WO2003036086A1 - A linear self-adjustable compression pump of variable displacement for power transmission of infinite ratios including a variable geometry crankshaft - Google Patents

Abstract

This invention refers to the creation of piston-bearing linear pumps of a variable displacement through mechanical variable geometry crankshaft systems which also operate as an actual torque load gauge. The operational principle of these crankshaft systems, is based on one hand on the centrifugal force which acts on the mechanical swiveling parts such as the crankshaft webs (G1-G2) and the swinging buttons (47-47b) as well as the symmetric arms (72) and, on the other hand, on the presence of a load acting as a centripetal force so that all the above parts exist under a controlled swinging. Those swinging parts are under direct or indirect connection with the pistons of the pumps in question, which start to reciprocate. Nevertheless, all pumps using crankshaft webs are easily modified from pumps of constant displacement to power producing engines.

Description

A LINEAR SELF-ADJUSTABLE COMPRESSION PUMP OF

VARIABLE DISPLACEMENT FOR POWER TRANSMISSION OF

INFINITE RATIOS INCLUDING A VARIABLE GEOMETRY CRANKSHAFT.

This invention refers to a pump of high compressive capacity whose displacement and as a consequence its volumetric provision is self- adjusting upon the modj-fying passage of its pistons and always on a correspondence to any load alterations. This fact enables the pump to operate on great mechanical power through infinite linear relations even in operation conditions of decreased entrance power. It consists of the main body of the pump on a radial lay-out of the cylinders and of different crankshaft systems of a variable geometry, which differ in shape, lay-out and operation method, the differences of which result in various solutions for the manufacture of linear pumps of variable displacement. The normal type of pump used for the transmission of large amount of mechanical power is the one bearing rotating pistons which by their rods are comiected to a rotating disc based on a constant plate placed at an mclined position against the pump rotation axis, while its supply changes only when, upon a manual movement, the angular position of the mclined and not rotating plate changes. Nevertheless, any manual correction of the angle of mclination of this mclined plate does not result into a linear modification, because this method is not connected to the load.

Another type of a high compressive capacity pump in use is the piston bearing pump of multiple stages. This pump is used for gas condensation such as the atmospheric air or oxygen at high pressure bottles for several industrial or other uses. It may have three pressure stages such as low, medium, high, while the pistons of these stages differ in both their diameter and stroke.

The invention, which will be described, refers to the construction of a piston bearing pump of great sucking and compressive capacity for both hquids and gases of any density. The intention of this pump is to be able to unlimitedly transmit mechanical power through a mechanical system of a linearly adjustable stroke of its pistons or its detailed volumetric supply per umt time.

Furthermore, and under a constant input power, such a pump may handle a variable load, the level of which will be recognized by a system of variable displacement in real time as its crankshaft system also operates as a load torque gauge. So, this pneumatic mechanical system constantly adjusts the angular velocity of the axis of the final exit of the transmitted power, so that any manual intervention to the system during its operation to be unnecessary. Such a pump will have a broad use in several branches of the modern technology, starting from light industries to the aerospace industry; i.e pumping complexes of a high manometric head, tank and tankers transfusion systems, heavy hydraulic industrial macliinery of metal fMshing, manufacture of multiple use aero-compressors as well as for drilling, and oil rrrining. Additionally, it has a direct application to the hydraulic systems of vessels, airplanes and space ships. Of great interest is also the idea to create hydraulic systems for the transmission of mechanical power for auto-motion starting from light to very heavy cate illar vehicles, or light and heavy vehicle and cateφillar vehicles. Such a system of power transrnssion taking power from one engine which may be hydrauHcally modified and linearly handled doesn't need to cooperate with gear boxes and differential gears.

The invention will be presented through different alternatives accompanied by the corresponding figures.

Concise presentation of the figures of the proposed pumps A-B-C-D.

