding axial movement of the inner ring relative to the outer ring is shown in Fig. 16.
- an axial restraint system 90 comprising a threaded shaft 91, a threaded mounting hole 92, a bearing pad 93, a lock nut 94, a hex head 95, and a recess 96.
- the axial restraint mechanism 90 is installed such that the bearing pad 93 is positioned in the recess 96.
- the bearing pad positioned within the recess allows clearance for the shroud 34 to be mounted to outer ring 42 without accidental contact of the bearing pads with the inner ring..
- the threaded shaft 91 is rotated to bring the bearing pad in contact with the inner ring.
- the mechanism is releasably fastened by tightening lock nut 94.
- the bearing pad is manufactured from an ultra high molecular weight plastic material.
- a compressor may include six such axial restraint mechanisms positioned circumferentially equally spaced about the inner ring.
- a positioning drive mechanism 121 for rotating inner ring 40 circumferentially within outer ring 42 is described with reference Fig. 10.
- Outer ring 42 has fixedly attached thereto rack gear 123 which extends radially outwardly from outer ring 42.
- rack gear 123 In gearing relation with rack gear 123 is pinion gear 124 which is driven via pinion axle 126 by actuator 128.
- Actuator 128 is selected and controlled to effect movement of inner ring 40 in relation to outer ring 42 between a first fully open position and a second partially closed position and any number of intermediate positions therebetween.
- Axle 126 is housed in a containment housing 130 which hermetically seals axle 126 from compressor interior 132 and which prevents leakage of fluid out of compressor 10 through containment housing 130.
- the tangential and circumferential loading on the rings by the refrigerant flow within the diffuser causes the inner ring to have the propensity to chatter back and forth within the outer ring. Excess movement or chattering of the inner ring would cause the rack gear 123 and the pinion gear 124 to fret and also cause other parts to wear. Preloading the inner ring via the roller assemblies as discussed herein earlier prevents movement of the inner ring as well as chattering under normal operating conditions. In cases of abnormal conditions, such as operating in a surge, a secondary mechanism is needed to prevent motion of the inner ring.
- the present invention provides for a drive mounting system to prohibit adverse movement and chattering of the inner ring by preventing the backlash between the rack gear and the pinion gear via adjustment of the relative center positions of the pinion gear and the rack gear utilizing the axle containment housing 130.
- the axle housing outer surface 125 is concentric about housing centerline 127 and housing bore 129 is concentric about housing bore centreline 131.
- the housing centerline 127 and the housing bore centerline 129 are offset by 1.5mm (.060 inches).
- wrench flats 135 and adjustment slots 134 of the positioning drive mechanism Referring to Fig. 11 there is shown wrench flats 135 and adjustment slots 134 of the positioning drive mechanism.
- the backlash between the rack gear 123 and the pinion gear 124 is removed by rotating the drive positioning mechanism by placing a wrench (not shown) across wrench flats 135. Once minimal backlash is achieved the positioning drive mechanism is fixed in place by the tightening of cap screws 133. Once the backlash is eliminated the tendency for the inner ring to move is discharged directly by the actuator through the gear system.
- the flow of fluid through diffuser 14 in a second partially closed position in relation to the fully open position flow rate is determined by the ratio of the minimum cross-sectional area of a flow channel of a diffuser in a partially closed position to the minimum cross-sectional area of a flow channel (defined by complementary channel sections 44 and 46) in a fully open position.
- This minimum flow channel area known as the "throat area” will generally be determined by the smallest diameter of the flow passage 52 of the inner ring channel 44 when 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 diffuser 14 is in a second partially closed position. For example, if a diffuser channel has a minimum area (throat area) of 81mm 2 (1/8 sq.
- the volumetric flow rate of fluid through a diffuser in the partially closed position will be about 50% of the flow rate as in the fully open position.
- the flow rate of fluid through compressor 10 when diffuser 14 is in a second, partially closed position will generally be between about 10% and 100% of the flow rate of fluid through compressor 10 when diffuser is in the first fully open position.
- a second partially closed position (Fig. 4) at least about 10% the volume of flow as in the fully open position should flow through diffuser 14 so as to prevent excessive thermodynamic heating, excessive noise and a degradation in the efficiency of the compressor.
- the amount of relative rotation between the two ring sections should be limited to an amount of rotation necessary to effect a second partially closed position.
- the rings should not be adjustable to completely close off a flow of fluid therebetween. The degree of allowable 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 inlet flow channel sections 44, 46 in the ring sections 40 and 42 in relation to the volume of the ring sections 40 and 42.
- R 2 defines the radius of the impeller tip
- R 3 defines the outside radius of inner ring 40
- R 4 defines the outside radius of outer ring.
- a diffuser having a mechanism to provide positive positioning of the inner ring corresponding to a first fully open position and a second partially closed position.
- Cavity 137 is machined in outer ring 42 to accommodate rack gear 123.
- Rack gear 123 is accurately mounted to inner ring 40 in a tongue and groove fashion wherein the rack gear is provided with a circumferential groove 143 adapted to receive tongue section 139 of inner ring 40.
- To determine the fully opened position the inner ring is positioned within the outer ring and the rings are rotated relative to one another until flow passages 52 are fully aligned with outer flow channels 46.
- the rack gear is mounted to the inner ring with gear face 145 in contact with full open stop 140 of cavity 137.
- Bolts (not shown) are then installed through gear mounting holes 142 and securely and tightened into threaded holes 138 in the inner ring.
