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.

Many conventional hydraulic gear pumps have intermeshing, external spur gears. Hydraulic fluid is carried by the gears, in the spaces between the teeth and enclosed by the pump casing, from an inlet side to an outlet side. Such pumps produce a pulsating discharge, occasioned by a tooth entering a space between teeth on the co-operating gear, causing the hydraulic fluid therebetween to be expelled. This pulsating discharge gives rise to a pressure ripple which may cause resonant vibration of mechanical components, which in turn produces noise. It would be desirable to reduce the magnitude of the pressure ripple, by smoothing out the fluid discharge, in order to reduce the noise.

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.

For transfer pumps, which operate at relatively low pressure and/or handle viscous fluids, such methods are quite satisfactory. There is likely to be some back-flow from the outlet side to the inlet side through the intermeshing teeth, but this is generally not of concern. However, such methods are not satisfactory for high pressure hydraulic pumps, and/or those handling fluids of low viscosity. Such pumps could be high specification transfer pumps but are more typically power generating hydraulic pumps, that is to say pumps in which the hydraulic fluid is used as a working fluid to operate hydraulic equipment. 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.

Thus, the industry standard 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.

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. In a 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. At speed, the torque variations are not of great significance; the noise generated by the flow ripple is. At low speed or when stationary, 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.

We have now determined that certain helical gears can be used for high specification applications, both as gear pumps and as gear motors, and give substantial benefit over corresponding apparatus having spur gears. In both applications there are substantial benefits in noise reduction; and in gear motors, better start-up characteristics.

In accordance with the present invention there is provided 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.

As the gears rotate the continuous driving contact breaks at the leading end of a helical tooth, and substantially at that instant there is already, or there is almost, continuous driving contact between the succeeding pair of teeth.

It should be noted that the term "substantially at the instant" implies that a small deviation either way is

permitted from the leading end coming out of engagement with its co-operating tooth at precisely the instant that continuous driving contact across the succeeding tooth is achieved.

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.

To meet the requirements of the present invention as defined above it is preferred that 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. Suitably the helical advance is in the range 0.5-2, preferably 0.85- 1.2, most preferably 1. Suitably 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. This implies that very small amounts of 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. Suitably 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.

Thus, in embodiments of the present invention having a very small degree of underlap, 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. In practice the gears preferably each have at least 13 , preferably at least 17, gear teeth.

Suitably the fluid displacement per revolution is from 5 to 500 ml, preferably from 15 to 300 ml.

In hydraulic apparatus as defined above, 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.

Preferably, 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.

When the hydraulic apparatus of the present invention is a gear pump, in addition to the pressure relief recesses there are preferably also provided fluid feed recesses which can communicate with the spaces between the teeth as they move out of mesh at each end of the meshing pair of gears, in order to permit fluid to enter the enlarging space between teeth from the inlet side of the pump as a tooth leaves that space, 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.

Preferably, 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.

Preferably, also, 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.

Preferably, 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.

Preferably, also, 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.

When the hydraulic apparatus is a gear pump, preferably 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.

In hydraulic apparatus in accordance with the present invention, axial forces are generated which urge the gears in opposite directions. To combat these axial forces, and counter tipping moments which would adversely affect the sealing efficiency, extra area is provided in the lower part of one seal plate and the upper part of the other seal plate.

Preferably 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.

When the hydraulic apparatus is a reversible gear motor, respective pressure relief recesses communicate with both ports of the motor at each end of the gears.

It is desirable that a gear motor be able to run in both directions of rotation. Therefore the form of seal plate loading described above for a gear pump cannot be used. A modified form of seal plate loading has to be used to provide a similar degree of pressure loading regardless of which port is under pressure. Such types of pressure plates have been in use with spur gear motors for a considerable period of time, and are equally applicable to helical gear motors.

In gear motors the use of pressure loaded floating seal plates can be disadvantageous. Because the motor is pressurised before rotation can commence floating plates loaded against the end faces of the gears reduce the output torque until sufficient speed is attained to generate dynamic film lubrication. Such reduction of torque can be serious if high starting torque is a requirement.

Hence it can be advantageous if the seal plates are supported over their periphery in the gear housing providing a small axial clearance between the gears and plates. In addition the plates can be pressure loaded in strategic places so that deformation away from the gear faces by internal pressure is prevented. Thus a controlled internal axial clearance is provided.

eliminating drag during starting without excessive reduction in volumetric efficiency.

Such technology has been used successfully with spur gear motors and is equally applicable to helical gear motors.

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.

It would be possible to use double helical gears, axially spaced on their respective shafts, in order to avoid having to design to combat axial forces on the gears. However, such double helical gears would introduce their own design problems and their use, whilst possible in accordance with the present invention, is not preferred.

Suitably, 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.

The invention will now be further described, by way of example, with reference to the accompanying drawings, in which:-

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.

As shown in Fig. 1, 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 and seal plates will now be described in greater detail with reference to Fig. 2.

As shown in Fig. 2, 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.

With the helical advance and contact ratio chosen for this embodiment, at most instants during the operation, there is continuous driving contact, and so a continuous seal, across the entire width of a pair of intermeshing teeth. The contact line runs diagonally across the tooth flanks from the involute generation circle to the leading end. As the gear pair rotates the continuous seal will break at the leading end of the appropriate tooth. At this instant, there is almost, but not quite, continuous driving contact across the entire width of the succeeding pair of teeth. In fact, at this instant, there will be a small gap adjacent to the trailing end of those teeth. However, that gap immediately disappears as rotation continues, so that once again there is continuous driving

contact, and a continuous line of sealing, across intermeshing teeth. Thus, 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.

