Hydrodynamic Torque Converters

Abstract

A hydrodynamic torque converter or converter-coupling, either on its own or as part of a motor vehicle automatic transmission, wherein the impeller blades are adjustable over a range, so as to provide over said range, variation in the effective diameter of the impeller.

Description

This invention relates to hydrodynamic torque converters or converter couplings.

Every hydrodynamic torque converter, or converter-coupling, must include within its stationary housing three, functionally essential, and coaxially disposed members, namely the impeller, the turbine, and the reactor. The latter, in the case of a converter, is a non rotatable fixture connected to the converter housing, or some rigid extension thereof. In the case of a converter-coupling the reactor is, however, mounted on a one-way clutch, so that it can rotate freely in the same forward direction as the impeller and the turbine, but not in the reverse direction. This gives the machine a duality of function whereby, for turbine speeds from stall up to a predetermined coupling speed, it behaves exactly as a torque converter by effecting torque multiplication, but, for all turbine speeds equal to and greater than the coupling speed, the torque ratio becomes unity, that is to say torque multiplication ceases and the machine functions purely as a hydrodynamic coupling. Without the reactor member, torque multiplication would be a physical impossibility in any hydrodynamic machine. The above description is pertinent to the simplest type of basic machine in which each of the three essential members has only one element, or, in other words, has only one row of blading. The theoretically possible multi-element variants of the basic machine are, however, as numerous as the permutations which can be obtained from combining one or more element of each of the three members, and the one or more ways by which each such element can be connected into the kinematic chain of the machine. However, the number of these combinations which have been tried out, survived the experimental development stages, and emerged as practical machines, are comparatively few. The design objective, in most of these necessarily more complex machines, is usually to extend the torque multiplication range. More explicitly such extension may include firstly, improvement in stall torque, secondly, improvement in torque multiplication throughout the entire torque conversion speed range, and thirdly, extension of this torque conversion speed range. Multi-element turbine, and/or multi-element reactor machines are not uncommon, but the impeller member is usually, though not invariably, a single element.

The present invention is applicable either to any basic machine or to any variant thereof as generally previously described. The main aspect of the invention lies in the construction of the impeller and the way in which it functions is substantially modified so as to effect a substantial change in the performance of the machine as a whole.

According to the invention a hydrodynamic torque converter or convertor-coupling has an impeller, having within prescribed limits, an infinitely variable effective diameter. This is achieved by providing for the controlled and sensibly radial movement, in unison, of the impeller blades.

For geometrically similar converters running at equal input and equal output shaft speeds, the impeller reaction torque, and the power absorbed by the impeller, are both directly proportional to D.sup.5, where D can be any nominated diameter of the torus, but is usually either the outside diameter, or the design-path diameter. The aforementioned sensibly radial movement of the blades, will in itself effect a change in the geometrical configuration of the impeller, but, within the required and prescribed limits of this movement, the foregoing mathematical functional relationship will remain approximately true. A significant fact in this relationship is that due to the value 5 of the index of D, only a small change of the latter will produce a large change in both the torque and power absorbed by the impeller. Consequently such a variable diameter impeller has a potentially high rate of response which can make it a particularly effective means of maintaining a continuous state of balance in any power transmission system, where input speed and torque from the prime mover, and reaction torque due to the load on the output shaft, can all vary simultaneously, substantially, and rapidly. A case in point is the transmission system of a motor vehicle, since the engine throttle opening can be subject to sudden and random variation, while simultaneously large variations in tractive effort may be required to accommodate changes in road gradient. The invention thus also extends to motor vehicle transmissions.

Although existing types of hydrodynamic torque converters have many features which are advantageous when incorporated into an automobile transmission system they, none the less, suffer from some disadvantages one of which is their restricted capacity for torque multiplication. Road vehicles range from those having the high power/weight ratio of about 250 HP per ton to those having the low power/weight ratio of about 10 HP per ton, and even less for large commercial freight carrying vehicles. From experimental development of transmission systems for this large range of vehicles has emerged the fact that, if a gearbox is the only means of torque multiplication employed, then the required number of gear ratios ranges from four for passenger cars to about 12, and sometimes even more, for heavy commercial vehicles. The starting torque multiplication capability of the former is about 3.6 and of the latter is about 14. Now a type of converter can be built which is unconventional inasmuch as it has a counter rotating reactor which is kinematically connected into the output drive system from the turbine and, which by this means, can produce a very high stall torque multiplication of 10. Also the more conventional type of converter having a fixed reactor, can be built to provide a stall torque multiplication as large as 7. However, neither of these types of converter is suitable for an automotive transmission system because of their poor torque capacity for a given size. Of the converters which are suitable for automotive application, multielement machines can have a stall torque ratio greater than 4, but for single element machines the ratio is not usually greater than 3 and, for converter-couplings, having a high coupling efficiency, the ratio is only about 2. The reasons for this restricted torque multiplication capability are well known and will not therefore be stated here. However, on account of this limitation, known types of converter can supplement but not replace the gearbox in an automotive transmission system. Nontheless the incorporation of any presently known type of converter into the transmission system will usually reduce the number of gear box ratios required. This invention has for its primary objective to increase the stall torque ratio and to improve the torque multiplication throughout the conversion speed range so as to eliminate entirely the need for the gearbox in the transmission system of a high power weight ratio vehicle i.e. most passenger cars and public service passenger vehicles. There is some doubt, however, as to whether the invention could enable the torque conversion range to be extended sufficiently to eliminate the need for the gearbox in low power weight ratio commercial vehicles but I do not intend this doubt to be a limitation of the invention. In any case it would be advantageous in reducing the number of gearbox ratios required in these lower power/weight ratio vehicles. The elimination of the gearbox, however, gives rise to a second aspect of this invention.