- the rack gear and the cavity are sized to provide for a predetermined amount of closure of the pipe diffuser. For example they may be sized such that difference between the rack gear angular width and the cavity provide for a 10% open position. In this example the required travel of the rack gear is 10 degrees, the rack gear angular width is 35 degrees and the corresponding cavity angular width is 45 degrees.
- a positive stop is created between the rack gear and the cavity to accurately and repeatably position the rings at points corresponding to a fully open position and a partially closed position.
- the positive stops also allow for field retrofit of actuator 128 without the need to adjust the position of the inner and outer rings.
- FIG. 5 showing a performance diagram for a compressor having a variable pipe diffuser integrated therein.
- the performance diagram of Fig. 5 includes a plurality of performance plots 60, 62, 64, 66 and 68, each corresponding do a discreet positioning between inner and outer ring sections 40 and 42.
- Each performance plot, e.g. 60 is characterized by a surge point, e.g. 70, which is the point of maximum available pressure. Operating a compressor at a flow rate at or below the surge point will likely result in a surge condition, as discussed in the Background of the Invention section herein.
- plot 60 may correspond, for example, to a first, fully open position
- plot 62 may correspond to an intermediate 2 degree partially closed position
- plot 64 may correspond to an intermediate 4 degree partially closed position
- plot 68 may correspond to a maximum 8 degree partially closed position.
- adjusting ring sections 40 and 42 toward a closed position has the effect of adjusting the surge point e.g. 70, 72 in a performance plot for a compressor toward a lower flow rate.
- a surge condition can be avoided during periods of low flow demand by adjusting diffuser rings 40 and 42 toward a closed position.
- surge point pressure available from compressor 10 remains essentially stable when diffuser rings 40 and 42 are adjusted toward a closed position.
- an operating condition requiring a low flow rate and high compressor pressure can be satisfied by adjusting diffuser rings 40 and 42 toward a closed position.
- An operating condition requiring a low flow rate and a high pressure ratio relative to the full load operating pressure ratio (e.g. 90% of full load) is common in the case where there is a large difference (e.g. about 50° F or more) between the ambient air temperature and indoor temperature, but occasional light loading in a building being cooled.
- a relatively high compressor pressure ratio (e.g. above about 2.5) is required by the refrigerant saturation pressures corresponding to the condenser, and evaporation temperatures, but only a reduced flow rate e.g. 25% of full load is needed to remove the heat generated within the building.
- Fig. 7 shows a performance diagram for a compressor having both adjustable guide vanes and a variable pipe diffuser. It is seen that efficiency of a compressor can often be optimized by combining an adjustment of guide vanes 33 with an adjustment of diffuser rings 40 and 42.
- dash curves 111, 112, 113, 114, 115, and 116 show performance plots for a compressor having a variable diffuser in a fully open position for various positioning of inlet guide vanes 33
- solid curves 101, 102, 103, 104 and 105 show performance plots for a compressor having partially closed (here, there is about 40% of original flow rate in the closed position) diffuser rings at various guide vane positioning.
- a compressor operates at optimum efficiency when operating at the "knee" (e.g. 81 at Fig. 6) of the performance plot characterizing performance of the compressor.
- the operating condition requiring, for example, a pressure of about 0.7 maximum, and a flow rate of about 0.3 maximum would be most efficiently satisfied by a compressor operating in accordance with plot 104, realized by adjusting diffuser rings 40 and 42 to a closed position and by adjusting guide vanes 33 to a 10 degree position.
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Description
- The present invention relates to centrifugal compressors in general and in particular to a backlash adjustment mechanism for a variable pipe diffuser for a centrifugal compressor.
- One of the major problems arising in the use of centrifugal vapor compressors for applications where the compressor load varies over a wide range is flow stabilization through the compressor. The compressor inlet, impeller and diffuser passages must be sized to provide for the maximum volumetric flow rate desired. When there is a low volumetric flow rate through such a compressor, the flow becomes unstable. As the volumetric flow rate is decreased from a stable range, a range of slightly unstable flow is entered. In this range, there appears to be a partial reversal of flow in the diffuser passage, creating noises and lowering the compressor efficiency. Below this range, the compressor enters what is known as surge, wherein there are periodic complete flow reversals in the diffuser passage, destroying the efficiency of the machine and endangering the integrity of the machine elements. Since 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 machine efficiencies over a wide operation range. DE-A-3529281 discloses a device for changing the flow direction of an airflow entering the suction-side opening of a compressor. US2,382,913 describes a compressor having a diffuser with a plurality of diffuser vanes with a fixed portion and a moveable portion. In U.S. Pat. No. 4,070,123, the entire impeller wheel configuration is varied in response to load changes in an effort to match the machine performance with the changing load demands. Adjustable diffuser flow restrictors are also described in U.S. Pat. No. 3,362,625 which serve to regulate the flow within the diffuser in an effort to improve stability at low volumetric flow rates.
- A common technique for maintaining high operating efficiency over a wide flow range in a centrifugal machine is through use of the variable width diffuser in conjunction with fixed diffuser guide vanes.
- U.S. Pat. Nos. 2,996,996 and 4,378,194, issued to a common assignee, describe variable width vaned diffusers wherein the diffuser vanes are securely affixed, as by bolting to one of the diffuser walls. The vanes are adapted to pass through openings formed in the other wall thus permitting the geometry of the diffuser to be changed in response to changing load conditions.
