This invention relates in general to hydraulic machines such as multi-cylinder motors and pumps wherein the fluid is delivered to the cylinders through a flat or cylindrical distributor. The term "machine" as used hereinafter should be taken as designating a device capable of operating either as a motor or as a pump.
These well known machines comprise as a rule a rotary cylinder body or "barrel", an impeller or swash plate rigid with the driving shaft and carrying a number of swivel-mounted connecting-rods operatively connected to the pistons, and a distributor plate.
This invention is concerned more particularly with a device for balancing the axial forces acting upon one of these members.
It is known that the force with which the cylinder barrel bears against the distributor plate ranges as a rule from 5 to 10 percent (although certain cases are known where slightly different values are attained) of the thrust exerted by the hydraulic fluid against the bottoms of the various cylinders. This force is absorbed in general by dynamic "shoes" disposed at spaced intervals on the outer periphery of the barrel or distributor plate, which shoes give a bearing force depending on various parameters such as fluid viscosity, relative speed of said shoes in relation to the plate, surface condition, etc.
Thus, the barrel engages the distributor plate in the inoperative condition and moves away therefrom as it picks up speed.
This device operates satisfactorily when the barrel has been separated from the distributor plate. On the other hand, when starting the machine the pressure distribution on the sealing faces disposed on either side of the high-pressure port is unknown. In fact, when the barrel and the distributor plate are in mutual contact, the fluid cannot penetrate into the interface formed between the surfaces. In this case the machine, when starting, must overcome a considerable frictional resistance. Therefore, it is obvious that there is latent risk of jamming and abnormal wear by friction during the starting period, until the dynamic shoes become gradually more efficient with the increasing speed of the machine and will permit the "lifting" of the cylinder barrel and a more adequate pressure distribution.
In fact, "mixed" starting conditions are encountered due to the relative evenness of the contacting faces of the barrel and distributor plate. However, in the case of a motor starting under load, there is no sufficient time for a proper distribution of pressure on the sealing surfaces and the starting efficiency will be relatively moderate. This effect, also referred to as "sticking" of the barrel to the distributor plate, is such that a relatively high pressure may be necessary for starting the motor, the pressure resuming a normal value at the end of the starting period. Now this constitutes a major inconveniences in the case of slow hydraulic motors starting under load, as used inter alia for propelling vehicles. This effect is rendered still more unpleasant and detrimental by the fact that the unequal pressure distribution in the various portions of the sealing surfaces permits a certain "wedge" or skew lifting of the barrel, which constitutes a source of considerable leakage during the starting period.
Machines are also known wherein the barrel is supported by a hydrostatic bearing consisting of a controlled delivery of fluid under pressure at the sealing interfaces formed between the high-pressure and low-pressure ports, on the one hand, and the inner space of the casing, on the other hand ; however, it is noted that the nature of the pressure distribution is unknown, at least under one portion of the sealing surfaces, so that the inconveniences mentioned hereinabove in connection with dynamic bearings are also observed in the present instance.
This method of balancing pressures by using hydrostatic bearings is attended by other inconveniences.
Thus, the fluid supply grooves leading to the hydrostatic bearing must be very narrow (about 0.01 inch to about 0.02 inch), thus preventing the construction of motors having very small dimensions.
The sealing surfaces between the ports and the inner space of the machine casing must be much larger than in the case of balancing by means of hydrodynamic shoes since the latter do not require the machining, in the sealing lips, of grooves similar to those feeding the hydrostatic bearing. This is attended by a considerably greater separating force between the barrel and the distributor plate. To increase the bearing force of the barrel the cross-sectional area of the orifices formed through the bottoms of the cylinders are reduced, with the consequence that the inner losses of pressure of the motor are increased. It is also noted that a single radial groove is not capable of ensuring a correct feeding of the ends of the circular groove. Finally, a serious inconvenience of this balancing method lies in its heavy fluid consumption due to the fact that a gap in excess of 0.01 mm (0.0004 inch) must necessarily be provided between the sealing surfaces of the barrel and distributor plate in order to avoid an excessive sensitivity to fluid pollution.