The controlled infinitely variable radial movements of the impeller blades is accomplished without interruption to power transmission. This complete continuity and infinite variability of power transmission make it possible to provide a system for governing the engine so that it will always operate along any predetermined speed-torque or speed-power characteristic, such for example, as the optimum engine utilization curve. This invention does not provide any specific means of achieving such a governing system but it is believed that such a system will be well understood by those skilled in the art.

It is highly desirable that all the blades of the impeller move exactly in unison. Otherwise both the mechanical and hydrodynamic balance of the machine could be seriously and prohibitively disturbed. To move all the blades radially inwards at any engine speed up to and marginally greater than that at which full throttle BHP is developed requires a very substantial total force. For example, at this speed the centrifugal force on each of say 20 blades of the converter shown in the accompanying drawings would be more than 1 ton. The system required to effect the required blade movements must, therefore, in order to provide an adequate margin of available force be capable of developing a total force of say 30 tons. This can most effectively be achieved by hydrostatic means, which may mean a penalty in terms of size, weight, complexity, and cost but I feel that this penalty can be offset by improved performance and efficiency.

An embodiment of the invention will now be described, by way of example, with reference to the accompanying drawings, wherein;

FIG. 1 is a longitudinal sensibly vertical sectional elevation of a vehicle transmission including a converter arrangement according to the invention;

FIGS. 2, 3 and 4 are related views showing details of the impeller core ring, the top and bottom parts of FIG. 2 being respectively views from opposite sides of the core ring, the top and bottom parts of FIG. 3 being sectional views taken on radial planes DO - OE in FIG. 2 and FIG. 4 being a circumferential section through an impeller entrance blade, the section being on line CC of FIG. 3;

FIGS. 5 and 6 are related views showing details of the impeller shell, the top and bottom parts of FIG. 5 being sectional views taken on the radial planes FO--OG in FIG. 6, and the top and bottom parts of FIG. 6 being quarter views from opposite sides of the shell;

FIG. 7 is a part view, in section of the (high pressure) HP cylinder and its circumferential ring, the section being taken on the plane K-L in FIG. 1 which is common to the center lines of all the impeller blade pistons;

FIGS. 8, 9 and 10 are related views of the composite component which comprises the impeller blade and the impeller blade piston, FIGS. 8 and 9 being respectively side and end elevations, and FIG. 10 being a view looking radially outwards; and

FIGS. 11 and 12 are related views of front ring of the (medium pressure) MP cylinder cover the top and bottom parts of FIG. 11 being both radial sections taken on the planes MO-ON of FIG. 12, and FIG. 12 being an elevational view of the rear face of the ring, but also including a part section of the cylinder cover taken on line RS in FIG. 1.

In FIG. 1 the input or front end of the transmission is located at the left hand end of the figure, while the output or rear end of the machine is at the right hand end of the figure.

In the interest of consistency and ease of understanding I have used the expressions "front" and "rear" in this description.

FIGS. 1 to 12 inclusive, illustrate and exemplify a specific method of applying this invention to a converter-coupling, the turbine and reactor members of which are fashioned substantially in accordance with well known practice.

Referring more particularly to FIG. 1, the turbine shell 4 is rivetted to its flange 5 and the latter has a forward extending hub which is rotatably mounted in the bush 3. This bush is a press fit in the thimble 2. The thimble provides a register for the flange on the rear end of the input shaft 1, also for the front plate 8 of the impeller shell and also for the thrust bearing 6. The turbine shell flange 5 is splined onto the front end of the output shaft 0, the rear end of which is rotatably mounted in the bearing 13. The inner race of the latter is clamped between a shoulder on the output shaft and the front face of the hub of the output flange 12 by means of a set screw and a disc. The output flange 12 is splined to its shaft 0. From the foregoing description it is apparent that the turbine and its shell 4 and its shell flange 5 and the output shaft 0 and its flange 12 comprise, as thus assembled, a composite rotating unit, which will hereafter be referred to as the Output Assembly.

