Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
  • F04C15/00—Component parts, details or accessories of machines, pumps or pumping installations, not provided for in groups F04C2/00 - F04C14/00
  • F04C15/06—Arrangements for admission or discharge of the working fluid, e.g. constructional features of the inlet or outlet
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
  • F04C18/00—Rotary-piston pumps specially adapted for elastic fluids
  • F04C18/30—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members
  • F04C18/34—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F04C18/08 or F04C18/22 and relative reciprocation between the co-operating members
  • F04C18/344—Rotary-piston pumps specially adapted for elastic fluids having the characteristics covered by two or more of groups F04C18/02, F04C18/08, F04C18/22, F04C18/24, F04C18/48, or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in group F04C18/08 or F04C18/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the inner member
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
  • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
  • F01C21/00—Component parts, details or accessories not provided for in groups F01C1/00 - F01C20/00
  • F01C21/08—Rotary pistons
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
  • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
  • F04C2/00—Rotary-piston machines or pumps
  • F04C2/30—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members
  • F04C2/34—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members
  • F04C2/344—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the inner member
  • F04C2/3441—Rotary-piston machines or pumps having the characteristics covered by two or more groups F04C2/02, F04C2/08, F04C2/22, F04C2/24 or having the characteristics covered by one of these groups together with some other type of movement between co-operating members having the movement defined in groups F04C2/08 or F04C2/22 and relative reciprocation between the co-operating members with vanes reciprocating with respect to the inner member the inner and outer member being in contact along one line or continuous surface substantially parallel to the axis of rotation