A. A Pump with a crankshaft of a gyroscopic precession modifying the volumetric capacity of the cylinders in relation to the load is depicted at figures -1- through -3-. It is based to the principle of the gyroscopic system «GARDANO» while the power transrnssion from engine (K) to crankshaft (G) which is defined as (Gi) or (G ) is succeeded by the use of a joint of variable geometry by a system of an asymmetric rotating mass (W) which by the above joint with the crankshaft the system (W-Gι,2) acts as a symmetric Watt which through the centrifugal force -F- developed by the rotation of the (W) alters the precession angle of the crankshaft, drawings - 1- through -3-.

The crankshaft (Gi) on one hand ends at a ball bearing following the figure -1-. When the ball cavity (3) belongs to the axle body (Gi) then the ball roller (4) is caulked at the area of the shell (33). Such a support consisted of a ball roller, permits the said axle to a precession movement without rotation.

B. The same operational principle is valid for the pump of type -B- bearing two inverted centrifugal crankshafts (Gi) or (G2) comiected to the symmetric rod (24) with radial slipping centrifugal rotating loads (25), as well as the use of angular displacement connections (36), figures-4-5-6-.

C. Pump bearing a crankshaft of concentric rotation which modifies its volumetric capacity through the swinging buttons as depicted at figures -7- through -23-. The engine power (K) enters the concentrally rotating crankshaft (Z) bearing constant dilated cam (E), figures -7-, -8-, -9-. A swinging button (47) is placed at each cam (E) as the one of figure -10- bearing a piercing through gate (48) which is pierced by the cam (E) of the crankshaft (Z) according to figures -13- and -14-. The side -A'- of the gate (48) due to the fact that it is bigger than the corresponding side of the cam (E), at a definite rotation of (Z) and provided the action of a centrifugal force at the system, the button (47) is swinging according to the figure -14-.

D. A pump with a crankshaft of homocentric rotation modifying the volumetric capacity of each cylinder through a system of symmetric rotating masses of variable inertia due to a centrifugal power as depicted at figures -24- to -28-.

Analysis of the pump systems of the types -A-, -B-, -C-,-D-

Type A'. Pump with a centrifugal gyroscopic crankshaft of variable radius. The power entrance appears at the axis (1) which is supported and rolled on a part of the shell (33) while extended ends to two transversal pins (2) whose intention is the support of the system of an asymmetric (W) of variable movement of inertia, as depicted at figures -1- and -2-. The second axile under the indication (Gi) characterized as crankshaft, is extended following the imaginary axis of the first one (as depicted according to figure -1-), while its support is performed in two different ways which, nevertheless, gives the axle (Gi) the possibility to a gyroscopic movement.

From one side, the crankshaft (Gi) ends to a ball articulation according to figure -1-. When the ball cavity (3) belongs to the axle body (Gi) then the ball bearing (4) is caulked at the part of the shell (33). Such a support consisting of a ball roller, permits the said axle to provide a gyroscopic movement even when it does not rotate.

The asymmetric Watt is embodied to the part (8) bearing the joints (9) and (10). The joint (9) is supported by a transversal axle (2) belonging to the power axle (1) while the joint (10) accepts the support of the jacket (11) bearing transversal pins (12). The jacket (11) is pierced through by the body of the crankshaft (Gi) being also able to an axial slipping, and to rotate around (Gi).

The system bears two parts (8) with embodied folding masses (W) as depicted at the top view plan of figure -2-.

The operational principle.

Upon rotation of the power axis (1), around the asymmetric Watt a centrifugal power -F- is created a force which through the part (8) and the shell (11) pushes the crankshaft (Gi) resulting it a precession angle and sets it to a gyroscopic movement following figure -3-, while the other part of the crankshaft is based on a ball bearing (4) which is fixed on the shell (33), figure -1-. The presence of a ball bearing at the system is simply indicative as to the understanding of the operation of a gyroscopic movement, while the needs for all motions are covered by the articulated joint of angular changes (36) depicted at figure -4-, or by the use of a angular change joint of Rzeppa type. According to the proposed solution the crankshaft (Gi) is gyroscopically swinging without being rotated, while when all gyroscopic movements are zeroed and it is aligned with the power axle (1), then on one hand the engine (K) including the axle (1) will continue to rotate while, on the other, they will be simultaneously connected to a completely still and not swinging nor rotating crankshaft (Gi), figure -1-

Crankshaft gyroscopically rotating.