The housing 14 of the output shaft bearing 13 is fitted with a cover 15 in which the oil seal 16 is mounted. The front face of the output shaft bearing housing 14 is registered onto the rear face of the flange 28 and this, in turn, has its front face registered into the rear face of the cover 27 which accommodates a (low pressure) LP gear pump. This gear pump comprises the body 31 and its functional rotating members which are the driven externally toothed gear 33 and its internally toothed mating gear 32. This LP gear pump is fitted with an unloader piston 34 and its spring 35 both of which are accommodated in the cover 27. The front face of the LP gear pump cover 27 is registered and bolted onto the rear cover 26 of the main housing 25. The latter has a flange at its front end, whereby it is grounded, that is to say bolted onto some rigid fixture, such for example, as the vehicle engine block. The flange 28 is splined onto and also by means of the nut 30 locked onto a very stiff tubular member 29. The latter is concentric with and accommodates within its bore the output shaft 0. It extends forward as a cantiliver and its front end is splined into the inner ring 19 of the oneway clutch on which the reactor 17 is mounted. Midway along its length this tube 29 is also splined and registered into a hub which is integral with the body 68 of a multi-plate clutch. From the foregoing description and from FIG. 1 it is apparent that the aforementioned components 19, 29, 68, 28, 14, 15, 27, 26 and finally the main housing 25 are all rigidly interconnected so that they form a composite unit which is effectively grounded by means of the flange at the front end of the main housing. This composite stationary unit will hereinafter be referred to as the Housing Assembly.

The reactor 17, the core ring and shell ring of which are integral with its aerofoil section blading, is conventional in both its fabrication and method of mounting. Components 18 and 19 are respectively the outer and inner rings and 20 and 21 are the end rings of a one way clutch, whilst components 22 are circumferentially equi-spaced sprags or rollers. Components 23 and 24 are the thrust bearings which ensure that correct running clearances are maintained at the fluid transfer faces from impeller to turbine and from reactor to impeller.

The Output Assembly, the Housing Assembly and the reactor as described above are of known configuration. The design and construction of the impeller and the mechanism for providing and controlling the radial movement of its blading, however, is to the best of my knowledge, novel and will now be described.

From FIGS. 2, 3 and 4 which are the related views showing details of the impeller core ring 9 it will be seen that this component has the general configuration of a spoked wheel, which has both a hub and a rim. The rim is quite broad in the radial dimension. It is apparent from FIG. 1 that the impeller is a mixed flow unit, since, in transit through the impeller, the flow changes from an axial direction at the entrance to a radial direction at the exit. The parts of the impeller which are analogous to the spokes of a wheel are actually fixed aerofoil section blades at the impeller entrance. The blade circumferential section, as shown in FIG. 4 is also cambered so as to provide the optimum entrance angle (.theta.) most suitable to receive the discharge flow from the reactor. The direction of this discharge depends on the turbine speed and varies through a very wide angle. As is apparent in FIG. 4 the trailing edge of each of these fixed impeller blades is recessed, in the form of a V-shaped groove, which extends radially inwards until it breaks through into the bore. This groove also extends radially outwards, where, in cutting through the convex rear face of the rim, it actually becomes a V-bottomed but parallel sided slot the depth of which runs out to zero at the outside circumference of the rim. The width between the parallel sides of this slot, and the angle of its V-bottom are such as to provide an accurate sliding fit for the front edge of the impeller blade. This blade is shown in detail in FIGS. 8, 9 and 10. As can be seen in the bottom half of FIG. 2, which is a quarter rear view, there are 20 of these V-bottomed parallel sided radial slots equi-spaced in the rear face of the impeller core ring. When placed in its slot in the core ring each blade forms an axially rearward and a radially outward extension of the rear edge of the corresponding fixed blade at the impeller entrance. The fixed and radially slidable parts of each blade thus comprise functionally one composite blade which has continuous smooth surfaces conducive to streamline flow.

The impeller shell 7 is shown in FIGS. 5 and 6. The shell hub has a forwardly extending stepped diameter spigot which is an accurate shrink fit into the mating bore of the core ring. Machined in the rear face of the shell are 20 equi-spaced radial slots, which at the center, where they break through into the bore, and as seen in the lower half of FIG. 5, are very deep and V-bottomed, but, where they break through at the circumference, are quite shallow and square bottomed. Core ring and shell have to be jig assembled so that their radial slots are in perfect alignment. Once assembled there is little tendency for the accuracy of the initial alignment to be disturbed even when the impeller is fulfilling its normal function under load. Nonetheless, to provide a positive registration between the core ring and shell, which will completely inhibit any slight angular displacement between them the two components are finally locked together by means of three grub screws. The holes for these are drilled and tapped into the front face of the hub of the assembled shell and core ring. They are parallel with the bore and equi-spaced on a pitch circle which is actually the interface between the shell spigot and its mating bore in the core ring. When thus assembled, and locked together, the core ring and shell are functionally one rigid composite unit into the rear face of which the 20 impeller blades can be assembled. If all the blades are now positioned in their slots so that they are equi-distant from the axis of the bore of the core ring and shell assembly, it will be seen that the whole assembly, core ring, shell and blades, has the configuration of the impeller of a centrifugal pump. In other words the assembly has the configuration of a normal type of torque converter impeller except for the fact that the blades do not extend radially outwards to the interface between the impeller exit and turbine entrance. The 20 passages which are thus formed between core ring, shell and blades comprise the axially rearward, at entrance, to radially outward, at exit, mixed flow section of the impeller. All surfaces of the passages are smooth and continuous and conducive to streamline flow. Now if the impeller, as assembled in accordance with the foregoing description, was rotated about the axis of its bore, then each blade would be subjected to centrifugal force and, if not constrained, would slide radially outwards to the full extent permitted by its slot. For automotive transmission purposes the impeller is usually directly connected to, and therefore runs at, the same speed as the engine, and it is necessary to be able to run the latter at speeds up to and marginally greater than the speed at which it develops full throttle BHP. At such speed the centrifugal force on each blade would be large, for example, in excess of 1 ton in the particular machine shown in FIG. 1. The corresponding required restraining force on each blade can most conveniently be provided by hydrostatic means.

