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
  • F04C2/00—Rotary-piston machines or pumps
  • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
  • F04C2/12—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type
  • F04C2/14—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
  • F04C2/16—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
• • 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
  • F04C2/02—Rotary-piston machines or pumps of arcuate-engagement type, i.e. with circular translatory movement of co-operating members, each member having the same number of teeth or tooth-equivalents
• • 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
  • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
  • F04C2/082—Details specially related to intermeshing engagement type machines or pumps
  • F04C2/084—Toothed wheels

Definitions

• This invention relates to a hydraulic apparatus having intermeshing, external helical gears.
• the hydraulic apparatus could be a gear pump, producing pressurised fluid from a rotary drive, or a gear motor, producing rotary motion from pressurised fluid.
• Helical gears have been proposed for gear pumps used in transfer applications.
• An example of a transfer application is one in which the fluid is transferred from a source to a site at which it is used.
• Helical gear pumps of this type tend to be quieter than spur gear pumps because fluid is expelled from the spaces between the teeth in a gradual, uniform manner, so that the pressure ripple is reduced.
• a primary desideratum of such helical gear pumps is to transfer a large volume of fluid in a small envelope size. This makes it desirable to maximise the ratio between the outside diameter of the gear (the tip circle) and the pitch circle, and to use small numbers of teeth.
• the teeth themselves are shaped to be a close fit in the spaces between teeth, to reduce undesired back- flow from the outlet side to the inlet side through the intermeshing teeth.
• Such pumps may be required to deliver pressures of up to 320 bar (32 x 10 6 Pa) , seldom less then 100 bar (10 x 10 6 Pa) , and tend to use low viscosity hydraulic fluids whose viscosity frequently does not exceed 20 centistokes (2 x 10 "5 m 2 s " ') at normal operating temperatures.
• gear pumps for high specification use, in particular for power applications, are gear pumps which use spur gears. These provide good sealing against back-flow and are employed for high specification use, in spite of their noise.
• gear motors Similar considerations apply in relation to gear motors.
• An additional consideration is that of starting torque.
• a gear motor is supplied with fluid under pressure and thereby is induced to rotate.
• spur gear motor the instantaneous variations of displacement which occur throughout the meshing cycle produce a flow ripple exactly as in a pump, inducing noise, and in addition creating variations in output torque.
• the torque variations are not of great significance; the noise generated by the flow ripple is.
• the torque delivery can be problematical. Whereas a pump cannot generate pressure when stationary, a motor may be required to produce full output torque when stalled and supplied with pressurised fluid.
• Spur gear motors may suffer from start-up difficulties in this regard.
• hydraulic apparatus having two intermeshing gears having external helical teeth, mounted for rotation within a casing between an inlet side and an outlet side, wherein substantially at the instant at which continuous driving contact is achieved across the entire width of a pair of co-operating teeth, the leading end of the preceding tooth comes out of its engagement with its co-operating tooth.
• the gear teeth preferably have flanks of involute profile.
• the geometry is substantially such that the continuous driving contact extends diagonally across the tooth flanks from the involute generation circle to the leading end.
• the contact ratio in the transverse plane is in the range 1.5-3, preferably 1.85-2.2. Most preferably the contact ratio in the transverse plane is in the range 1.95-1.99.
• the helical advance is in the range 0.5-2, preferably 0.85- 1.2, most preferably 1.
• the contact ratio in the transverse plane exceeds the helical advance by 0.8-1.2, preferably by about 1. It can be 1 exactly but it is found preferable that it is slightly less than 1, most preferably 0.95-0.99.
• underlap can be desirable in aiding the release of fluid trapped in the diminishing volume of a tooth space which is receiving a tooth, thus preventing, or helping to prevent, excessive deleterious pressure being generated by entrapment of fluid.
• the underlap is such as to permit release of trapped fluid, but not sufficiently long-lasting to permit back-flow of fluid from the outlet side to the inlet side within the speed range for which the apparatus is designed.
• the hydrodynamic properties of a hydraulic fluid, even one of very low viscosity, are such that very small amounts of underlap do not give rise to any substantial back-flow of hydraulic fluid, from the outlet side to the inlet side. Back-flow would obviously be deleterious to the volumetric efficiency of a gear pump but also is believed to have an adverse effect on the amplitude of the pressure ripple, and hence on noise generation, of a gear pump or gear motor.
• the continuous driving contact is achieved across the entire width of a pair of intermeshing teeth very slightly after the leading end of the preceding tooth comes out of engagement with its co ⁇ operating tooth. This means that there are never quite two continuous lines of driving contact across two adjoining pairs of intermeshing teeth so that entrapment of fluid should not occur.