A different way to support one end of the crankshaft consists of a joint through a transversal axle according to drawing -2- piercing through of the crankshaft while it becomes still at the skins of the forked part (6) ending at an axle7) which is based and rolled on a part of a box (33) of the pump body, figures -2-3-.

At its end and following its imaginary axis the crankshaft bears a splined coupling (T) cooperating with a jacket (11a) with a corresponding acceptor, whose purpose is not to permit a free rotation of the jacket unless its axial displacement so the axle (1) torque to be also transmitted to the crankshaft which due to its operational differences, it will hereinafter be characterized as crankshaft (G2). This transversal joint support to (G_) mcluding a part (6) permits (G2) to rotate as well while its other end is swinging and making an orbit around the area. The body of the part of the swinging (Gz) due to the fact that it is of a circular cross-section, accepts the seat of the ring (13a) on its surface which may be served from both its sides by axial ball bearings (16) in order not to be axially displaced. The system at a dynamic operational condition.

When the crankshaft (Gi) or (G_ is released from any centripetal influence, then, provided a dynamic operational condition of the system, any given power axle (1) rotation will create a centrifugal tension of the balance weights (W) which by the parts (8) will break the unified imaginary axis -0-A-A-O- running through the parts (1-G2-7) where a precession angle O-A will be created at the arm (G2), figure -3-.

The ring (13a) bears at its perimeter ball connections (14) by which the connecting rods are joined (15) which are also connected to the pistons (17) which are swinging at the cylinders (18) belonging to the body of the pump. The connecting rod joint method on one hand by the pistons and on the other by the accessory (13a), can be obtained by various technical methods without altering the essential characteristics of the invention, as well as the head of each cylinder of the pump is coupled by sucking and forcing conductors with the corresponding inlet and exit valves according the apphcation use.

The variable displacement. The role of each pump of a variable displacement intends to extend its effectiveness further from the capacity of conventional pumps so that both its compressive capacity and volumetric supply to be always in connection with the real load, either if it pumps liquids or compresses gases. Any given power corresponds to a given centrifugal force -F- provided by the Watt rotation.

The Watt following the part (8) acting as a first class lever modifying -F- to mechanical work, intends to adjust the radius of the crankshaft (G2) by diagrarnmatically following the load changes. a) If the resistance imposed at the piston's head (17) is the minirnum one symbolized with -P-, then the centrifugal force -F- sets the crankshaft at its maximum radius so that the pistons run their maximum stroke, figure -3- b) If the resistance imposed at the piston is the maximum one symbolized by -Pi-, then the position of the crankshaft by an equivalent -Fi- will create the minimuiri recoil of the pump pistons at every operational cycle, so that when the piston course is minimized the system torque is maximized.

Furthermore, when the load exceeds the permitted limits and becomes -Pa- then the load sets the crankshaft (G2) at a complete ahgning to the unified imaginary axis -O-A-A-O- where any piston recoil is ceased and although under a zeroed radial displacement a maximum -F- centrifugal force is created, the engine (1) cluding the part (8) of the jacket (11) and the crankshaft (G2) are continuously rotating. The crankshaft mechanism of the system -A- bearing the crankshaft (G2) may be transformed into a constant crankshaft whose schematic presentation will be identical to the variable geometry of figure -3- which is at a maximum precession angle. In this case the part (8) of the asymmetric Watt will not modify the radius but it will be embodied to the axle (1) and then the Watt (W) will react as a flywheel of a constant moment of inertia. This crankshaft further to its use with constant displacement pumps, may be also used for the apphcation of internal combustion engines for power production.

B\ A pump including a pair of inverted symmetric crankshafts with centrifugal slipping rotating loads.

This pump is based on the operational principle of type -A- except that it bears two crankshafts whose gyroscopic movement is accomplished by their connection to a rod which during its rotation sets them to a diametrically opposite symmetric lay-out, as depicted by figure -4-.