To achieve this each blade is, as shown in FIGS. 8, 9 and 10, composite in construction and comprises the blade proper 10 (FIG. 1) which is rigidly and permanently attached to a cylindrical part 11 which is a hydraulic piston. Each blade is accurately machined all over, and its two faces are ground parallel so that the finished blade is of the exact uniform thickness required to slide freely but, without sideplay in the slots of the impeller core and shell assembly. The piston is also machined all over and its diameter ground to very close limits. Blade and piston are then accurately jig assembled so as to ensure that they are in perfect alignment in each of the elevational views as depicted in FIGS. 8 and 9. The two parts are then permanently united into a composite whole by brazing or other means which will provide the necessary mechanical strength without detriment to the surface finishes of blade and piston, or to the accuracy of the jigged dimensions and alignments.

Turning again to FIG. 1 37 is the (high pressure) HP cylinder block which as depicted in the part sectional view in FIG. 7 has 20 equi-spaced radial bores to receive the impeller blade pistons. These cylinders can only be bored, and finally ground, to the necessary close limits, in a radially inwards direction. Thereafter, their outward facing open ends have to be closed off by means of the high tensile steel ring 38 which is shrunk onto the outside diameter of the cylinder block. The HP cylinder block 37 and its ring 38 are therefore inseparable and comprise a composite member the front face of which is registered and bolted onto the rear face of the impeller shell. This registration must ensure, not only the axial alignment but also the angular positioning of the cylinder block 37 with respect to the impeller shell 7 so that the axis of each of the 20 equi-spaced HP radial cylinders is accurately aligned with the radial center line of the corresponding impeller blade slot. Extreme accuracy of these alignments is necessary in order to ensure that the two components of each composite blade and piston unit can respectively slide freely in their mating slot and mating cylinder. Now each HP radial cylinder communicates directly with a corresponding longitudinal cylinder. There are therefore 20 of the latter equi-spaced on a common pitch circle, and each having its axis parallel with the main axis of the machine. Each HP longitudinal piston 36 has the same volumetric displacement as its corresponding HP radial impeller blade piston 11 but is not necessarily of the same bore and stroke. In fact, in the particular design shown in FIG. 1 the stroke of the horizontal pistons is double that of the radial pistons. The whole of the interspace enclosed between each radial and corresponding longitudinal piston is completely filled with either a high viscosity oil or a nonsolid grade of grease which functions as a hydrostatic medium. Both the high viscosity of the medium and exceptionally high accuracy of fit of each piston in its bore is essential in order to completely inhibit leakage of the hydrostatic medium when it is subjected to the extremely high pressures which are required to hold the blade-piston unit in position, or to move it radially inwards against the centrifugal force acting upon it, when the impeller is driven at high speeds. This hydrostatic pressure can be of the order of several tons per square inch and it is therefore essential that the piston block 37 is fabricated from a high tensile material such as cast steel. To ensure that the 20 blade-piston units move precisely in unison the rear end of each HP longitudinal piston is abutted against the annular front face of the large stepped diameter and short stroke floating piston 40. This is designed as the (medium pressure) MP hydrostatic piston and it has a sliding fit both inside its cylinder 43 and also on the outside cylindrical face of the tubular member 41. Both these sliding surfaces of the MP piston are fitted with sealing rings but these are of a conventional type used for hydrostatic pressures of the order of 1 ton per square inch and are only shown diagrammatically in FIG. 1. The MP cylinder 43 is in the form of an annular ring which is clamped between the rear face of the HP cylinder block 37 and the front face of the MP cylinder cover 44 by means of a large number of high tensile tie bolts 47 equi-spaced on a common pitch circle. The internal front face of cylinder cover 44 has a forwardly extending boss onto which the ring 45 is clamped by means of the flanged rear end of the tube 41. The front end of this tube is not only registered into the rear face of the impeller shell 7 but it also has 20 projections or castellations 42, similar to those of a dog tooth clutch, which are a push fit into the impeller shell blade slots. The bore of the cylinder cover 44 houses a bearing 51 the inner race of which is mounted on a boss which projects from the front end of the body 68 of the multiplate clutch. The cylinder cover 44 is also bored to accommodate the driven rotor 53 and the two intermeshing followers (not shown in FIG. 1) of a conventional type screw pump. Such a pump is capable of developing delivery pressures of up to 1 ton per square inch. The pump shaft 54 is mounted in the thrust bearing 56 and the radial bearing 55 and is fitted, at its rear end, with the planetary gear 58. The latter meshes with the sun gear 64 which is integral with the sleeve 65. The screw pump shaft bearings 55 and 56 are housed in the ring 57. The ring 57 and the planetary gear housing 59 and the ring 46 are successively registered, one into the other, and the latter onto the rear face of the MP cylinder cover 44 to which they are all secured by means of the bolts 39. Pressed into the bore of the ring 46 is a shell 62 the rear end of which is fitted with the vanes 63 so that it functions as a centrifugal pump. To the rear face of the planetary gear housing 59 is registered and bolted the cover 60. To cover 60 is registered and rivetted the rearwardly extending tubular member 61 which is rotatably mounted in the bearing 52. This bearing is mounted in the rear cover 26 which is part of the Housing Assembly. From the foregoing description and from FIG. 1 it can now be seen that all the following components comprise one rigid assembly; 1 Input shaft 8 Impeller shell front plate 7 Impeller shell 9 Impeller core ring 37 HP cylinder block 38 HP cylinder block ring 43 MP cylinder 41 MP cylinder tube 44 MP cylinder cover 45 MP cylinder cover ring front 46 MP cylinder cover ring rear 57 Screw pump bearing housing ring 59 Screw pump planetary gear housing 60 Screw pump planetary gear housing cover 61 Screw pump planetary gear housing cover tube 62 LP centrifugal pump impeller shell The above assembly which is driven by and at the same speed as the input shaft 1 will hereinafter be referred to as the Input Assembly. Mounted within this assembly and rotating with it but having a degree of freedom of relative movement are: 10 and 11 Impeller blades and piston units 36 HP pistons 40 The MP piston 53 and 54 Screw pump rotor and shaft 58 Screw pump planetary gear