The present invention while avoiding the inconveniences listed hereinabove provides a hydraulic motor having a starting torque substantially equal to the rated torque while permitting maintaining internal pressure losses at a value of the same order as that of motors having the cylinder barrel balanced by hydrodynamic shoes.
To this end, the high-pressure hydraulic machine according to this invention, of the type comprising a multi-cylinder barrel and a distributor plate revolving in mutual contact, wherein the cylinders of said barrel and the orifices through which said cylinders open into the interface between said barrel and said plate are disposed on a circle about the barrel axis together with the corresponding fluid inlet and outlet ports of the distributor plate acting as a hydrodynamic bearing, is characterized in that the mean diameter of the imaginary circle containing the centers of said orifices and ports is greater than the diameter of the imaginary cylinder containing the axes of said machine cylinders, whereby the resultant balancing force exerted by said barrel on said distributor plate be directed internally of the bearing surfaces of said hydrodynamic bearing which are preferably coincident with the barrel axis.
Another object characterizing this invention consists in providing a hydrostatic bearing of which the fluid consumption is negligible in comparison with the other motor leakages, with means for regulating or controlling the thickness of the fluid film separating the barrel from the distributor plate.
To this end, the hydraulic machine according to this invention is also characterized in that the distributor plate comprises a hydrostatic bearing fed with fluid under pressure for balancing the barrel in the inoperative position and that means for controlling the supply of pressure fluid to said hydrostatic bearing are provided in order to limit or reduce said supply during the normal operation of the machine.
Balancing devices according to this invention will now be described by way of example with reference to the attached drawing, in which :
FIG. 1 is a fragmentary longitudinal section of a motor with its distributor, the section being taken along the line I--I of FIG. 3 ;
FIG. 2 is a plan view of the distributor plate, as seen in the direction of the arrows II--II of FIG. 1 ;
FIG. 3 is an end view of the cylinder barrel, taken in the direction of the arrows III--III of FIG. 1 and
FIG. 4 is a plan view showing a modified form of embodiment of the distributor plate.
Referring first to FIG. 1, showing a cylinder block 1 rotatably mounted on a distributor plate 2, it will be seen that in this cylinder block a plurality of cylinders 3 are disposed at spaced angular intervals about the axis of rotation (0) of this block, the centers of said cylinders lying on a common circle having a diameter d0, so that the cylinder axes are parallel to said axis (0).
Each cylinder has slidably mounted therein a piston 4 operatively connected through a ball joint 6 to a connecting-rod 5 having its other end (not shown) provided with another ball joint attached to a rotary impeller or swash plate as well known in the construction of hydraulic motors. Driven bodily with this impeller or swash plate the cylinder 1 revolves inside a case or body 7 in a bearing 8.
The bottoms of these cylinders 9 open to the outside of the barrel and the latter is provided to this end with a series of spaced orifices 10 lying on a mean circle having a diameter d3 and limited by circles having diameters d1 and d2, respectively. These orifices 10 contact the distributing surface of plate 2 formed with a pair of arcuate ports 14 and 15 constituting the former the inlet port and the latter the exhaust or delivery port.
The distributor plate 2 is also provided with a number of peripheral hydrodynamic "shoes" 16 from which a bearing force is derived which depends on the viscosity of the hydraulic fluid and also on the velocity of motion in relation to the barrel.
Sealing surfaces 17, 18 bounded by circles having the diameters denoted d6 and d7 respectively are disposed on either side of ports 14 and 15.
The operation of this device will be better understood when considering the following remarks :
For each cylinder under pressure, it is assumed that the force pressing the barrel 1 against the distributor plate 2 is substantially equal to the force acting upon the piston.
The hydraulic reaction force per cylinder is produced by the pressure acting upon a surface defined by the angle a =(2π1n) (n being the number of cylinders of the pump) and by a pair of circles having the diameters d6 and d7 respectively, which are the outer diameters of the sealing surfaces. Due to the pressure drop across the sealing surfaces, it may be assumed that the reaction surface (S) will be defined by the diameters d4 and d5 equal to (d1 + d7 /2) and (d2 + d6 /2 ), respectively, and will in all cases be limited by the angle a = (2π/n), this surface (S) being shown by a thin-shaded area in FIG. 3.