The shell (19) of the pump is divided in to two parts and when it is reassembled is supported by screws (20). At the section point, every part of the shell bears a sheath accepting the support of an external cornice (21) of each one of them via a ball bearing. The two ball bearings (21) serve the seating of the spool - section (22) which may be also transformed as an internal rotating cornice of the ball bearings. The spool (22) is drilled at its center creating a cavity (see figure -6-) which is dilated to two diametrically opposite points (FL) in order to accept the forwarded bosses of the bar (24). The spool bears a transversal passing through hole at these points where the axle (23) gets fixed; the length of which is bigger than the external perimeter of the spool (22) following figure -6-. The purpose of this axle is to accept the seat of a symmetric bar (24) which transversally penetrates it while this bar executes a partial rotation around the axle (23). The oblong parts of the bar (following figure -4-) bear two canals (28) (recesses) where the first crosses the second by 90°. These canals serve to the slipping displacing seat of a centrifugal load (25) of a disc shape. The disc-load (25) also following the imaginary axis passing through the center of its two flat surfaces, bears support pins (26) which are seated while partially rotating at the center of the parallelepiped bearings (27) whose parallelepiped surfaces are both adjacent and slipping to the corresponding parallelepiped surfaces of the canal (28) situated at the body of the bar (24), while the canal having the maximum dimension easily enters the disc body (25). A toothed cornice (29) is placed at the external perimeter of the spool (22); this cornice is connected to a toothed wheel (30) fransmitting the engine power to the spool (K), figure -4-. Such a power may also be directly transmitted axially towards the crankshafts via angular variation joints (36) according to figure -2- and figure -3- utilizing subpart (7).

At its external perimeter, the cornice (29) bears two diametrically opposite canals (31), the width of which corresponds to the diameter of the transversal axis (23), figure -6-. As the cornice is placed under compression on the perimeter of the spool, it protects on one hand the axis from any displacements and on the other hand the cornice on the spool from a possible slipping upon the engine torque transmission.

The disc-load (25) bears a hole (32) of a circular cross section with a direction from the perimeter to the center. The free end of the crankshaft (Gi) enters this hole (32), which while such coordinating accessories facilitate the disc to rotate around the crankshaft. The other end of the (non rotating) crankshaft is seated at the boss of the plug (34) sealing the side gates of the pump as it gets stabilized by the screws (35). The seating of this end on the plug is accomplished through an articulated joint of angular variations (36) perntitting an axle to execute a gyroscopic precession without being rotated, figure -6-. Every crankshaft, penetrates a ball bearing through its center whose (at a half cross-section view) the passing through hole bears a splined coupling cooperating with the key (39) being able for an easy axial displacement (on the crankshaft).

The ball bearing has a nest (38) with radial joints which through the pins (40) are put in gear with the connecting rods (41) whose other end through the pin (42) is put in gear with the piston-section (17) whose head is at the LDC of the cylinder (18).

Operation principle. When the system is still and the engine not operating, the system (Gι-24- Gi) may come at a complete ahgning, that is a single joint imaginary axis may pass through all three above mentioned essential elements.

This ahgning is accurately accomplished by a pair of pins (43) fixed at a spool which are parallel to the transversal axis (23) and symmetrically placed as to the position and distance between them (one of which is depicted at figure -5-) mtending to cease any further rotation of the bar (24) further to the one which aligns it to the two crankshafts (Gi).

At this horizontal ahgning the loads (25) are slipping to the center of the syrnmetric bar (24). When the spool starts to rotate, the centrifugal force which will appear at the system (Gι-24-Gι) through the displacing loads

(25), will create a torsional moment to the bar (24) acting clockwise and will simultaneously put the crankshafts at a gyroscopic precession, figure - 4-. The reason why this rod will clockwise movement is because the neghgible quantity of the mass weight (25a) asymmetrically placed at a point of the discs' (25) perimeter, always provides the system with the same centrifugal direction after each full ahgnment of the crankshaft assembly.

Nevertheless, such an ahgnment as the above mentioned, may arise due to the presence of an excessive load at the exit of the pump while the engine operates.