- Fixedly mounting the diffuser blades to one of the diffuser walls presents a number of problems particularly in regard to the manufacture, maintenance and operation of the machine. Little space is afforded for securing the vanes in the assembly. Any misalignment of the vanes will cause the vane to bind or rub against the opposite wall as it is repositioned. Similarly, if one or more vanes in the series has to be replaced in the assembly, the entire machine generally has to be taken apart in order to effect the replacement.
- The efficiency of a compressor could be greatly enhanced by varying the outlet geometry of the diffuser. In U.S. Patent number US5,807,071, commonly assigned, a variable geometry pipe diffuser is disclosed. A variable geometry pipe diffuser (which may also be termed a split-ring pipe diffuser) splits the diffuser into a first, inner ring and a second outer ring. The inner and outer rings have complementary inlet flow channel sections formed therein. 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 outer ring are rotatable respective one another. The rings are rotated to improve efficiency for varying pressure levels between a fully open position and a partially closed position. In the partially closed position the misalignment of the exit pipes of the diffuser causes an increase in noise. Rotation of the rings past an optimum design point results in excessive noise and efficiency degradation.
- The geometrical tolerances within a centrifugal compressor are small. At the same time the loads within the compressor are large and dynamic in nature. In a split ring pipe diffuser the problem of maintaining tolerances in the face of the dynamic loading becomes quite onerous. There are both axial (thrust) loads and circumferential loads on the ring pair that need to be managed. The diffuser rings must be able to rotate relative to one another and at the same time tight control over their relative position must be maintained in order to ensure proper alignment of the flow channels and the ultimate efficiency of the compressor. The cost of maintaining the necessary tolerances in a split ring diffuser is generally very high.
- Another problem with split ring diffusers is premature part wear. Lubricants are generally not used within the gas flow regions of centrifugal compressors to preclude contamination of the gases. The dynamic loads imposed upon the split ring diffuser by the gas flow exiting the impeller cause wear in the components of the diffuser to be accelerated by the absence of lubricating oil.
- The drive system for accurately positioning the rings relative to one another must, among other things, be rigid to avoid any fretting of components. Because of circumferential loading on the rings there is a propensity for the inner ring to oscillate relative to the outer ring which could cause compressor instability, part wear and could adversely affect efficiency. This causes several problems that need to be overcome. A drive system is needed that is capable of preventing the relative movement between the inner and outer rings. A bearing concept is also needed which would allow for the relative rotation of the two rings and also be capable of withstanding the circumferential and thrust loads while maintaining tight geometric tolerances between the rings. There is also a need to provide a positioning system that includes positive minimum and maximum stops to avoid unnecessary noise and efficiency degradation as well as simple field retrofit. In addition, there is a need for the drive and bearing systems have a long operating life and be easy to install and adjust properly.
- In accordance with the present invention there is provided compressor comprising a backlash adjustment mechanism for use in a variable geometry pipe diffuser for a centrifugal compressor, as claimed in claim 1.
- A rack gear is attached to the inner ring of a variable pipe diffuser in a preferred embodiment of the present invention. In a preferred embodiment, a pinion gear is attached to a rotation drive means mounted to the casing of a centrifugal compressor in meshing arrangement with the rack gear. In a preferred embodiment, the rotational drive means is operated to rotate the inner ring relative to the outer ring between a position where the diffuser pipes are fully open and a position where the diffuser pipes are partially closed. In addition, in a preferred embodiment, the drive means is capable of aligning the pipes of the diffuser to a plurality of positions between the fully open and partially closed positions. In a preferred embodiment, a travel limiter is also provided to positively limit the travel of the inner ring at the fully open and the partially closed positions. The backlash adjustment mechanism has a housing concentric about a first centerline and a bore therethrough concentric about a second centerline. The pinion gear is mounted to a drive shaft passing through the bore. The housing is rotatably mounted to the casing. The housing of the present invention is rotatably operable about the second centerline to adjust the backlash between the rack gear and pinion gear.
- In the drawings, wherein like numerals are used to indicate the same elements throughout the views;
- Fig. 1 is a cross-section side view of a compressor incorporating a backlash mechanism according to the invention having a variable pipe diffuser;
- Fig. 2 is a perspective view of a variable pipe diffuser incorporating a backlash mechanism according to the invention;
- Figs. 3 and 4 are cross-sectional front views of a variable pipe diffuser incorporating a backlash mechanism in accordance with the invention in a first, fully open, and a second, partially closed position, respectively;
- Fig. 5 is a top view of a compressor having a variable diffuser incorporating a backlash mechanism of the present invention;
- Fig. 6 is a cross-section view of a ring support mechanism incorporating a backlash mechanism of the present invention taken substantially along line 6-6 in Fig. 5;
- Fig. 7 is a cross-section view of a ring support mechanism incorporating a backlash mechanism of the present invention taken substantially along line 7-7 in Fig. 6;
- Fig. 8 is a cross-section of a roller assembly incorporating a backlash mechanism of the present invention;
- Fig. 9 is a cross-section view of an axle incorporating a backlash mechanism of the present invention;
- Fig. 10 is a cross-sectional view of a positioning drive mechanism incorporating a
backlash mechanism of the present invention of
detail area 10 in Fig. 1; - Fig. 11 is a top view of a positioning drive mechanism incorporating a backlash mechanism of the present invention;
- Fig. 12 is a perspective view of a rack gear incorporating a backlash mechanism of the present invention.