This surface (S) is generally equal to 0.9 or 0.95 times the piston area ; in other words, when the pressure prevailing in the cylinder is exerted against this surface, the resultant force will be according to circumstances 0.9 to 0.95 times the force acting on a piston.
In the pumps and motors now available commercially the mean diameter d3 of orifices 10 is not clearly determined and varies between the value d0 of the diameter of the circle containing the centers of the bottom orifices of the cylinders and considerably lower values (0.75 d0 and even less). A simple calculation will prove that the resultant of the reaction forces of the hydrodynamic shoes 16 will be considerably out-of-center in relation to the barrel axis. It may also be proved in this case that it lies at a distance which may exceed d. /2. As a result, only one or two dynamic shoes are loaded. In the case illustrated in FIG. 2, assuming that 15 is the high-pressure port, only the shoes 16a and 16b would receive a sensitive load. This is attended by two inconveniences :
the barrel takes an inclined or skew position in relation to the distributor plate :
This inclination is attended by very reduced play on the side of shoes 16a and 16b ; in other words, the sensivity to pollution of the fluid will be considerable. On the other hand, important plays will develop on the side of the low-pressure port 14.
the fluid temperature between the shoe and the barrel increases to considerable proportions (up to 200° C temperature increments are easily observed) ; this is attended by the twofold inconvenience of abnormally heating the component elements, thus reducing their useful life while decomposing (or carbonizing) the fluid.
According to the present invention, the average diameter d3 of orifices 10 and corresponding ports 14, 15 is greater than the diameter d0 of the circle containing the centers of the cylinder orifices ; under these conditions, the resultant of the reaction forces of the dynamic shoes may be caused to become coincident with the axis of rotation of the barrel. In fact, without going so far the values of d3 may be selected to be such that the resultant passes through a point G3 located at a distance d9 of the axis of rotation, which is considerably inferior to the mean radius of the dynamic shoes.
Thus, all the dynamic shoes will share the supporting action, the clearance between the barrel and the distributor plate will remain substantially constant and the temperature gradient under the shoes having the highest load will be such that the fluid temperature will remain within reasonable and permissible limits. Consequently, the volumetric efficiency of the pump or motor is improved while reducing its sensivity to fluid pollution and increasing its useful life.
The foregoing applies to both hydraulic pumps and motors, but the operation of these machines is improved only when their rotational speed is relatively high, for the pressure distribution under the sealing surfaces of the ports is then as contemplated hereinabove. Pumps are always driven at relatively high speed and started under no-load condition. On the other hand, hydraulic motors are frequently operated at very low speeds and under maximum pressure. In this case the distribution of pressures on the sealing surfaces is not known and the consequence of the considerable friction thus developed may lead to a low starting torque and a high degree of wear.
Considering now the motor in its inoperative condition with the barrel contacting the distributor plate, we have seen that at a given moment the force urging the barrel against the distributor plate was equal to the force exerted on each piston, multiplied of course by the number of cylinders under pressure.
On the other hand, the reaction force may be equal, at the minimum, to the product of the pressure by the surface area bounded by the angle a and diameters d1 and d2, multiplied by the number of cylinders under pressure at the time considered. This pressure force may be considerably lower than that obtained when the motor revolves at high speed. According to this invention, an additional reaction effort is provided for cancelling the friction forces. Let :
m be the maximum number of cylinders under pressure ;
Sc the cross-sectional area of a cylinder (or piston) ;
Sri the surface bounded by the angle a and diameters d2 ;
the additional reaction surface contemplated will be :
Sr2 ≥ m × Sc - m × Sr1 = m (Sc - Sri)
Moreover, due to the fact that the surfaces denoted Sri are smaller than surfaces (S) (active reaction surfaces when the motor is operating at high speed), the center of gravity G1 of Sri will not be coincident with the above-defined center of gravity G2. It will lie at a distance d8 greater in relation to the axis of rotation of the barrel. To balance the barrel under these conditions it is contemplated, as the surface Sr2, a surface area such as illustrated at S2 in FIGS. 1 and 2, of which the center of gravity is at a distance d8 from the barrel axis. In this case there are two surfaces S2 since the motor can rotate in either direction, the induction and exhaust ports 14, 15 being then inverted.