• the numbers of teeth in the intermeshing gears is determined by the contact ratio. Preferably the number is minimised to give the largest possible ratio between outside diameter and pitch circle diameter and thereby maximise displacement per unit of face-width. However the need to achieve a desired contact ratio results in a greater number of teeth than would be typical in a spur gear hydraulic apparatus. In itself the greater number of teeth helps to reduce the amplitude of pressure ripple.
• the gears preferably each have at least 13 , preferably at least 17, gear teeth.
• the fluid displacement per revolution is from 5 to 500 ml, preferably from 15 to 300 ml.
• pressure relief recesses can communicate with the spaces between the teeth as they move into mesh at each end of the meshing pair of gears, in order to permit fluid to escape from the decreasing space between teeth into the outlet side of the apparatus as a tooth enters that space, the disposition of the pressure relief recess at one end of the gears relative to that at the other end thereof being staggered by an amount corresponding to the helical advance.
• Such recesses are regarded as important in embodiments in which no underlap is provided - that is to say, where at all times there is continuous driving contact across the entire width of two adjoining pairs of intermeshing teeth.
• Such recesses are also regarded as desirable in embodiments in which a small amount of underlap is provided. It is believed that if such recesses were not provided the underlap might have to be sufficiently large to present a risk of a substantial back-flow of fluid from the outlet to the inlet.
• each pressure relief recess is a slot straddling the common tangent to the pitch circles of the gears and communicating at one end with the high pressure port of the apparatus, the other end of said slot having a basic end surface beyond which there extends a subsidiary slot disposed wholly on that side of the contact line remote from the power input or output gear.
• U.S. Patent Specification No. 4548562 relates to a helical gear pump and states that in order to avoid a hydraulic lock that would reduce pumping efficiency and create pump noise as well as undesirable hydraulic pressure forces in a pump, it is common practice to use an end plate on each axial side of the meshing pump gears and to provide a recess in the end plates to permit discharge of the fluid trapped in the gear tooth space thus providing communication between that space and the adjacent port. This provides a pressure relief that prevents a build-up of pressure in the trapped fluid in the tooth space of the meshing gear teeth.
• Said specification aims to eliminate the need for using end plates with pressure relief recesses, but Figure 4 thereof nonetheless illustrates the known recesses said to be no longer required. There is no suggestion that these are staggered from end to end by an amount corresponding to the helical advance; and they are shown as having a simple square-ended shape.
• the hydraulic apparatus of the present invention is a gear pump
• the pressure relief recess and the fluid feed recess at each end of the gears being so shaped and positioned as to avoid communication with each other by way of said spaces, and the disposition of the fluid feed recess at one end of the gears relative to that at the other end thereof being staggered by an amount corresponding to the helical advance.
• each fluid feed recess is a slot interconnecting the bearings for the shafts carrying the gears and being spaced from the basic end surface of the associated pressure relief recess by an amount corresponding to the helical advance.
• the basic end surface of the pressure relief recess at that end of the gears containing the leading ends of their teeth is disposed in the plane containing the axes of both gears.
• at least the effective portions of those edges of the fluid feed recesses adjacent to the pressure relief recesses are straight and at right angles to the common tangent to the pitch circles of the gears.
• At least the effective portion of that edge of the fluid feed recess at that end of the gears containing the trailing ends of their teeth is disposed in the plane containing the axes of both gears.
• the meshing pair of gears is retained between floating pressure-loaded end seal plates in which the pressure relief and fluid feed recesses are formed.
• Elastomeric seals with suitable back-up arrangements are accommodated in grooves in the backs of the plates to isolate areas which are subject to pressurised fluid, the area and shape defined by the seals being arranged to bring the seal plates squarely into contact with the end surfaces of the intermeshing gears so that the overall closing force is greater by a small amount than the separating force generated by pressure within the casing of the apparatus.
• the working surfaces of the seal plates, where provided, and the surfaces of the gears in contact therewith are treated, for example by a peening/tumbling process, to provide minute cavities to retain hydraulic fluid, in order to provide a wear-resistant hydrodynamic film during operation.
• respective pressure relief recesses communicate with both ports of the motor at each end of the gears.
• the seal plates are supported over their periphery in the gear housing providing a small axial clearance between the gears and plates.
• the plates can be pressure loaded in strategic places so that deformation away from the gear faces by internal pressure is prevented.
• a controlled internal axial clearance is provided. eliminating drag during starting without excessive reduction in volumetric efficiency.
• Helical gear pumps and motors as described herein can be provided with sleeve bushing type bearings or anti- friction roller bearings in similar fashion to spur gear pumps and motors.