- Fig.13 is a perfomance diagram for a variable pipe diffuser incorporating a backlash mechanism according to the present invention;
- Fig. 14 is a performance diagram for a compressor having inlet guid vanes only;
- Fig. 15 is a performance diagram for a compressor incorporating a backlash mechanism according to the present invention having a variable pipe diffuser and inlet guide vanes; and
- Fig. 16 is a cross-sectional view of a compressor having an axial restraint mechanism and incorporating a backlash mechanism according to the present invention.
-
- Referring now to Fig 1, there is shown a
centrifugal compressor 10 as part of an HVAC system (not shown) having animpeller 12 for accelerating refrigerant vapor to a high velocity, adiffuser 14 for decelerating the refrigerant to a low velocity while converting kinetic energy to pressure energy, and a discharge plenum in the form of acollector 16 to collect the discharge vapor for subsequent flow to a condenser. Power to theimpeller 12 is provided by an electric motor (not shown) which is hermetically sealed in the other end of the compressor and which operates to rotate ahigh speed shaft 19. - Referring now to the manner in which the refrigerant flow occurs in the
compressor 10, the refrigerant enters the inlet opening 29 of thesuction housing 31, passes through theblade ring assembly 32 and theguide vanes 33, and then enters thecompression suction area 23 which leads to the compression area defined on its inner side by theimpeller 12 and on its outer side by thehousing 34. After compression, the refrigerant then flows into thediffuser 14, thecollector 16 and the discharge line (not shown). - A variable
geometry pipe diffuser 14 which can incorporate a backlash adjustment mechanism according to the present invention includes a first,inner ring 40 and a secondouter ring 42, aring support mechanism 35, and apositioning drive mechanism 121. Referring to Figs. 3 and 4 the inner and outer rings have complementaryflow channel sections flow channel section 44 of theinner ring 40 has acomplementary channel section 46 formed inouter ring 42.Inner ring 40 andouter ring 42 are rotatable with respect to one another. In a preferred embodiment,inner ring 40 rotates circumferentially within a stationaryouter ring 42. - When one ring is rotated with respect to the other, the alignment between each pair of complementary inlet flow channels of the inner and outer rings changes as seen with reference to Figs. 3 and 4.
Rings outer rings channel sections - In Fig. 5 a
ring support mechanism 35 is shown The mechanism shown illustrated the use of three such mechanisms spaced circumferentially equidistant about the diffuser. Referring now to Figs. 6 and 7 the ring support mechanism includes aninner bearing slot 41 and acutout 43 disposed ininner ring 40, aroller assembly 54, aroller axle assembly 36 and anouter bearing slot 45 disposed in the outer ring. The roller assembly as shown in Fig. 8 includes aroller 55 having anouter bearing surface 56, and a pair of thrust bearing surfaces 57. The axle assembly as shown in Figs. 6 - 7 includes anaxle 37 and anaxle bolt 39. As seen in Fig. 9axle 37 includes ahex head 38 and anaxle body 47, anaxle body centerline 48, an axle bore 49 and anaxle bore centerline 50. In addition theaxle 39 includes a pair ofshoulders axle bore centerline 50. - Another problem with split ring diffusers is premature part wear. Lubricants are generally not used within the gas flow regions of centrifugal compressors to preclude contamination of the gases. The dynamic loads imposed upon the split ring diffuser by the gas flow exiting the impeller cause wear in the components of the diffuser to be accelerated by the absence of lubricating oil. Due to the non-availability of lubricating oils in most compressors it is usually necessary to take steps to minimize friction and fretting wear. Accordingly, in certain embodiments of the present invention and as described herein below, component interfaces are hard coated, parts are manufactured from ultra high molecular weight plastic materials, the ring assemblies are preloaded and backlash is eliminated from the gears of the positioning drive system.
- Referring now to the manner in which the inner ring is assembled and its movement. The
outer ring 42 is stationary with respect to the suction housing and three sets ofring support mechanisms 35 are installed into the outer ring by positioning theroller assembly 54 within the bearingslot 45 of the outer ring, passing the axle through the mounting hole 58 and the roller assembly and then installing theaxle bolt 39 through the axle and loosely threading theaxle bolt 39 into threaded holes 59 in the outer ring. Theinner ring 40 is installed inside of the outer ring with thecutouts 43 of the inner ring circumferentially aligned with the bearingslot 45 and theroller assemblies 35 and then rotating the inner ring clockwise as shown in Fig. 7 to position the roller assemblies within the bearingslot 41. With the inner ring installed within the outer ring the ring support mechanisms are employed to properly center and position the inner ring by rotating the axle through the use of a wrench placed on thehex head 38. Theaxle body centreline 48, on which theroller 55 is mounted is offset fromaxle bore centerline 50, which is concentric with theshoulders hex head 38 causes the roller assembly to rotate about the shoulders within the outer ring and causes the roller assembly to be radially displaced relative to the outer ring. Once the inner ring is properly centered within the outer ring the hex head is further rotated to preload theouter bearing surface 56 of the roller assemblies against the inner ring. Theaxle bolt 39 is then tightened. The preload conditioned is preferred because it prevents the inner ring from movement due to tangential and circumferential loads. In an embodiment of the present invention theroller 55 and theinner ring 40 are aluminum and both theouter bearing surface 56 and theinner bearing slot 41 are hardened to prevent wear. The roller assemblies restrain movement of the rings in the axial direction due to thrust loads by positioning the thrust bearing surfaces 57 within the hardenedinner bearing slot 41 and the relatively softouter bearing slot 45. Thethrust bearing surface 57 of the roller assembly must allow for the rotation of the inner and outer rings and at the same time withstand the thrust loads produced by the compressor. In a preferred embodiment thethrust bearing surface 57 is manufactured from ultra high molecular weight plastic 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 slot and are utilized to carry the thrust loads of the compressor and to adjust 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 pre-assemble the inner and outer rings and transport them to the compressor for finally assembly. - A mechanism for limiting and precluding axial movement of the inner ring relative to the outer ring is shown in Fig. 16. There is shown an