To the rear face of the rear cover 26 of the Housing Assembly is registered and bolted the cover 27 which houses a conventional type of low pressure gear pump. Components 32 and 33 are respectively the internally toothed annular gear and the mating externally toothed gear, which are mounted in the pump body 31. This pump is driven by means of dogs in the bore of the externally toothed gear 33 which engage with mating dogs on the rear end of the tubular member 61. The flow from this pump is discharged into a space in the cover 27 so that the delivery pressure acting on the annular stepped face of the unloader piston 34 causes it to lift against the action of its spring 35. This displacement of the piston permits the oil discharged by the pump to flow through a number of equi-spaced radial holes, near the rear end of the sleeve 65 and so into the annular interspace between this sleeve and the tube 29. After entering this interspace the oil then flows forward until it passes outwards through a number of equi-spaced radial holes near the front end of the sleeve 65 and so enters the impeller 63 of the LP centrifugal pump. This pump, which functions as a flow booster, discharges the oil forwards into the annular interspace between its shell 62 and the body 68 of the multiplate clutch. The oil thus enters and floods the pocket which is formed in the ring 46 at the rear end of the rotor 53 of the screw pump. In a manner, yet to be described, this pump only runs intermittently and for short periods. At all other times the discharge from the LP centrifugal pump therefore continues to flow forward and so passes through the interspaces between the elements of the bearings 51 into the annular interspace between the tube 29 and the tube 41. The oil then continues to flow forward until it passes through the interfacial clearances between the shells of the turbine and reactor and the shells of the reactor and the impeller. Thus it enters the torus of the converter where it is entrained with the oil which is circulating therein as the hydrodynamic medium. An equal rate of outflow from the torus must occur at the circumferential interfacial clearance between the impeller and the turbine. This oil is then discharged centrifugally through holes in the shell periphery and located just to the rear of its flange. The oil then gravitates to a sump, as the bottom of the housing 25, from whence it is piped back, via an oil cooler, to the suction port of the LP gear pump, and thus recommences the circulation cycle. The circulatory flow of oil, apart from fulfilling the usual function of continuously bleeding off and cooling the hydrodynamic medium from its working space within the converter torus, is sufficient in volume to ensure that the rear inlet end of the screw pump 53 is always kept fully flooded. Also, dependent on the rating and loading of the spring 35 of the unloader valve, oil is always available at a predetermined pressure at the inlet port of the spool valve 71. This is a normally closed, two position on/off, solenoid operated valve with spring closure. When the solenoid is energized this valve admits oil, via the radially drilled hole in the flange 28 and the long forwardly extending hole in the wall of the tube 29 to the working space at the head of the piston 69. The consequent displacement of the piston against the action of the clutch spring 70 effects engagement of the clutch. This locks the sun gear sleeve 65 (which is otherwise free to rotate on its bearings 66 and 67) to the stationary tube 29. Thus in travelling round the sun gear 64, which is integral with the sleeve 65, the planetary gear 58 is compelled to rotate on its own axis and thus drive the screw pump rotor 53. This pump, which has the capability to produce a delivery pressure of up to about 1 ton per square inch, then begins to pump the oil which is continuously available at its flooded rear end into the space in the cylinder cover front ring 45. This space, as shown in FIG. 12, extends in an arc on either side of the pump rotor axis so as to accommodate two non-return valves 