Definitions

• the invention relates to a vane cell machine for liquids, consisting of a slotted rotor mounted in a stator, in which radially displaceable vanes are slidably arranged, which can be pressed against a stator inner wall by centrifugal, spring or other compressive forces, crescent-shaped expanding or constricting delivery cells are formed and the liquid entry occurs through a hollow central stator and the vane cells are filled from the inside out.
• Vane cell machines are built as constant pumps or motors or as variable pumps or motors. Vane cell machines are also used as volumetric counters. The advantages of vane cell machines are their even flow rate and their low-noise operation. Problems arise from the respective hydraulic radial and axial bearing loads.
• the hydraulic radial bearing loads in vane cell machines with rotor lengths equal to the working range of the vanes result from the product of the projection area, formed from the rotor and the outstanding vane, and the hydraulic pressure, ie the differential pressure acting on the rotor.
• Small radial loads result from the vane friction on the stator and in the Rotor slots and the weight of the rotor.
• the rotor shafts and bearings are either dimensioned to a large extent or an attempt is made to compensate for them by complex and aerodynamically disadvantageous multi-stroke pump or motor designs.
• the hydraulic axial bearing loads can be avoided by symmetrical formation of the axial hydraulic active surfaces of the rotor, the hydraulic pressures exerted on the active surfaces having to be the same.
• the rotor rests on one side in the area of the stator, the hydraulic pressure being more effective on the opposite side, so that there is no axial force compensation. This would be remedied by an axially immovable design of the rotor bearing with a precisely identical setting of the front rotor gaps, which is, however, complex.
• pneumatic cushions are provided in some recesses in the end faces of the rotor according to DE-A 21 33 455, which lie between the impeller blades and are fed with compressed air through channels which are worked into the side covers of the housing in a circular arc, so that when the rotor is axially displaced, pressure differences occur between the pneumatic cushions located on both sides of the rotor, which exert pushing forces in the direction of a central position.
• a comparatively complex solution is also in the DE 31 20 350 proposed for a vane machine, in which the shaft rotor is designed with two large axially displaceable liners which are acted upon in axially displaceable bearing bushes on the rear and front sides by the delivery pressure in pressure-loaded gaps in order to equalize the pressure to bring about on the shaft rotor so that the bearing loads and friction losses are minimized.
• Disadvantages are the large and costly number of precision parts in the hydraulic working area, the relatively large gap lengths required between the high-pressure and low-pressure areas and the resulting poor efficiency of the wing cell machine.
• the shaft emerging for the input or output from a rotary piston machine causes axial bearing loads due to the pressure difference at the shaft seal and, in the case of constant ring seals, also due to the spring force of the same, unless a compensation is provided on the opposite side by a symmetrical design.
• rotary piston pumps are known for example from DE-AS 12 36 641.
• a stator cavity of constant cross section a cylindrical rotating rotor with a plurality of essentially radial slots, in which blades slide, is supported, several conveyor cells being formed between the stator and the rotor by a correspondingly wavy design of the cross-sectional contour of the stator cavity the conveying or working medium is fed in and out via tangential opening channels, of which the suction or low pressure side, each on one side of a wing are sensitive, leading to a central rotor cavity, while the high-pressure side, each on the other side of each vane channels are each connected in a continuous longitudinal channel of the rotor assigned to each vane.
• the longitudinal channels are in turn connected to an annular groove made in a stator side wall, which is connected to the high-pressure side connection opening of the pump or the motor.
• DE-AS 12 36 941 further proposed that as pressure-side channels on the relevant side of each wing a plurality of grooves running in the wing sliding direction are provided, which are incorporated into the corresponding wall of the associated rotor slot, an annular groove being additionally arranged on both sides of the rotor in the side walls of the stator facing the rotor end walls, into which the pressure-side side Open the longitudinal channels of the rotor, the ring grooves being connected to the pressure connections of the pump or motor.
• the rotor cavity in which the low-pressure side bores of the rotor lead, is part of a central longitudinal bore of a shaft connected to the rotor.
• this rotary piston machine is expensive to construct.
• volumetric efficiency and the dry suction capacity (with an empty pump) of vane pumps are determined by the gap losses, the magnitude of which - assuming the same product to be conveyed, the same manufacturing accuracy and pressure difference - depends on the gap lengths. Therefore, with a comparable flow rate, low-speed pumps with a correspondingly large cyclical pump volume and gap lengths have poorer volumetric efficiencies and less dry suction capacity than fast-rotating pumps with a correspondingly smaller cyclical pump volume and gap lengths. These technical relationships mentioned also limit the possibilities for constructively improving the volumetric efficiency and the dry suction capacity due to the necessary speed limitation by the holding pressure.
• rotary piston machines for liquids require large-dimensioned shafts and bearings due to the large projection surface formed from the rotor and protruding vanes, which are subjected to the differential pressure, unless the rotary piston machines are designed as double-stroke vane pumps or motors, each of which have two inlet and outlet openings for the liquids, a measure which is technically complex in terms of production and leads to an increase and thus a deterioration in the holding pressure in pumps.
• the invention consists in that the rotor of the wing cell machine is shaft-free and tubular and is extended on both sides beyond the working area determined by the wing and is mounted with the extensions in the outer stator and has continuous wing slots from the inside to the outside diameter and that the jacket of the stator in the area of the rotor extensions on the surface is acted upon by the operating pressure and / or relieved of hydraulic active surfaces directed against the rotor for at least partial compensation or avoidance of radially occurring forces.
• the operating pressure becomes effective in the bearing gaps rotor / outer stator located there, which leads to further bearing loads.
• recesses (effective surfaces) in the stator casing which are relieved of the operating pressure reduce this radial load part considerably.
• the invention according to claim 2 provides a solution, which is characterized in that the wing cell machine is provided with a hollow central stator, the channels for filling the enlarging wing cells through radial recesses in the wings and / or are formed in the wing slots and the central stator has operating pressure applied to the surface against the rotor has surfaces for at least partial compensation of radial forces, the hydraulically loadable recesses being able to be replaced by small bores which are subjected to operating pressure and which produce larger pressure-acting surfaces directed against the rotor in the rotor / outer stator bearing gaps.
