Classifications

• • 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
  • F04D29/05—Shafts or bearings, or assemblies thereof, specially adapted for elastic fluid pumps
  • F04D29/051—Axial thrust balancing
  • F04D29/0516—Axial thrust balancing balancing pistons
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
  • F04D17/00—Radial-flow pumps, e.g. centrifugal pumps; Helico-centrifugal pumps
  • F04D17/08—Centrifugal pumps
  • F04D17/10—Centrifugal pumps for compressing or evacuating
• • 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
  • F04D27/008—Stop safety or alarm devices, e.g. stop-and-go control; Disposition of check-valves
• • 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
• • 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
  • F04D29/05—Shafts or bearings, or assemblies thereof, specially adapted for elastic fluid pumps
  • F04D29/056—Bearings
• • 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
  • F05D2210/00—Working fluids
  • F05D2210/10—Kind or type
  • F05D2210/12—Kind or type gaseous, i.e. compressible
• • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
  • Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
  • Y10S—TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
  • Y10S415/00—Rotary kinetic fluid motors or pumps
• • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
  • Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
  • Y10S—TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
  • Y10S417/00—Pumps

Definitions

• the present invention relates generally to centrifugal compressors and, more specifically, to balancing thrust in such compressors.
• a compressor is a machine which increases the pressure of a compressible fluid, e.g., a gas, through the use of mechanical energy.
• Compressors are used in a number of different applications, including operating as an initial stage of a gas turbine engine.
• Gas turbine engines in turn, are themselves used in a large number of industrial processes, including power generation, natural gas Hquification and other processes.
• centrifugal compressors in which the mechanical energy operates on gas input to the compressor by way of centrifugal acceleration which accelerates the gas particles, e.g., by rotating a centrifugal impeller or rotor through which the gas passes.
• Centrifugal compressors can be fitted with a single rotor, i.e., a single stage configuration, or with a plurality of rotors in series, in which case they are frequently referred to as multistage compressors.
• Each of the stages of a centrifugal compressor typically includes an inlet conduit for gas to be compressed, a rotor which is capable of providing kinetic energy to the input gas and an exit pipe which converts the kinetic energy of the gas leaving the rotor into pressure energy.
• Multistage centrifugal compressors are subjected to an axial thrust on the rotor caused by the differential pressure across the stages and the change of momentum of the gas turning from the horizontal to the vertical direction.
• This axial thrust is normally compensated by a balance piston and an axial thrust bearing. Since the axial thrust bearing cannot be loaded by the entire thrust of the rotor, a balance piston is designed to compensate for most of the thrust, leaving the bearing to handle any remaining, residual thrust.
• the balance piston is normally implemented as a rotating disc or drum which is fitted onto the compressor shaft, such that each side of the balance disc or drum is subjected to different pressures during operation.
• the diameter of the balance piston is chosen to have a desired axial load to avoid its residual load from overloading the axial bearing.
• Conventional oil-lubricated bearings are typically designed to withstand axial thrust forces on the order of four times the maximum residual axial thrust which are expected to occur during abnormal, e.g., surging, conditions.
• the compensation provided by the balance piston may not be sufficient to avoid bearing overload.
• some types of centrifugal compressors are more likely than others to experience such gas condition variations, e.g., in gas storage applications for multistage centrifugal compressors employing parallel operation, wherein the difference in axial thrust between the first and second sections of the compressor, linked to flow coefficient difference, may not be readily compensated for by the balance piston.
• conventional oil-lubricated bearings are typically designed to withstand axial thrust forces on the order of four times the maximum residual axial thrust which are expected to occur during abnormal, e.g., surging, conditions.
• AMBs active magnetic bearings
• AMBs operate by based on electromagnetic principles to control axial and radial displacements within the compressor.
• AMBs include an electromagnet driven by a power amplifier which regulates the voltage (and therefore the current) into the coils of the electromagnet as a function of a feedback signal which indicates displacement of the compressor's rotor inside the device.
• AMBs have the desirable attribute that they do not require oil as a lubricant, reducing overall maintenance of the compressor system and, potentially, removing the requirement to provide seals between the impellers and the bearing.
• AMBs also have the drawback that they are not capable of handling as much axial thrust as the conventional oil-lubricated bearings.
• Exemplary embodiments relate to systems and methods for dynamically balancing axial loads in centrifugal compressors to reduce residual axial loads on the bearings used therein.
• a sensor or probe detects a parameter associated with the axial load acting on the bearing. Based on the detected parameter, the pressure in a balance chamber is controlled to adjust the compensating axial force generated by a balance drum.
• Advantages according to exemplary embodiments described herein include, for example, a reduction in the residual axial forces acting on the bearings across different operating conditions.
• advantages are not to be construed as limitations of the present invention except to the extent that they are explicitly recited in one or more of the appended claims.
• a centrifugal compressor includes a rotor assembly including at least one impeller, a bearing connected to, and for rotatably supporting, the rotor assembly, a stator, a balance drum disposed between said at least one impeller and said bearing, a balance chamber, defined at least in part by an outboard side of said balance drum, and having a balance line connected thereto, a sensor for sensing a parameter which is associated with an axial load on said bearing, and a control valve for varying a pressure within the balance chamber based on the sensed parameter.
• a method for dynamically balancing axial load acting on a bearing in a centrifugal compressor includes the steps of detecting a parameter associated with the axial load, and controlling a pressure in a balance chamber proximate a balance drum in the centrifugal compressor based on the detected parameter to dynamically balance the axial load acting on the bearing.
• Figure 1 is a schematic view of a multistage-type centrifugal compressor, which can be provided with dynamic balancing mechanisms according to exemplary embodiments;
• Figure 2 depicts static axial load balancing in a centrifugal compressor
• Figure 3 depicts dynamic axial load balancing in a centrifugal compressor according to an exemplary embodiment
• Figure 4 is a flowchart illustrating a method for dynamic load balancing according to an exemplary embodiment.
• FIG. 1 schematically illustrates a multistage, centrifugal compressor 10 in which such thrust balancing systems may be employed.
• the compressor 10 includes a box or housing (stator) 12 within which is mounted a rotating compressor shaft 14 that is provided with a plurality of centrifugal impellers 16.