72. These permit the oil delivered by the screw pump to enter the MP cylinder 43. The consequent forward displacement of the MP piston 40 produces an equal displacement of the twenty HP pistons 36 each of which produce a commensurate radially inwards movement of its corresponding blade piston 11. All the impeller blades are thus moved radially inwards, in unison, and the effective diameter of the impeller thereby reduced. Since this reduction in the effective diameter of the impeller is brought about merely by operation of the spool valve 71 it would obviously be a simple matter to connect the electrical leads of the spool valve solenoid to any suitable governor wherever located, whether on the engine or in the engine throttle control system or elsewhere. Conversely an increase of the diameter of the impeller, by moving the blades radially outwards, is easy to effect because this is their natural tendency anyway, under the influence of the centrifugal force acting upon them. It is only necessary to provide the constraint to ensure that they move strictly in unison. All that is necessary to do this is to release a volume of oil from the MP cylinder which is commensurate with the required increase in effective diameter of the impeller. The blades will then automatically move radially outwards. Furthermore, they will be constrained to move in unison because the centrifugal force acting on each blade will continuously pressurize the hydrostatic medium enclosed between the blade piston 11 and its corresponding longitudinal HP piston 36. The rear ends of all the latter will therefore be pressed up against and maintained in continuous contact with the annular common front face of the MP piston 40 as it moves in a rearward direction. A simple mechanism for automatically controlling the release of oil from the MP cyliner is shown in FIG. 1. The ring 45 is bored so as to expose the head of a piston valve 48 to the hydrostatic pressure in the MP cylinder. This valve is loaded by means of the compression spring 49. The head of the piston is fitted with a throttling bush or choke 50. The aperture in the latter is normally sufficient to ensure that the pressures on the two sides of the piston 48 are equalized and this valve is therefore retained in the closed position by the action of its spring 49. The bottom of the valve cylinder is however connected by a communicating hole to the annular port of a small normally closed piston valve 73. The latter can, however, move radially outwards under centrifugal force and against the action of a compression spring 74. Component 75 is an interchangeable weight. By selecting an appropriate size of weight, and a spring of appropriate rating it can be arranged that the valve 73 will open at any predetermined speed of the Input Assembly, since as previously described, the ring 46 in which the valve 73 is radially mounted, is a part of this assembly. But the aperture of the valve 73 is large as compared with the aperture of the choke 50 in the head of the piston valve 48. A pressure differential will thus be created between the two sides of the latter which will be sufficient to ensure that it opens rapidly, and so as to fully unmask an annular port in the piston valve cylinder wall. This will permit the rapid discharge of oil from the MP cylinder into the main housing rear cover 26 from whence it will gravitate into the sump of the main housing 25 and so re-enter the LP circulation system. This discharge of oil from the MP cylinder will permit a commensurate radially outwards movement of the impeller blades thus producing an increase in the effective diameter of the impeller. Since, as already stated, the impeller reaction torque is proportional to D.sup.5, the prime mover will be slowed down rapidly, so that the valves 73 and 48 will close successively and the MP piston 40 will thus be hydraulically locked in some new position to the rear of its previous position. This means that the impeller blades are also hydraulically locked in their new position of correspondingly increased effective diameter of the impeller.