• This measure is simpler to manufacture, causes comparatively lower gap losses and thus improves the volumetric efficiency.
• this vane cell machine is of simple construction, a comparatively complex additional shaft bearing and the resulting friction forces being avoided from the outset, as well as axial and radial hydraulic forces can be minimized.
• the radial channels for filling the feed cells are formed by radial recesses in the vanes and / or in the wing slots, which consistently from the outside diameter to the rotor longitudinal bore as the inside diameter of a bilaterally beyond the working area determined by the vanes ⁇ projecting shaftless rotor, the liquid entering axially through the hollow rotor axis and the filling of the enlarging feed cells in the radial direction through a window in the rotor axis and in the further course through recesses in the rotor slots and / or in the wings.
• the radial filling of the vane cells from the inside via the rotor slots also has the advantage that the stroke volume of the vanes in the rotor slots is included in the cyclical working volume of the pump or motor without a special filling process for this stroke volume against the centrifugal force, as is the case with the tangential or axial filling of the vane cells known from the prior art is required.
• the rotor axis which also serves as a liquid inlet and as a bearing for the rotor, advantageously enables hydraulic and in particular radial pressure compensation by hydraulic support against the rotor axis in pumps and motors.
• the recesses are preferably acted upon by the operating pressure given by the liquid, so that no further pressure sources or controls are necessary.
• the recesses in the stator casing outside the wing working area are arranged opposite the rotor outer casing, that is to say symmetrically with respect to a vertical surface passing through the wing working area.
• the recesses lie in the casing of a stator pin, which defines the central opening of a rotor Pipes through and rests on this sealing.
• the last-mentioned embodiment has the advantage that the recesses can also be at the same height as the wing working area, which may result in a reduction in height. Combinations of the embodiment mentioned are equally possible.
• the rotor part projecting beyond the wing working area has the same or a reduced outer diameter in comparison to a diameter in the wing working area.
• a reduced diameter outside the vane working area has the advantage that the rotor is axially centered when the vane machine is running.
• stator parts that laterally delimit the working space toward the axis of rotation are slightly chamfered to expand the working space.
• bevels are slightly wider on both sides than it corresponds to the axial mobility of the rotor in the stator, so that when the vane machine begins to rotate, the vanes that come out by centrifugal force immediately center the rotor to the work area and, because of the lack of axial forces, this also without additional forces Friction on the wings is maintained.
• the rotor is tubular and has a longitudinal bore in which an even number of vane slots ends openly and in which diametrically opposite vanes are firmly connected to one another or are formed in one piece.
• the rotor can also receive a stator pin in the tube opening, which is hollow on the inside and in the area of the radial through the rotor through slots for the movable wing has a window and the wings and / or the rotor slots have radial recesses. This arrangement enables a partial compensation of the radial hydraulic forces on the rotor.
• the rotor is preferably coupled on one of its end faces to an axially fixed shaft as an input or output connection, the shaft being received in the stator housing.
• the stator bore is made in a pitch circle radially outward over the area going through the maximum radial deflection of the vanes, so that there is a connection between two or more feed cells via the recess created thereby. This measure makes it easier to fill the conveyor lines.
• stator jacket transition area between the enlarging and the shrinking conveyor cells or the guiding of the vanes in the area between two conveyor cells is preferably arranged centrally with respect to the axis of rotation, so that the vanes do not undergo any radial movement in this area which is loaded with differential pressure To run.
• the inner stator outer jacket has depressions which can be acted upon hydraulically by the pump delivery pressure or the inlet pressure of the motor at least partial compensation of the radial hydraulic bearing load.
• the rotor parts projecting beyond the wing working area preferably have a reduced outer diameter in comparison to the rotor diameter in the wing working area. As a result, the rotor is axially centered during operation.
• the stator sheath that laterally delimits the working area of the wing is conical in the area of the non-pressurized blades, so that the blades when Start sliding into the axially centered position.
• the rotor is connected directly or via a coupling on the end face opposite the inlet opening to a shaft as a drive or output device, the shaft being guided in a sealed manner into the stator housing.
• 1 shows a vertical section through a vane cell machine
• 2 shows a vertical section along line II-II in FIG. 1
• FIG 3 shows a partial view of a longitudinal section through a vane cell machine with conically shaped transition areas between the vane working area and the adjacent stator jacket
• FIG. 4 shows a sectional view through a vane machine with a tubular rotor, the diametrically opposite vanes of which are connected to one another,
• FIG. 6 is a sectional view of a vane machine with a rotor which has a central bore in which a stator pin is fitted
• FIG. 8 shows an embodiment in which the central stator is designed for the inlet connection, in a vertical section
• FIG. 10 shows a longitudinal cross section of a further embodiment of a vane machine, and 11 shows a cross section at the height of the wing working area perpendicular to the section according to FIG. 10,
• FIG. 13 shows a vertical section along line XIII-XIII in FIG. 12.
• the vane machine which is preferably designed as a single-stroke vane machine, which is designed as a pump in the embodiment shown in FIGS. 1 and 2, has a shaftless rotor 1, which in the axial direction either has an outer diameter 2 with a constant circumference, as in the vane type beits Society 15, or a reduced circumference 3 outside.
• the rotor 1 is fitted sealed in a stator 4 outside the wing working area.
• the stator has recesses 5, which are designed according to their position and size such that the operating pressure of the liquid acting here leads to a partial or complete hydraulic force compensation, also taking into account the friction and weight forces.
• the recesses 5, viewed in the axial direction are located in front of or behind the wing working area 15 and are arranged symmetrically thereto.
• the vertical end faces or jumps in diameter 6 in an existing in the upper half of FIG. 2 The jump in diameter of the rotor 1 also serves to center the rotor, which results in gaps 7 of equal size between the rotor end face and the opposite stator end face during running operation.