• apparatus in accordance with the present invention is used to supply, in the case of a gear pump, or be driven by, in the case of a gear motor, hydraulic fluid at a pressure of at least 100 bar (10 x 10 6 Pa) , preferably at least 220 bar (22 x 10 6 Pa) , typically up to 320 bar (32 x 10 6 Pa) , or even beyond.
• Fig. 1 shows a gear pump in schematic plan view
• Fig. 2 is a perspective view of gears and seal plates used in the gear pump of Fig. 1;
• Figure 3 is a diagrammatic view, on that end containing the leading ends of their teeth, of two intermeshing helical gears of a pump;
• Figure 4 is a similar view on that end of said gears which contains the trailing ends of their teeth;
• Figure 5 is a diagrammatic end view on a considerably larger scale, on that end of said gears containing the leading ends of their teeth, of the meshing zone of said gears, showing high and low pressure relief slots at both ends thereof;
• Figure 6 is a view on substantially the same scale as Figures 3 and 4 of the inner face of a seal plate for that end of the gears containing the leading ends of their teeth, the slots in the seal plate being similar but not identical to those shown in Figures 3 to 5; and
• Figure 7 is a view corresponding to Figure 6 of the inner face of a seal plate for that end of the gears containing the trailing ends of their teeth, the slots in the seal plate being similar but not identical to those shown in Figures 3 to 5.
• the gear pump comprises a casing 2, made up of two components 4, 6 bolted together, to retain two helical gears 8, 10.
• Each gear 8, 10 is mounted on a shaft, the shaft for the gear 8 being extended so that a projecting portion 12 can be driven by a prime mover.
• Two seal plates 14, 16, each in the shape of a figure eight, are provided. Seal plate 14 provides a seal between the casing and the parallel end faces at one end of the gears 8, 10. Seal plate 16 provides a seal between the casing and the parallel end faces at the other end of the gears 8, 10.
• the gear teeth are arranged to fit with close tolerance within the casing. Rotation of the gears causes them to carry hydraulic fluid in the spaces between the gear teeth, entrapped by the casing, from an inlet or suction side, to an outlet or pressure side.
• the gears are helical gears having eighteen external gear teeth.
• the helical advance is exactly 1, so that the leading end 18 of each tooth is in line with the trailing end 20 of the preceding tooth.
• the contact ratio in the transverse plane is slightly less than 2.
• Each tooth has flanks of involute profile, identical on both sides thereof and terminating at a narrow upper land or tip which is the arc of a circle centred on the axis of the gear.
• the bottom lands or roots between teeth are radiussed to avoid fouling.
• underlap there is a very small amount of underlap in the design. This is present to aid the release of fluid trapped in the diminishing space between teeth, as a tooth moves into that space. If the underlap were not present, because the contact ratio was higher (for example 2.2) there would at every instant in the operation be two continuous lines of contact across adjoining pairs of teeth.
• seal plates 14, 16 are of a generally known type and so need not be described in full detail here. Briefly, they are floating pressure-loaded seal plates. Elastomeric seals with suitable back-up arrangements are accommodated in grooves 22 in the backs of the seal plates to isolate areas which are supplied with pressurised fluid, the area and shape defined by the seals being arranged to bring the seal plates squarely into contact with the end faces of the intermeshing gears so that the overall closing force is greater by a small amount than the separating force generated by the pressurised fluid in the pump casing. As a result, leakage across the end faces of the gears is reduced to a minimum.
• recesses are provided on the inner faces of the seal plates to permit hydraulic fluid to flow respectively from and to the space between teeth. These recesses are described in detail hereinafter.
• the working surfaces of the seal plates, and the abutting surfaces of the gear teeth, are treated by a peening/tumbling operation to provide minute cavities, to retain hydraulic fluid. This then forms a hydrodynamic film when the pump is running, to prevent wear.
• a pump of the type shown in the drawings is easily capable of operating at the high pressures required for power-generating hydraulic pumps. It could of course also be used for transfer duties, where the lack of back-flow of the pump may be advantageous, without being essential.
• the hydraulic pump shown is able to operate at pressures up to and perhaps exceeding 320 bar (32 x 10 6 Pa) , with low viscosity hydraulic fluids whose viscosity frequently does not exceed 20 cs (2 x 10 "5 m 2 s "1 ) at normal operating temperatures. It could typically revolve at 2000 r.p.m. and deliver 300 ml of hydraulic fluid per revolution. Its power output could be 500 h.p.
• the seal plates 14 and 16 are each also provided, on their faces adjacent the gears 8 and 10, with a pressure relief recess and a fluid feed recess, which communicate with the spaces between the teeth in the meshing zone at each end of said gears.