axial restraint system 90 comprising a threadedshaft 91, a threaded mountinghole 92, abearing pad 93, alock nut 94, ahex head 95, and a recess 96. During assembly of the diffuser theaxial restraint mechanism 90 is installed such that thebearing pad 93 is positioned in the recess 96. The bearing pad positioned within the recess allows clearance for theshroud 34 to be mounted toouter ring 42 without accidental contact of the bearing pads with the inner ring.. Oncehousing 34 is installed the threadedshaft 91 is rotated to bring the bearing pad in contact with the inner ring. With the bearing pad properly positioned the mechanism is releasably fastened by tighteninglock nut 94. In a preferred mechanism the bearing pad is manufactured from an ultra high molecular weight plastic material. A compressor may include six such axial restraint mechanisms positioned circumferentially equally spaced about the inner ring. - A
positioning drive mechanism 121 for rotatinginner ring 40 circumferentially withinouter ring 42 is described with reference Fig. 10.Outer ring 42 has fixedly attached thereto rackgear 123 which extends radially outwardly fromouter ring 42. In gearing relation withrack gear 123 ispinion gear 124 which is driven viapinion axle 126 byactuator 128.Actuator 128 is selected and controlled to effect movement ofinner ring 40 in relation toouter ring 42 between a first fully open position and a second partially closed position and any number of intermediate positions therebetween.Axle 126 is housed in acontainment housing 130 which hermetically sealsaxle 126 fromcompressor interior 132 and which prevents leakage of fluid out ofcompressor 10 throughcontainment housing 130. The tangential and circumferential loading on the rings by the refrigerant flow within the diffuser causes the inner ring to have the propensity to chatter back and forth within the outer ring. Excess movement or chattering of the inner ring would cause therack gear 123 and thepinion gear 124 to fret and also cause other parts to wear. Preloading the inner ring via the roller assemblies as discussed herein earlier prevents movement of the inner ring as well as chattering under normal operating conditions. In cases of abnormal conditions, such as operating in a surge, a secondary mechanism is needed to prevent motion of the inner ring. The present invention provides for a drive mounting system to prohibit adverse movement and chattering of the inner ring by preventing the backlash between the rack gear and the pinion gear via adjustment of the relative center positions of the pinion gear and the rack gear utilizing theaxle containment housing 130. The axle housing outer surface 125 is concentric about housing centerline 127 and housing bore 129 is concentric about housing bore centreline 131. In an embodiment of the present invention the housing centerline 127 and the housing bore centerline 129 are offset by 1.5mm (.060 inches). Referring to Fig. 11 there is shownwrench flats 135 andadjustment slots 134 of the positioning drive mechanism. After installation of the positioning drive mechanism into thesuction housing 31 the backlash between therack gear 123 and thepinion gear 124 is removed by rotating the drive positioning mechanism by placing a wrench (not shown) acrosswrench flats 135. Once minimal backlash is achieved the positioning drive mechanism is fixed in place by the tightening of cap screws 133. Once the backlash is eliminated the tendency for the inner ring to move is discharged directly by the actuator through the gear system. - The flow of fluid through
diffuser 14 in a second partially closed position in relation to the fully open position flow rate is determined by the ratio of the minimum cross-sectional area of a flow channel of a diffuser in a partially closed position to the minimum cross-sectional area of a flow channel (defined bycomplementary channel sections 44 and 46) in a fully open position. This minimum flow channel area, known as the "throat area" will generally be determined by the smallest diameter of theflow passage 52 of theinner ring channel 44 whendiffuser 14 is in a fully open position, and will be controlled by thewidth 53 at the interface between the inner andouter rings diffuser 14 is in a second partially closed position. For example, if a diffuser channel has a minimum area (throat area) of 81mm2 (1/8 sq. in) in a second partially closed position, and a minimum area (throat area) of 161mm2 (1/4 sq. in.) in a fully open position then the volumetric flow rate of fluid through a diffuser in the partially closed position will be about 50% of the flow rate as in the fully open position. The flow rate of fluid throughcompressor 10 whendiffuser 14 is in a second, partially closed position, will generally be between about 10% and 100% of the flow rate of fluid throughcompressor 10 when diffuser is in the first fully open position. - In a second partially closed position (Fig. 4), at least about 10% the volume of flow as in the fully open position should flow through
diffuser 14 so as to prevent excessive thermodynamic heating, excessive noise and a degradation in the efficiency of the compressor. To this end, the amount of relative rotation between the two ring sections should be limited to an amount of rotation necessary to effect a second partially closed position. In other words, the rings should not be adjustable to completely close off a flow of fluid therebetween. The degree of allowable 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 inletflow channel sections ring sections ring sections - Continuing with reference to Fig. 4, R2 defines the radius of the impeller tip, R3 defines the outside radius of
inner ring 40, and R4 defines the outside radius of outer ring. By making the thickness, defined by the Quantity T = R3-R2 ofinner ring 40 no larger than is necessary to block a desired portion (e.g. 50% of flow) of flow throughouter ring channels 46, the flow of fluid throughdiffuser 14 can be efficiently controlled. Rotation of the inner ring with respect to the outer ring will reduce the diffuser throat area before any diffusion has taken place, thus preventing flow acceleration after diffusion. Also, the smaller the inner ring thickness, T, the smaller the turning angles of the flow through diffuser in the partially closed position. Both of the above-described effects tend to improve compressor efficiency under part-load operating conditions. - Referring now to Figs. 5 and 12 a diffuser is shown having a mechanism to provide positive positioning of the inner ring corresponding to a first fully open position and a second partially closed position.