• these jumps in diameter 6 serve to center the rotor to the working space, the disadvantage described above of the one-sidedly greater effectiveness of the hydraulic pressure being able to be accepted by contacting the opposite side, since the End face 6 is kept small as the effective area by a small difference in diameter.
• the gaps 7 between the end faces of the rotor 1 and the stators 4 on both sides are ensured for hydrostatic force compensation at the same pressure.
• the rotor 1 has radially extending slots 8, in which the vanes 9 are slidably guided.
• the space in the guide slots 8 below the blades 9 is in each case connected to the wing cell located in front in the direction of rotation, in the present case by radial recesses 10 in the blade and / or recesses 11 in the rotor. Since, when the vane cell machine is at a standstill, as shown in FIG.
• the vanes 9 which are moved outwards by centrifugal force during operation can be immersed in the rotor, and the freely movable rotor 1, which is not reduced in diameter, is axially displaced on one side against an end face of the stator 4 can be, whereby the wings 9 are prevented from coming out when the vane cell machine starts up, which can lead to the jamming of the vanes on the relevant stator inner wall, are in the area the wing 9 not acted upon by differential pressure, the stator inner jacket parts 12, which laterally delimit the working space 15, are conical or slightly beveled towards the axis of rotation, widening the working space.
• stator inner jacket parts 12 extend slightly further on both sides than corresponds to the axial mobility of the rotor 1 in the stator, so that when the vane cell machine begins to rotate, the vanes which come out by centrifugal force immediately center the rotor 1 to the working space 15 and this is maintained even without additional friction on the wings 9 in the absence of axial forces.
• the input and output connection of the vane cell machine takes place via a shaft 13 protruding into the stator housing 4 and sealed there, which shaft is connected to the rotor axially without reaction via a coupling 14.
• the rotor 1 is of tubular design, a stator pin 16 protruding centrally into the tube opening, the stator pin 16 being firmly connected to the other stator parts.
• the hydraulic operating pressure in the region of the pipe slots is not effective on the rotor.
• the remaining hydraulic and radial forces caused by weight and friction are caused by recesses 17 acted upon by the hydraulic operating pressure on the surface of the central stator pin 16 in pumps in the area of the decreasing vane cells and in motors and volumetric counters recesses 18 in the area of the enlarging ones Vane cells depending on the The size and position of the recesses are partially or completely balanced.
• stator pin 16 which is hollow up to the end of the working space width 20.
• This stator has a window 21 in the working area 15 of the expanding wings 9, radial recesses 10 and 11 being provided in the wings 9 and / or in the rotor 1, through which the expanding wing cells are filled with the assistance of the centrifugal force become.
• the recesses 10 and 11 are arranged in the direction of rotation on the back of the blades and / or in the rotor directly behind the blades.
• the vane machine shown in FIGS. 10 and 11 essentially consists of a rotor 111 which is mounted on a hollow shaft 110 as an inner stator 100 and which is arranged in its stator 112 so as to be rotatable and surrounded by it.
• the stator 112 can be formed in two parts, in particular with a component 113 integrated with the hollow shaft 110.
• the rotor 111 has a reduced diameter to the side thereof and lies with its outer lateral surface in a sealed manner against the inner stator jacket. Between the end faces 114 and 115 of the rotor, a gap 116 and 117 is formed to the opposite end face of the stator, which is pressurized.
• an axial bore 118 and a radial bore 118 ' provide between the gaps 116 and 117.
• the rotor is connected directly or via a coupling (not shown) to a shaft 119, which is rotatably mounted in a sealed manner in the stator housing or in the input or output machine.
• the hollow shaft 110 is designed as an end-face inlet opening which is accessible in the direction of the arrow 120 and which is connected via a window opening 121 of the hollow shaft via corresponding recesses in the rotor to radially extending groove-shaped recesses 122 in the rotor and recesses 123 in the wing.
• the vanes 124 are located in radial slots 125 of the rotor 111.
• the inner stator 100 is provided on its running surface with depressions 126 which are acted upon hydraulically by the pump delivery pressure or the inlet pressure of the motor and are arranged in size and position such that the radial hydraulic Bearing load is partially or fully balanced.
• stator inner jacket has additional recesses 129 which protrude in a crescent shape beyond the maximum radial deflection (curve 128).
• a transmission area 130 is provided, in which the vanes 124 do not execute any radial movement when rotating in the direction of the arrow 131. 10 and 11 works as follows:
• the liquid flowing in in the direction of arrow 120 is guided through the window opening 121 into the groove-shaped radial recesses 122, 123 radially outward into the conveying cells 127 and is discharged essentially tangentially in the direction of arrow 132.
• the ingress of liquid through the hollow axis and the filling of the enlarging vane cells from the inside to the outside takes place largely through the supply of energy from the drive and leads to low holding pressures even at high speeds.
• a hydraulic pressure equalization can be created by simple design measures.
• the tubular rotor 201 is slidably supported in the two bearings 202 and 203.
• the single-stroke cam ring 204 forms the working space 205 and is firmly connected to the bearings 202 and 203.
• This 3-part outer cylindrical stator is inserted into the pump housing 207 with a liquid-carrying or flowable gap 206 and is sealed at both ends to the pump housing, for example by O-rings 208.
• the pressure outlet 209 located in the cam ring acts upon the gap 206 with the respective operating pressure of the pump when it is transferred to the corresponding outlet port 218 of the housing 207.
• one or more radial bores 210 are arranged in the bearings 202 and 203, which have opposite pressure forces acting on the rotor within the bearing areas and lead to partial or full pressure equalization.
• the inner stator 213 is fitted in a contactless manner but with a narrow gap in the inner diameter of the rotor 201.
• the filling of the same takes place via the inlet bore 214 of the inner stator 213 which runs through to the drive side and the window 215 in the region of the expanding vane cells.
• the inlet pressure is effective on both end faces of the rotor.
• the bearings in the circumferential effective area of the hydraulic radial pressure forces are provided with recesses 211 which are connected to the low-pressure side via the gaps 217 and the bores 214 and 218, so that in the region of the recesses 211 only a small storage length 212 remains, which is sufficient for sealing and storage.
• the hydraulic operating pressure acts directly on the inner stator without loading the rotor and, in addition, the gap between the rotor and the inner stator is pressurized via the rotor slots, which contributes to a further partial pressure equalization.