The function that the automatic centrifugally operated pilot valve 73 fulfills in operating the piston valve 48 and so unloading the MP cylinder could equally well be fulfilled by a solenoid operated valve. The latter would, in this case, however, have to be mounted, so that it was not influenced by centrifugal force acting upon its moving parts. That is to say the valve would have to be mounted with its axis horizontal and parallel with the main axis of the machine. It would preferably be housed in the MP cylinder cover 44 alongside the corresponding piston valve 48 which it controls. The only complication would be that since the MP cylinder cover is a part of the rotating Input Assembly it would be necessary to connect the solenoid leads to their switch by means of slip rings. This would not, however, be a problem since only low voltages would be involved. The leads from the slip ring bushes could be extended and connected into the same governor control system as already mentioned. This would then mean that all variations to either increase or to decrease the effective diameter of the impeller would be under the control of a single unified governing mechanism. Governing of the engine to ensure its operation along any prescribed speed-torque or speed-power characteristic would then be feasible.

Since automotive transmission systems are subject to sudden large changes of engine torque and power due to random operation of the throttle controls, and also can be simultaneously subject to sudden large changes in load due to abrupt variations in road gradient, it is essential that any torque converting device should have a high rate of response. Otherwise undesirably large variations in engine speed could occur. Particularly, is it desirable to avoid setting up, even transiently, conditions where the engine would tend to either race or to stall. To this end it is desirable to provide for high rates of charging and discharging the MP cylinder 43 in order that all required changes of the effective diameter of the impeller are implemented rapidly. To achieve this it is desirable to provide two or more screw pumps 53 and two or more MP cylinder discharge valves 48 all equi-spaced on a common pitch circle. The arrangement shown in FIG. 12 includes three pumps and three discharge valves. Quite apart from providing for an improved response rate two or more pumps and discharge valves are also essential to ensure that the rotating Input Assembly, of which they are a part, is dynamically balanced, and also to ensure that the loadings of the sun gear 64 and the bearings 66 and 67 of the sun gear sleeve 65 are symmetrical.

There is a further advantage in having a multiplicity of MP cylinder discharge valves inasmuch that it can easily be arranged for their operation to occur at pre-determined stepped speed intervals. Thus dependent on the magnitude of any particular increase in engine speed so one, two, three or more valves would operate simultaneously to effect the required correction. That is to say the response rate would be approximately proportional to the magnitude of the required reduction in engine speed.

The impeller blade control system as described in the foregoing paragraphs incorporates two important practical features, inasmuch that both the HP and the MP hydraulic systems are specifically designed with the object of circumventing sealing problems.

The HP system is a "closed" system. That is each blade-piston 11 and its associated longitudinal piston 36 completely enclose and seal off a specific volume of the hydrostatic medium, and the inhibition of its leakage from the system is dependent on the high accuracy of fit of both the blade-piston and the longitudinal HP piston in their respective bores. There are no piston rings or valves would could malfunction in service.

The MP system is not a "closed" system inasmuch as the volume of oil in the working space behind the piston in the MP cylinder is variable. Oil has to be pumped into the cylinder at pressures of up to about 1 ton per square inch, and the cylinder, being a part of the Input Assembly, is rotated at engine speed. Consequently, but for the fact that the MP screw pump is itself also a part of the Input Assembly, some difficult and possibly insuperable sealing problems could arise. As it is, however, the MP screw pump is able to deliver oil directly into the MP cylinder, and by this means an otherwise difficult sealing problem is entirely circumvented. The fact that the pump is a part of the rotating Input Assembly does, however, necessitate the rather more complex arrangements for its drive by means of the planetary and sun gears 58 and 64 and the sleeve 65 and the multiplate clutch.

There is also an advantageous design feature with regard to the impeller blades 10. Referring to FIGS. 8 and 9 each blade has an extension 77 which is the same thickness as the blade and is rectangular in section. Previously it was explained that the 20 equi-spaced radial slots in the rear face of the impeller shell, where they break through at the circumference, are quite shallow and square bottomed. This part of each slot provides a slideway for the extension 77 of each blade. The purpose of this extension is to fill the blade slot which would otherwise be exposed in the front face of the impeller shell when the blade was in any position other than at the extreme outward end of its range of movement. The fact that the impeller blade extension 77 always fills the blade slot, means that the front inside face of the shell, up against which the oil circulating in the torus always presses, is a continuous smooth surface, for all positions of the blades. The objects of this is to avoid creating turbulence of the oil circulating in the torus.

Claims

I claim:

1. A hydrodynamic torque converter or converter-coupling comprising a reactor, an impeller having a core-ring and a shell between which is an interspace for fluid flow, a plurality of equi-spaced radially disposed and radially fixed blades at the impeller entrance, the radially fixed blades being attached to the impeller core-ring and receiving the discharge from the reactor, the core-ring and shell each having a plurality of radially extending slots, a plurality of blades slidably mounted in the slots to move radially, the movable blades being angularly equi-spaced and dividing the interspace into flow sectors, means for causing the movable blades to move radially in unison to provide an impeller of variable effective diameter, and the slots being disposed to hold each blade in exact co-planar alignment with the trailing edge of a different fixed blade of the impeller core-ring.

2. A converter according to claim 1 wherein each slidable impeller blade has a rearward extension which protrudes beyond the rear face of the impeller shell, and wherein the means for causing the movable blades to move in unison includes a plurality of high pressure pistons, the rearward extension of each blade having secured to it a radially and outwardly disposed high pressure piston for radially constraining the movable blade, said movable blade and piston forming one composite rigid unit.

3. A converter according to the claim 2 wherein the high pressure pistons of the blade-piston units are arranged to slide in equi-spaced radial cylinders which are bored in an annular-shaped, high pressure piston block which is concentrically registered with and rigidly secured to the rear face of the impeller shell, whereby the axes of the radial cylinders lie in a common plane which is at right angles to the converter axis.