Cavity 137 is machined inouter ring 42 to accommodaterack gear 123.Rack gear 123 is accurately mounted toinner ring 40 in a tongue and groove fashion wherein the rack gear is provided with acircumferential groove 143 adapted to receivetongue section 139 ofinner ring 40. To determine the fully opened position the inner ring is positioned within the outer ring and the rings are rotated relative to one another untilflow passages 52 are fully aligned withouter flow channels 46. With the rings in this position, and the ring support mechanism adjusted as described herein above, the rack gear is mounted to the inner ring withgear face 145 in contact with fullopen stop 140 ofcavity 137. Bolts (not shown) are then installed throughgear mounting holes 142 and securely and tightened into threadedholes 138 in the inner ring. The rack gear and the cavity are sized to provide for a predetermined amount of closure of the pipe diffuser. For example they may be sized such that difference between the rack gear angular width and the cavity provide for a 10% open position. In this example the required travel of the rack gear is 10 degrees, the rack gear angular width is 35 degrees and the corresponding cavity angular width is 45 degrees. With the rack gear thusly positioned a positive stop is created between the rack gear and the cavity to accurately and repeatably position the rings at points corresponding to a fully open position and a partially closed position. The positive stops also allow for field retrofit ofactuator 128 without the need to adjust the position of the inner and outer rings. - Operation and use of the compressor as described above can be understood with reference to Fig. 5 showing a performance diagram for a compressor having a variable pipe diffuser integrated therein. The performance diagram of Fig. 5, includes a plurality of performance plots 60, 62, 64, 66 and 68, each corresponding do a discreet positioning between inner and
outer ring sections - For purposes of illustration,
plot 60 may correspond, for example, to a first, fully open position,plot 62 may correspond to an intermediate 2 degree partially closed position,plot 64 may correspond to an intermediate 4 degree partially closed position, andplot 68 may correspond to a maximum 8 degree partially closed position. - It is seen that adjusting
ring sections - It is helpful to understanding to compare performance diagram of Fig. 5, for a compressor having a variable diffuser to the performance diagram shown in Fig. 6 corresponding to a compressor having adjustable inlet guide vanes only. In Fig. 6, plots 80, 82, 84, and 86 and 88 correspond to discreet positioning of
guide vanes 33 in increasingly closed positions. It is seen that closingguide vanes 33, like the closing ofdiffuser ring sections inlet guide vanes 33 toward a closed position. - However, it is seen from the performance diagram of Fig. 6 that adjusting
guide vanes 33 toward a closed position has the further effect of lowering the head pressure available fromcompressor 10 at the surge point. Hence, a low flow rate operating condition requiring a relatively high pressure cannot be satisfied by adjustingguide vanes 33 alone. - By contrast, it is seen from the performance diagram of Fig. 5 that surge point pressure available from
compressor 10 remains essentially stable when diffuser rings 40 and 42 are adjusted toward a closed position. Hence an operating condition requiring a low flow rate and high compressor pressure can be satisfied by adjusting diffuser rings 40 and 42 toward a closed position. - An operating condition requiring a low flow rate and a high pressure ratio relative to the full load operating pressure ratio (e.g. 90% of full load) is common in the case where there is a large difference (e.g. about 50° F or more) between the ambient air temperature and indoor temperature, but occasional light loading in a building being cooled. In such a situation, a relatively high compressor pressure ratio (e.g. above about 2.5) is required by the refrigerant saturation pressures corresponding to the condenser, and evaporation temperatures, but only a reduced flow rate e.g. 25% of full load is needed to remove the heat generated within the building.
- Fig. 7 shows a performance diagram for a compressor having both adjustable guide vanes and a variable pipe diffuser. It is seen that efficiency of a compressor can often be optimized by combining an adjustment of
guide vanes 33 with an adjustment of diffuser rings 40 and 42. With reference to Fig. 7 dash curves 111, 112, 113, 114, 115, and 116 show performance plots for a compressor having a variable diffuser in a fully open position for various positioning ofinlet guide vanes 33, whilesolid curves plot 104, realized by adjusting diffuser rings 40 and 42 to a closed position and by adjustingguide vanes 33 to a 10 degree position. - While the present invention has been explained with reference to a number of specific embodiments, it will be understood that the scope of the present invention should be determined with reference to the appended claims.