It might be thought that the design could be such that, at the very instant that continuous driving contact is achieved across one pair of teeth, continuous driving contact is lost along the preceding pair of teeth. This would be a system with no underlap. However, it is not thought feasible or economical to attempt to design or manufacture with such precision. In practice we believe that providing a small amount of underlap is perfectly satisfactory. Due to the hydrodynamic properties of the hydraulic fluid, even of a low viscosity hydraulic fluid, such a small amount of underlap is most unlikely to give any significant back-flow of hydraulic fluid from the outlet side to the inlet side. It is only sufficient to allow fluid to leave the diminishing space between teeth.

It may in fact be possible to rely on underlap alone, in allowing release of fluid in the diminishing space between teeth as intermeshing occurs. However, in this embodiment further provision is made to allow release of such fluid. This further provision is achieved by the seal plates 14, 16.

These 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.

Because of the use of helical gears, axial forces are generated within the mating gear pair, tending to move the gears in opposite directions indicated by the arrows in Fig. 2. To combat these axial forces, extra area is provided in the lower part of seal plate 14 and in the upper part of seal plate 16. The object is to ensure that a tipping moment, which could adversely affect the sealing efficiency, does not arise.

On the inner faces of the seal plates, recesses (not shown in Figure 2) are provided 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. In fact 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.

Such a pump has been found to operate with high efficiency, but with a very marked reduction in the pressure ripple and hence in the noise generated. Typically the pressure ripple has an amplitude about one quarter of that generated by a comparably-sized modern spur gear pump produced to the highest available standards of precision.

Referring now to Figures 3 to 7, to illustrate that the helical advance is exactly 1, a series of teeth 30, 31, 32 and 33 on the drive gear 8 and a series of intermeshing teeth 40, 41, 42 and 43 on the driven gear 10 have herein been given the suffix A at their leading ends and the suffix B at their trailing ends. The common tangent to the pitch circles of said gears is shown at 50, and the contact line, that is to say the line along which the teeth of said gears progressively make driving contact with one another, is shown at 52. The gear 8 is rotated by a prime mover in the direction of the arrow 54, and drives the gear 10 in the direction of the arrow 56. The rotation of said gears causes them to carry hydraulic fluid in the spaces between their teeth, entrapped by the casing, from a suction or low pressure region 58 communicating with an inlet port in the casing to a high pressure region 60 communicating with an outlet port in

the casing. The plane containing the axes of both of the gears 8 and 10 is shown at 62.

In addition to the grooves 22, 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. As will now be described, 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. Thus 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, and 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. As best seen in Figure 5, 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. In similar manner but with a significantly different shape as well as disposition due to the afore¬ mentioned stagger resulting from the helical advance, 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. In similar manner but with at least its disposition significantly different due to the afore-mentioned stagger resulting from 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.

By virtue of the relative shapes and dispositions of the slots 70, 72, 74 and 76, 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

Figures 3 to 5, when engagement is established at the trailing ends of a pair of co-operating teeth 32 or 42, so as to complete continuous driving contact across the entire face-width of the pair of co-operating teeth 32 and 42 and the leading end of the preceding tooth 31 is

simultaneously about to come out of engagement with its co-operating tooth 41 so as to disestablish continuous driving contact across the entire face-width thereof, the helical spaces between said teeth (which are interconnected by the backlash between the gears 31 and 42) are only at that instant no longer vented by the slots 70, 72, 74 and 76. The instant before, the diminishing spaces between teeth could vent into the pressure relief slots. The instant after, the expanding spaces between teeth can receive hydraulic fluid from the fluid feed slots. Consequently the helical spaces between teeth can empty and fill very efficiently when those spaces respectively diminish and expand. This leads to quiet operation and a minimization of cavitation damage, as well as to maintenance of the pump's designed displacement of hydraulic fluid. 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.

If desired, the 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. To accommodate these alternative modes of operation, respective pressure relief recesses are arranged to communicate with both ports of the motor at each end of the gears.

In the present invention it is thought important that the contact ratio be high, preferably at least 1.5 and, more preferably, at least 1.85. This in turn implies the

use of greater tooth numbers than would be customary with a modern spur gear pump. 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. However, because of the reduced outside 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. Hence 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. The result is that wider helical gears (i.e. from end to end between the sealing plates) can be allowed without exceeding the maximum allowable bearing pressures. This compensates at least in part, and perhaps in whole, for the reduction in displacement per unit of face-width.

It is believed that the helical gear pump of the invention, as well as providing a reduced pressure ripple on the outlet side, 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.

Application of the present invention to gear motors as outlined previously is also envisaged.

The reader's attention is directed to all papers and documents which are filed concurrently with or previous to this specification in connection with this application and which are open to public inspection with this specification, and the contents of all such papers and documents are incorporated herein by reference.

Each feature disclosed in this specification (including any accompanying claims, abstract and drawings) , may be replaced by alternative features serving the same, equivalent or similar purpose, unless expressly stated otherwise. Thus, unless expressly stated otherwise, each feature disclosed is one example only of a generic series of equivalent or similar features.