• the rotor assembly 18 includes the shaft 14 and impellers 16 and is supported radially and axially through bearings 20 which are disposed on either side of the rotor assembly 18.
• the multistage centrifugal compressor operates to take an input process gas from duct inlet 22, to increase the process gas' pressure through operation of the rotor assembly 18, and to subsequently expel the process gas through outlet duct 24 at an output pressure which is higher than its input pressure.
• the process gas may, for example, be any one of carbon dioxide, hydrogen sulfide, butane, methane, ethane, propane, liquefied natural gas, or a combination thereof.
• sealing systems 26 are provided to prevent the process gas from flowing to the bearings 20.
• the housing 12 is configured to cover both the bearings 20 and the sealing systems 26, to prevent the escape of gas from the centrifugal compressor 10.
• the bearings 20 may be implemented as either oil-lubricated bearings or active magnetic bearings. If active magnetic bearings are used as bearings 20, then the sealing mechanisms 26 may be omitted.
• the centrifugal compressor 10 also includes the afore-described balance piston (drum) 28 along with its corresponding labyrinth seal 30.
• a balance line 32 maintains the pressure in a balance chamber 34 on the outboard side of the balance drum at the same (or substantially the same) pressure as that of the process gas entering via the inlet duct 22.
• this balance line 32 includes a control valve which can modulate the pressure in the balance chamber 34 based upon, for example, sensed axial loading on or near the bearing 20 as will be described below with respect to Figure 3.
• the balancing drum 28 is designed to exert an axial force in the outboard direction, the magnitude of which is based on the expected axial load of the impellers minus that of the motor. This is accomplished by, for example, designing the system such that the pressure Pu of the process gas on the inboard side of the balancing drum 28 is greater than the pressure Pe on the outboard side of the balancing drum 28, and by selecting a balancing drum of an appropriate size (diameter) to generate the desired balancing force.
• the pressure imbalance is developed and maintained by providing the balance line 32 between the balance chamber 34 and the main suction line associated with inlet duct 22 such that the pressure in the balance chamber is substantially the same as that on the inboard side of the impellers 16.
• the axial thrust compensation provided by the balance drum 28 would substantially offset the axial load placed on the bearings 20 by the impellers 16, or at least offset such an axial load enough that any residual load is within the design specifications of the bearings 20.
• operational variances within such compressors and/or the use of AMBs as bearings 20 may cause the residual loading to exceed the design tolerances of the bearings 20 for axial loading.
• Table 1 which illustrates results from an axial loading test for an exemplary six impeller centrifugal compressor 10 having a balance drum 28 with a diameter of 231 mm rotating at 17000 rpm. This test compressor was equipped with AMBs as the bearings 20 which were nominally rated for axial loading of between +/- 9000 N.
• a control valve 40 is placed on the balance line 32 to enable an automated control of the pressure Pe which is exerted on the outboard side of the balancing drum 28 as shown in Figure 3.
• the same reference numerals as used in Figures I and 2 refer to the same or similar components of a centrifugal compressor 10.
• the control valve 40 regulates the pressure in the balance chamber 34 to vary the reaction force generated by the balancing drum 28 as a function of, for example, the bearing 20' s displacement or the axial load on the bearing 20 as measured by a sensor or probe 42.
• the control valve 40 thus controls the value of the pressure Pe and, accordingly, the amount of compensatory axial load provided by the balancing drum 28. More specifically, by closing the control valve 40, the pressure Pe increases thereby reducing the amount of compensatory axial load provided by the balancing drum 28. Alternatively, by opening the control valve 40, the pressure Pe decreases thereby increasing the amount of axial load provided by the balancing drum 28. When the control valve 40 is completely open, the maximum amount of compensatory axial load is generated by the balancing drum 28.
• the amount of load provided by the balancing drum 28 is, according to exemplary embodiments, controllably variable, it may be desirable to design the balancing drum 28 such that its maximum compensatory axial load is larger than that provided by conventional static balancing drums (i.e., by providing a larger balancing drum 28 to the system) since it is possible in these exemplary embodiments to reduce the amount of compensation being provided by closing the control valve 40 as desired.
• control valve 40 is controlled based on a feedback signal from probe or sensor 42 regarding the amount of axial loading that the bearing 20 is experiencing at a given time. Measurements can be made periodically by the probe or sensor 42 and reported back to control logic 44 which is connected to control valve 40 to implement any desired control algorithm to open and close the valve 40 as needed to adjust for operating changes which result in more (or less) residual loading of the bearings 20.
• control logic 44 may be implanted as an ASIC, FPGA, computer, or other type of processor and may be implemented purely in hardware, purely in software or in some combination thereof.
• the sensor or probe 42 may be any of a number of different types.
• an induction sensor or probe such as a linear potentiometric displacement transducer (LPDT) can be used to measure displacement of the bearing 20 due to axial loading.
• LPDT linear potentiometric displacement transducer
• an eddy current sensor or probe may be a more appropriate implementation of sensor or probe 42.
• Other types of sensors e.g., piezoelectric or sensors which measure pressure at the oil film in the bearing, may be used as alternatives.
• control logic 44 can include a proportional integral derivative (PID) controller which automatically, in a closed loop, changes the pressure in the balancing drum chamber 34 as a function of the measured thrust on the machine.
• PID proportional integral derivative
• the control logic 44 can act on the valve 40 via a simple PID controller.
• the control system can be designed with a bias (hysteresis value) to avoid any thrust valve hunting.
• a test was performed to evaluate the arrangement according to exemplary embodiments illustrated in Figure 3 and to determine its ability to better control the residual load on the bearings 20.
• the test used the same type of centrifugal compressor 10 which was evaluated above to generate the results in Table 1, i.e., a six impeller centrifugal compressor running at 17000 rpm, except that the balance drum 28 was increased in size to have a diameter of 247 mm to provide for a slightly greater maximum compensatory axial load capability in this dynamic balancing arrangement.
• the results of the test are shown below in Table 2.
• exemplary embodiments enable for centrifugal compressors to be outfitted with smaller thrust bearings, since the axial loading on such bearings can be better controlled.
• such compressors are expected to have higher availability by reducing the residual load on such bearings.
• a method for controlling residual axial loading in such compressor systems according to exemplary embodiments can be performed as illustrated in the flowchart of Figure 4. Therein, at step 100, a parameter associated with the axial load on the bearing is detected. Then, at step 102, a pressure in the balance chamber proximate the balance drum in the centrifugal compressor is controlled based on the detected parameter to dynamically balance the axial load acting on the bearing.