Claims (8)
- A centrifugal compressor comprising a backlash adjustment mechanism having an annular radially disposed split ring diffuser, the diffuser including an outer ring (42), an inner ring (40) rotatably mounted within the outer ring, a rack gear (123) fixedly attached to said inner ring and a pinion gear (124) operably engaging the rack gear and being attached to a drive shaft (126) for rotation thereof the backlash adjustment mechanism comprising:a housing (130) having a cylindrical body positioned concentrically about a first centerline (127) and having a bore (129) disposed axially through the body positioned about a second centerline (131) offset from said first centerline;the drive shaft (126) rotatably disposed within the bore of the housing; andthe housing (130) mounted so as to be selectively rotatable on said first centerline (127) so as effect an adjustment of the backlash between the pinion gear (124) and the rack gear (123).
- A compressor comprising a backlash adjustment mechanism as claimed in claim 1 wherein the outer ring is a stationary member (42), the inner ring is a movable member (40),
and the drive shaft is rotatably disposed concentric with the second centerline (131). - The compressor comprising a backlash adjustment mechanism set forth in claim 1 or 2, wherein the housing includes an area adapted to engage a wrench for rotation thereof.
- The compressor comprising a backlash adjustment mechanism set forth in any of claims 1 to 3, wherein the housing includes a fastening means for releasably securing the housing at a selected adjustment position.
- The compressor comprising a backlash adjustment mechanism set forth in claim 4 wherein the fastening means comprises an annular collar disposed on said housing and having annular slots formed therein to accommodate fasteners passing through said slots and into an adjacent support structure.
- The compressor comprising a backlash adjustment mechanism set forth in claim 4 or 5 wherein the housing can be rotated over a continuous range of positions so as to provide for precise position adjustments.
- A compressor comprising a backlash adjustment mechanism as claimed in any one of claims 2 to 6
wherein the stationary member (42) has a mounting hole disposed therein; and
the housing (130) is mounted within the mounting hole and rotatably operable to effect the adjustment of the backlash between the pinion gear (124) and the rack gear (123). - A compressor comprising a backlash adjustment mechanism as claimed in any preceding claim further having a casing (34), wherein the annular radially disposed split ring diffuser is disposed within the casing (34), the outer ring (42) is fixedly attached to the casing (34),
the casing (34) has a mounting hole disposed therein;
the inner ring (40) is movably connected to the casing (34) such that the pinion gear (124) meshes with the rack gear (123) to drivably connect them; and
the housing (130) is mounted within the mounting hole and rotatably operable to effect the adjustment of the backlash between the pinion gear (124) and the rack gear (123).
Applications Claiming Priority (5)
Application Number | Priority Date | Filing Date | Title |
---|---|---|---|
US08/907,288 US5988977A (en) | 1997-08-06 | 1997-08-06 | Backlash adjustment mechanism for variable pipe diffuser |
US907288 | 1997-08-06 | ||
US08/907,319 US5895204A (en) | 1997-08-06 | 1997-08-06 | Drive positioning mechanism for a variable pipe diffuser |
US907319 | 1997-08-06 | ||
EP98630040A EP0896157B1 (en) | 1997-08-06 | 1998-07-31 | Drive positioning mechanism with backlash adjustment for variable pipe diffuser |
Related Parent Applications (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
EP98630040A Division EP0896157B1 (en) | 1997-08-06 | 1998-07-31 | Drive positioning mechanism with backlash adjustment for variable pipe diffuser |
Publications (2)
Publication Number | Publication Date |
---|---|
EP1321679A1 EP1321679A1 (en) | 2003-06-25 |
EP1321679B1 true EP1321679B1 (en) | 2004-10-06 |
Family
ID=27129487
Family Applications (2)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
EP98630040A Expired - Lifetime EP0896157B1 (en) | 1997-08-06 | 1998-07-31 | Drive positioning mechanism with backlash adjustment for variable pipe diffuser |
EP03004927A Expired - Lifetime EP1321679B1 (en) | 1997-08-06 | 1998-07-31 | Drive position mechanism with backlash adjustment for variable pipe diffuser |
Family Applications Before (1)
Application Number | Title | Priority Date | Filing Date |
---|---|---|---|
EP98630040A Expired - Lifetime EP0896157B1 (en) | 1997-08-06 | 1998-07-31 | Drive positioning mechanism with backlash adjustment for variable pipe diffuser |
Country Status (10)
Country | Link |
---|---|
EP (2) | EP0896157B1 (en) |
JP (1) | JP3091179B2 (en) |
KR (1) | KR100297572B1 (en) |
CN (1) | CN1279300C (en) |
AU (1) | AU737988B2 (en) |
BR (1) | BR9803641A (en) |
DE (2) | DE69818589T2 (en) |
ES (2) | ES2225802T3 (en) |
MY (2) | MY123718A (en) |
TW (1) | TW402666B (en) |
Cited By (2)