• said recesses are so shaped and positioned that they very effectively prevent a build-up of pressure in the trapped fluid in diminishing tooth spaces of the meshing gears 8 and 10, and aid the subsequent filling of expanding tooth spaces of the meshing gears 8 and 10, whilst avoiding communication with each other by way of said spaces, the disposition of said recesses at one end of said gears relative to those at the other end thereof being staggered by an amount corresponding to the helical advance.
• the recesses in the seal plate 16 at that end of the gears 8 and 10 containing the leading ends of their teeth consist of a pressure relief slot 70 and a fluid feed slot 72
• the recesses in the seal plate 14 at that end of said gears containing the trailing ends of their teeth consist of a pressure relief slot 74 and a fluid feed slot 76.
• the slot 70 straddles the common tangent 50 to the pitch circles of the gears 8 and 10 and communicates at one end with the outlet port of the pump, the other end of said slot having a basic end surface 78 disposed in the plane 62 beyond which surface there extends a subsidiary slot or "nose" 80 disposed wholly on that side of the contact line 52 remote from the power input gear 8.
• the slot 74 straddles the common tangent 50 and communicates at one end with the outlet port of the pump, the other end of said slot having a basic end surface 82 disposed at a distance from the plane 62 corresponding to the helical advance beyond which surface there extends a subsidiary slot or "nose" 84 disposed wholly on that side of the contact line 52 remote from the power input gear 8.
• the slot 72 interconnects bores 86 in the seal plate 16 which communicate with the bearings for the shafts carrying the gears 8 and 10, and its edge 88 adjacent to the slot 70 is straight and at right angles to the common tangent 50 and is spaced from the basic end surface 78 of the slot 70 by an amount corresponding to the helical advance.
• the slot 76 interconnects bores 89 in the seal plate 14 which communicate with the pump bearings, and at least the effective portion 90 of its edge adjacent to the slot 74 is straight and at right angles to the common tangent 50, is spaced from the basic end surface 82 of the slot 74 by an amount corresponding to the helical advance, and is disposed in the plane 62.
• the tooth spaces in the meshing zone can communicate to the maximum possible extent either with the low pressure region 58 or with the high pressure region 60 as appropriate, whilst direct communication between said regions by way of said tooth spaces is only just avoided. For example, as shown in
• a pressure sensor located in the outlet port of a gear pump constructed as described has shown a marked reduction in pressure harmonics associated with the pressure ripple.
• seal plates 14 and 16 can be omitted and the slots 70, 72, 74 and 76 formed directly in the pump casing.
• the invention is broadly applicable also to a reversible gear motor in which either port can temporarily constitute the inlet for high pressure hydraulic fluid.
• respective pressure relief recesses are arranged to communicate with both ports of the motor at each end of the gears.
• the contact ratio be high, preferably at least 1.5 and, more preferably, at least 1.85.
• Compared to a modern spur gear pump there will need to be a reduction in the outside diameter (tip circle) in relation to pitch circle diameter and hence a reduction in displacement per unit of face- width of the gears, perhaps of about 30% compared to that of a typical optimised spur gear pump of the same pitch circle diameter.
• the root diameter of a gear pump employing helical gears in accordance with the invention can be larger than that of an optimised modern spur gear pump.
• the gears can employ large diameter shafts, providing an increase in stiffness, in comparison with an optimised modern spur gear pump.
• the increase in stiffness is proportional to the cube of the respective diameters.
• wider helical gears i.e. from end to end between the sealing plates
• This compensates at least in part, and perhaps in whole, for the reduction in displacement per unit of face-width.
• the helical gear pump of the invention also provides advantages in relation to fluid intake from the inlet side.
• Spur gears give rise to a pressure ripple on the inlet side, which is lessened by the use of helical gears.
• Spur gears can also suffer from difficulties of entrainment of hydraulic fluid, particularly at high speeds, to the extent that the inlet side is sometimes pressurised. Entrainment, even at high speeds, is not expected to be a problem, using helical gears in accordance with the invention.

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• 1995
  • 1995-07-07 GB GB9700560A patent/GB2304155B/en not_active Expired - Lifetime
  • 1995-07-07 JP JP50418896A patent/JP3972072B2/ja not_active Expired - Lifetime
  • 1995-07-07 EP EP95943151A patent/EP0769104B1/de not_active Expired - Lifetime
  • 1995-07-07 KR KR1019970700047A patent/KR970704968A/ko not_active Application Discontinuation
  • 1995-07-07 DE DE69511870T patent/DE69511870T2/de not_active Expired - Lifetime
  • 1995-07-07 WO PCT/GB1995/001610 patent/WO1996001950A1/en not_active Application Discontinuation
  • 1995-07-07 AT AT95943151T patent/ATE184080T1/de active
  • 1995-07-07 AU AU28928/95A patent/AU2892895A/en not_active Abandoned

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