Landscapes

• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Structures Of Non-Positive Displacement Pumps (AREA)
• Control Of Positive-Displacement Air Blowers (AREA)
• Testing Of Balance (AREA)

Priority Applications (10)

Applications Claiming Priority (2)

Publications (2)

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Cited By (4)

Families Citing this family (9)

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• 2009
  • 2009-12-22 IT ITCO2009A000072A patent/IT1397707B1/it active
• 2010
  • 2010-12-16 KR KR1020127019136A patent/KR20120123351A/ko not_active Application Discontinuation
  • 2010-12-16 CA CA2785334A patent/CA2785334A1/en not_active Abandoned
  • 2010-12-16 JP JP2012545248A patent/JP5928827B2/ja not_active Expired - Fee Related
  • 2010-12-16 CN CN201080064545.9A patent/CN102762871B/zh not_active Expired - Fee Related
  • 2010-12-16 EP EP10793252A patent/EP2516866A2/en not_active Withdrawn
  • 2010-12-16 US US13/518,574 patent/US20130115042A1/en not_active Abandoned
  • 2010-12-16 MX MX2012007457A patent/MX2012007457A/es not_active Application Discontinuation
  • 2010-12-16 AU AU2010335267A patent/AU2010335267A1/en not_active Abandoned
  • 2010-12-16 RU RU2012127256/06A patent/RU2557143C2/ru not_active IP Right Cessation
  • 2010-12-16 BR BR112012015363A patent/BR112012015363A2/pt not_active IP Right Cessation
  • 2010-12-16 WO PCT/EP2010/070001 patent/WO2011076668A2/en active Application Filing

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