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US9157446B2 (en) | 2013-01-31 | 2015-10-13 | Danfoss A/S | Centrifugal compressor with extended operating range |
US10962016B2 (en) | 2016-02-04 | 2021-03-30 | Danfoss A/S | Active surge control in centrifugal compressors using microjet injection |
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JPH0683160U (en) * | 1993-04-30 | 1994-11-29 | 志 祥 侯 | Solder wire feeder |
JP2003106289A (en) * | 2001-10-01 | 2003-04-09 | Kondo Kogyo Kk | Multi-stage centrifugal fan |
US7356999B2 (en) | 2003-10-10 | 2008-04-15 | York International Corporation | System and method for stability control in a centrifugal compressor |
US7905102B2 (en) | 2003-10-10 | 2011-03-15 | Johnson Controls Technology Company | Control system |
FR2878912B1 (en) * | 2004-12-07 | 2008-08-15 | Peugeot Citroen Automobiles Sa | POWER SUPPLY ASSEMBLY FOR INTERNAL COMBUSTION ENGINE AND CORRESPONDING VEHICLE |
JP4952465B2 (en) * | 2007-09-13 | 2012-06-13 | 株式会社Ihi | Pipe diffuser centrifugal compressor |
TWI452208B (en) * | 2007-10-31 | 2014-09-11 | Johnson Controls Tech Co | Method of controlling the capacity of gas compression system |
US9353765B2 (en) | 2008-02-20 | 2016-05-31 | Trane International Inc. | Centrifugal compressor assembly and method |
US8052384B2 (en) * | 2008-06-17 | 2011-11-08 | Hamilton Sundstrand Corporation | Centrifugal pump with segmented diffuser |
EP2505849A1 (en) * | 2011-03-28 | 2012-10-03 | Siemens Aktiengesellschaft | Method and system for energy optimization of a centrifugal compressor |
CN103158024B (en) * | 2011-12-19 | 2015-04-22 | 沈阳理工大学 | Mechanism capable of adjusting transmission gap of gear rack |
CN104421209B (en) * | 2013-08-26 | 2017-02-08 | 珠海格力电器股份有限公司 | regulator structure and centrifugal compressor |
US10690148B2 (en) | 2015-07-22 | 2020-06-23 | Carrier Corporation | Diffuser restriction ring |
EP3688314A2 (en) | 2017-09-25 | 2020-08-05 | Johnson Controls Technology Company | Two piece split scroll for centrifugal compressor |
KR20210014450A (en) * | 2019-07-30 | 2021-02-09 | 현대자동차주식회사 | Varialble geometry turbocharger |
CN112780612A (en) * | 2021-01-29 | 2021-05-11 | 安徽应流航空科技有限公司 | Diffuser positioning and fixing structure |
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DE2402506C3 (en) | 1974-01-19 | 1979-01-18 | Daimler-Benz Ag, 7000 Stuttgart | Ball joint |
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DE3529281C1 (en) * | 1985-08-16 | 1986-09-11 | Daimler-Benz Ag, 7000 Stuttgart | Device for changing the direction of an air flow entering through the inlet of a compressor of an exhaust gas turbocharger of an internal combustion engine |
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-
1998
- 1998-07-14 TW TW087111412A patent/TW402666B/en not_active IP Right Cessation
- 1998-07-31 EP EP98630040A patent/EP0896157B1/en not_active Expired - Lifetime
- 1998-07-31 ES ES03004927T patent/ES2225802T3/en not_active Expired - Lifetime
- 1998-07-31 DE DE69818589T patent/DE69818589T2/en not_active Expired - Fee Related
- 1998-07-31 EP EP03004927A patent/EP1321679B1/en not_active Expired - Lifetime
- 1998-07-31 ES ES98630040T patent/ES2203914T3/en not_active Expired - Lifetime
- 1998-07-31 DE DE69826901T patent/DE69826901T2/en not_active Expired - Fee Related
- 1998-07-31 CN CNB2004100022331A patent/CN1279300C/en not_active Expired - Fee Related
- 1998-08-05 AU AU78775/98A patent/AU737988B2/en not_active Ceased
- 1998-08-05 KR KR1019980031864A patent/KR100297572B1/en not_active IP Right Cessation
- 1998-08-05 JP JP10220933A patent/JP3091179B2/en not_active Expired - Fee Related
- 1998-08-05 BR BR9803641-6A patent/BR9803641A/en not_active IP Right Cessation
- 1998-08-05 MY MYPI20042528A patent/MY123718A/en unknown
- 1998-08-05 MY MYPI98003568A patent/MY121001A/en unknown
Cited By (3)
Publication number | Priority date | Publication date | Assignee | Title |
---|---|---|---|---|
US9157446B2 (en) | 2013-01-31 | 2015-10-13 | Danfoss A/S | Centrifugal compressor with extended operating range |
US10184481B2 (en) | 2013-01-31 | 2019-01-22 | Danfoss A/S | Centrifugal compressor with extended operating range |
US10962016B2 (en) | 2016-02-04 | 2021-03-30 | Danfoss A/S | Active surge control in centrifugal compressors using microjet injection |
Also Published As
Publication number | Publication date |
---|---|
MY123718A (en) | 2006-05-31 |
DE69826901T2 (en) | 2006-03-09 |
AU737988B2 (en) | 2001-09-06 |
MY121001A (en) | 2005-12-30 |
CN1538072A (en) | 2004-10-20 |
KR19990023377A (en) | 1999-03-25 |
KR100297572B1 (en) | 2001-10-27 |
DE69818589T2 (en) | 2004-08-05 |
EP0896157A1 (en) | 1999-02-10 |
CN1279300C (en) | 2006-10-11 |
DE69826901D1 (en) | 2004-11-11 |
AU7877598A (en) | 1999-02-18 |
DE69818589D1 (en) | 2003-11-06 |
ES2225802T3 (en) | 2005-03-16 |
BR9803641A (en) | 1999-09-28 |
JPH11132197A (en) | 1999-05-18 |
TW402666B (en) | 2000-08-21 |
EP0896157B1 (en) | 2003-10-01 |
EP1321679A1 (en) | 2003-06-25 |
ES2203914T3 (en) | 2004-04-16 |
JP3091179B2 (en) | 2000-09-25 |
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