4. A converter according to claim 3 wherein each radial cylinder is in sole and direct hydraulic communication with a horizontally extending cylinder, the horizontal cylinders being bored in the same common high pressure cylinder block as the radial cylinders, and being equi-spaced on a common pitch circle about the axis of rotation of the converter.

5. A converter according to claim 4 wherein the converter has an input assembly comprising the input shaft, the impeller and all parts connected to the impeller for rotation therewith, the arrangement being that when the input assembly is driven at speed, the centrifugal force acting on each impeller blade-piston unit pressurizes the hydrostatic fluid enclosed in the working space between each radial and horizontal piston pair so that each horizontal piston is forced rearwards into abutment with the annular front face of a common short stroke and large diameter medium pressure hydraulic piston, the cylinder of which is coaxially mounted onto the rear face of the high pressure cylinder block so that under the said condition of abutment, the tips of the radially slidable impeller blades will all lie on a common circle about the converter axis, so that the impeller is dynamically balanced.

6. A converter according to claim 5 wherein the medium pressure cylinder has its rear end closed by a cover, onto the rear face of which are bolted a succession of concentric annular members, the rearmost of which is tubular and is rotatably mounted in a bearing in the converter housing, so that this bearing comprises the rear support for the input assembly.

7. A converter according to claim 6 wherein the tubular rear member of the input assembly protrudes rearwards through its bearing and has its rear end drivably connected with the inner gear member of a conventional type gear pump, the suction port of which is pipe connected to the sump in the base of the converter housing, and whereby this pump provides fluid at low pressure to serve the whole hydraulic system of the converter.

8. A converter according to claim 7 wherein the medium pressure cylinder rear end cover houses two or more medium pressure screw pumps equi-spaced on a common pitch circle about the converter axis, each pump having a rearwardly extending shaft on which is mounted a planetary gear wheel, said gear wheel meshing with a common sun gear wheel having an integral tubular shaft concentric with and rotatably mounted on a stiff cantilevered non-rotating tubular member which, at its rear end, is rigidly secured to the converter housing and, at its front end, provides a mounting for the one-way clutch of the converter reactor member, and, midway along its length, provides a seating for a bearing which supports the input assembly, and also a rigid mounting for the casing of a multiplate hydraulically operated clutch, the arrangement being such that when the clutch is disengaged, the sun and planetary gears all rotate bodily, but without relative movement, about the main axis of the converter so that the medium pressure screw pumps all remain inoperative, but when the clutch is engaged the sun gear is held stationary, the planetary gears in rotating about the sun gear are compelled to rotate on their own axes and so drive the medium pressure screw pumps, and each of said screw pumps delivers fluid through one or more non-return valves directly into the medium pressure cylinder, so that the resulting forward displacement of the medium pressure piston causes an equal forward displacement of each and every horizontal high pressure piston, each of which in turn, causes a commensurate radially inward displacement of its associated high pressure blade-piston, and thereby the impeller blades are each moved radially inwards by an equal amount and in unison so that throughout this movement while the effective diameter of the impeller is being reduced, the dynamic balance of the impeller is always maintained.

9. A converter according to claim 8 wherein the said conventional type gear pump has its output connected to supply a flow of cooled hydrodynamic fluid to the converter torus, to keep the suction end of each medium pressure screw pump fully flooded, and to provide fluid whenever required, through a pressure regulating valve and a normally closed, balanced type of solenoid operated on/off hydraulic valve to engage rapidly the multiplate clutch so that the medium pressure screw pumps become operative.

10. A converter according to claim 9 wherein the medium pressure cylinder rear end cover houses two or more spring loaded, normally closed, piston type, medium pressure cylinder unloader valves, equi-spaced on a common pitch circle about the converter axis, and that each such valve, independently of the others can be un-balanced so that it opens and exhausts the hydraulic fluid from the medium pressure cylinder and so permits the medium pressure piston and each of the horizontal high pressure pistons to move equally and in unison rearwards due to the hydrostatic pressure created by the centrifugal force generated by each impeller blade-piston unit, and so that the latter move radially outwards in unison and by the same amount and so that throughout this movement while the effective diameter of the impeller is being increased, the dynamic balance of the impeller is always maintained.

11. A converter according to claim 10 wherein each of the two or more medium pressure cylinder unloader valves is unbalanced and opened by means of an adjacent normally closed solenoid operated balanced type of on/off hydraulic valve, mounted longitudinally in the medium pressure cylinder rear cover.

12. A converter according to claim 11 wherein the volume of fluid in the medium pressure cylinder acts as an hydraulic lock capable of maintaining the effective diameter of the impeller at any particular value for any required period of time, as long as there is no flow of fluid into or from the medium pressure cylinder.

13. A converter according to claim 11 wherein the solenoid operated valves which respectively initiate charging and discharging of the medium pressure cylinder to initiate the decrease and increase of the effective diameter of the impeller, are connected into a separate and external control system designed to achieve automatic operation of the engine along any preselected speed power line, such as for example the optimum engine utilization curve.