In this case, the heads of the pistons (17) become still at the middle between the lower dead center «LDC» and the upper dead center «UDC» of the stroke, while this neutral point of the pistons is characterized as a zero recoil point.

The more the precession angle of a crankshaft deviates from the urirfied imaginary axis O-A-A-O, the more the ball bearing (37), figure -4- will axially tend to be displaced on the crankshaft (Gi) and to adjust its axial correspondence to the imaginary vertical axis piercing the center of the piston including its connecting rod (41) and the nest (38) of the ball bearing. Nevertheless, from the minimum to the maximum precession angle O-A which will be created, the reacting torque of the load appearing at the head of the piston will create a component tending by the connecting rod to displace the ball bearing (37) from the vertical imaginary axle piercing the center of each piston, connecting rod and ball bearing, and this will happen to every piston when the latter operating under load conditions.

In order to avoid this, the boss on which the articulated joint of angular changes is seated (36) is modified to a defector (44) which is internally modified to a cone (45) so that the crankshaft is able to freely move (according to the half cross section view) while this defector by its front view (46) co-adjacent to the nest (38) maintains the ball bearing to an imaginary axis vertically piercing the center of the piston (17).

A different method to put into gear the connecting rods (41) with the crankshaft (Gi), is the following: in place of the ball bearing (37), to place and fix (embody) a ring to the crankshaft bearing at its perimeter articulated joints of angular variations (36) which will be joined to the connecting rods. In this case the defector (44) will no longer have an operational use.

If bar (24) is embodied with spool (22) so to cease any rotation around its support axle (23), then the pump, on one hand, is modified to a constant displacement one, while, on the other hand it is transformable to an internal combustion engine for power production.

C\ A pump with an concentric crankshaft including swinging buttons.

The crankshaft (Z) is equipped with diametrically opposite equal eccentric buttons (E) of a dilated nature mtending to drive the swinging buttons which modify the radius by a centrifugal force upon a defined direction. Following figures -7-8-9, the crankshaft is depicted at a front, top and side view, a view referring to its broader schematic condition. Following figure -10- a swinging button (47) of a circular cross section bearing a piercing through gate whose shape creates a parallelogram (48) from which passes the constant button (E) of the crankshaft according to the cut section in figure -13-.

The width -L- of the swinging button corresponds to the width of the constant button (E) while one of the two dimensions of the gate (48) and more specifically -A'- is bigger than the corresponding -A- which belongs to the button (E) of the crankshaft. This difference of dimensions between the same sides of both the constant and movable button, is the reason why upon a centrifugal force effect during the system rotation, the button (47) may alter its rotation radius and start swinging.

The crankshaft (Z) cannot have odd buttons (E) because it will not be balanced, indicatively, the crankshaft of figures -7- and -8- bears two constant buttons on which the swinging buttons are placed (47). A washer (51) is placed at the area -a- on the concentric support axles (49) while the safety locks are placed (52) at the canal (50) by which the swinging buttons are supported, figure -13-. Following figure -10- a piercing through 4sided gate (48) is asymmetrically placed with respect to the imaginary center vertically piercing the swinging button of circular cross section (47). The asymmetric position of the gate intends to concentrate a bigger load mass at a certain point of the perimeter of the button (47), figure -10-, so that the centrifugal force to acquire one-way impulse which drives the swinging button always following the centrifugal direction. Due to the fact that the asymmetric position of the gate against the imaginary center of the button, is subject to distance restrictions, the button may be increased in height while weight may be deducted from the areas (V) situated diametrically opposite to the area where -F- acts and directs the button, figure - 11-. When the crankshaft (Z) is out of order, the swinging buttons present a picture of a symmetric cylinder, following figure -13- cross section view, and when the crankshaft starts to rotate (with a zeroed load), then the swinging buttons (47) are positioned according figure -14-. Following figures -15- and -16- the swinging button (47) is depicted as positioned on the constant button (E) of the crankshaft, while being in a static condition, from the imaginary center of the crankshaft maintains a radius -Ri- to all the points of its perimeter, while in a dynamic operational condition the distance of a certain point of its perimeter from the center of its concentric axle (49) is modified so that when the system rotates it describes a circle whose radius becomes -R_~-, figure -16-.