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• Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• General Engineering & Computer Science (AREA)
• Rotary Pumps (AREA)
• Hydraulic Motors (AREA)
• Centrifugal Separators (AREA)
• Soil Working Implements (AREA)

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Publications (1)

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ID=6883305

Family Applications (1)

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Cited By (3)

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Citations (4)

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• 1992
  • 1992-09-02 DE DE9211768U patent/DE9211768U1/de not_active Expired - Lifetime
• 1993
  • 1993-08-26 CA CA002143719A patent/CA2143719C/en not_active Expired - Fee Related
  • 1993-08-26 EP EP93919192A patent/EP0659237B1/de not_active Expired - Lifetime
  • 1993-08-26 JP JP06506841A patent/JP3129737B2/ja not_active Expired - Fee Related
  • 1993-08-26 KR KR1019950700840A patent/KR950703124A/ko not_active IP Right Cessation
  • 1993-08-26 DK DK93919192.0T patent/DK0659237T3/da active
  • 1993-08-26 AU AU49543/93A patent/AU684725B2/en not_active Ceased
  • 1993-08-26 DE DE59302390T patent/DE59302390D1/de not_active Expired - Fee Related
  • 1993-08-26 WO PCT/EP1993/002311 patent/WO1994005912A1/de active IP Right Grant
  • 1993-08-26 AT AT93919192T patent/ATE137306T1/de not_active IP Right Cessation
  • 1993-09-02 CN CN93118992A patent/CN1040786C/zh not_active Expired - Fee Related

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