With this arrangement, the barrel will safely move away from the distributor plate immediately as the motor is supplied with fluid under pressure.
However, if another additional provision were not provided, the motor could not operate normally, for when the barrel separates from the distributor plate the fluid pressure exerts a certain action under the sealing surfaces of the distribution ports, so that the hydraulic reaction would become greater than the force urging the barrel against the distributor plate ; to avoid this effect it is necessary not only to exert a pressure on the surface S2 when the barrel is inoperative, but also to discontinue the application of this pressure when the barrel is separated from the distributor plate. This is obtained in the manner illustrated in FIGS. 1 and 2, showing that each surface S2 is supplied through the corresponding induction port 14 or 15 via a pressure-loss generating laminar-flow duct consisting of a helical groove 19 machined in a cylindrical member 20 force fitted or assembled without play in a hole 21 adjacent to the fluid passage leading to said ports. Thus, when the barrel separates from the distributor plate, the fluid leakage through the outer periphery of surface S2 is such that the pressure across this surface drops rapidly to substantially zero value (for example the pressure loss may be calculated to cause the fluid pressure to be substantially zero when the barrel leaves the distributor plate by 2 to 3).
Also in this case it is proved that the fluid consumption may be reduced to practically negligible values (corresponding to less than 1/1,000 of the rated output of the motor). Finally, it will be seen that when the motor rotates at a relatively high speed, the hydrodynamic shoes disposed at the outer periphery of the barrel will take over the barrel supporting action. By means of proper calculation it will be possible, according to the desired operating speed, to determine the speed whereat the passage from hydrostatic lift to hydrodynamic lift (through the shoes) will take place. It will be seen that the pressure losses in this motor will not exceed those of a motor balanced by means of a hydrodynamic shoe, and also that this barrel balancing method is applicable to hydraulic motors of all sizes.
FIG. 4 shows another form of embodiment of this invention wherein the distributor plate 2a comprises four hydrostatic bearings S22 each supplied via a duct 23 having a separate pressure loss, said bearings being interposed between the peripheral hydrodynamic shoes 16c. With this arrangement a real adjustment of the barrel position in relation to the distributor plate can be achieved, in contrast to the preceding form of embodiment wherein the single additional reaction surface S2 definitely precluded this adjustment.
The above described device is applicable to all pumps and motors of the type comprising a multi-cylinder barrel and a flat or spheroidal distributor plate.
This arrangement is particularly advantageous in the case of hydraulic motors for which starting under full-load conditions is a frequent requirement. This is observed notably in the case of hydrostatic transmission motors of automotive vehicles (passengers vehicles, trucks and more particularly public works vehicles and agricultural machines). A specific feature characterizing the arrangement of this invention is that its component elements can be manufactured easily and economically, for example by using parts of sintered metal.
The balancing device is also applicable to the distributor plate :
of hydraulic, axial-piston pumps and motors wherein the distribution slide face is a surface of revolution of which the axis merges with the axis of rotation of the cylinder block,
of pumps and motors wherein the cylinder block held against but the distributor plate, either plane or of revolution, rotates.
This method of balancing axial forces is also applicable to the impeller or swash plate supporting the ball-joints or heads of the connecting-rods of the pistons of a pump or motor of the broken axis type, in case the axial stress transmitted through the piston rods are already absorbed by means of a hydrostatic bearing. A typical example of such impeller or swash plate is described in detail in the French Pat. application No. 69/19519 of June 12, 1969, the impeller or swash plate comprising on one face orifices registering with ports formed in a fixed slide face. These ports are supplied with a fluid under pressure. The other face of the plate comprises cavities for receiving the ball-shaped heads of the pistons-supporting connecting-rods.
This axial-force balancing method is also applicable to radial-piston pumps and motors wherein the fluid distribution takes place through the medium of a flat distributor plate. In this case the orifices such as 10 are connected to cylinders disposed at right angles to the axis of rotation of the cylinder block and the "floating" distributor plate is urged against the cylinder block by an external force derived in general from the fluid supply pressure.
Although a few forms of embodiment of this invention have been described, illustrated and suggested herein, it will readily occur to those conversant with the art that various modifications and changes may be brought thereto within the scope of the present invention, without departing from the basic principles thereof as set forth in the appended claims.