Swinging buttons and pistons with connecting rods.

For pumps whose network passages have valves (inlet-exit), the pistons have to bear a connecting rod. Pistons having a radial lay-out as the one of those pumps, their connecting rods through a central jacket (55) are joined to the swinging button (47), following figure -17-. This sheath bears three connecting rods (53) at a perimetric lay-out, per 120°, where their ends are put into gear with, on one hand, the pistons (54) and on the other hand, the central sheath (55) through the connections (56).

In order to avoid any axial displacement of the sheath by the swinging button (47), asymmetric driving plates (58) are used, also operating as additional loads which through the screws (59) are embodied to the button (47) and co-calculated as inert load of the swinging button. So, a centrifugal force is developed -F- equal to the one demanded by each specific centrifugal swinging system, as depicted at figure -18-, half cross section view.

Swinging buttons without connecting rods and two-way passage network without valves.

Figure -19-, presents a radial pump with two pistons (it is an indicative number) cooperating with a swinging button (47) adjacent through slipping pinions. It also a the two-way passage oil network A-B without inlet - exit valves comiected to the distributor (61) while it moves from the extension of the concentric axle (49).

The channels -A - B'- and -B - A- connect the heads (H) of the cylinders with the distributor (61) whose exit is connected to the hydraulic engine (62) by channel (I). The hydrautic engine may be either piston bearing or with a rotor gearing the axle (63) of final exit of the power. The intension of the distributor is to supply a continuous flow of hydraulic power to the engine (62) following the same mode of distribution of continuous return flow to the pump cylinders, so the distributor activates die proper networks at the proper time.

Regardless of the type of hydrauhc engine which will be used by the system in order to supply mechanical power to the moving wheels or the caterpillars of a vehicle, it is important that each variation of the external load will be handled by a linear self-adjustment of the volumetric supply of die pump to the hydrauhc engine. This adjustive capacity will permit the axis of the pump to rotate with a constant power regardless of the load level handled by the pump, including the case where an excessive load makes the swinging buttons out of order. In case of an excessive load (such as the fijll restriction of channel I), the heads of the pistons will become still at the half of their passage which is also the "point of zero recoil" while the swinging buttons will come to a symmetric radius -Ri-, following figure - 20-. The swinging button (47) as a free centrifugal mass when placed at a zero rotation radius under a maximum power, transfers the maximum -F- centrifugal force.

Operational principle.

The power of the engine (K) enters the axis (49) bearing constant dilated buttons (E) on which the swinging buttons are seated (47). Upon t e exercise of the centrifugal power, the swinging buttons tend to have an operational radius which will be a function of the load and the engine power.

The buttons (47), at their external perimeter, bear a driving canal (65) within which a wheel (66) both enters and rolls carried by every piston of the pump supported by the pin (67). When the hydrauhc liquid displacement period is completed by the piston (64), see figure -19-, the course followed by the liquid is -A-B'- which, through both the distributor and channel (I), the liquid entered the hydrauhc engine (62), rotated the axis (63), exited from the conductor -D- and entered the next piston of the pump by the distributor followed the course - A'-B-. The hydrauhc liquid pushed the second piston from the «UDC» to the «LDC» ma taining the original pressure from the compression to the refilling of this second cylinder while the pistons, due to a lack of connecting rods, will not result into sucking. When die lower piston starts to compress, the distributor will reverse the operation of its gates permitting a reverse of the flow with a course -B-A'-I-D-B'-A-. So, this results into a constant supply of the hydrauhc engine (62) with hydrauhc power. This pump is characterized as a "closed circuit" or "closed container". Conclusion:

When the load is decreased at a variable volume pump, its volumetric supply is increased (as to the hydrauhc engine whose revolutions are increased per unit time), by an inversely corresponding decrease of the manometric pressure of the pump as well as a decrease of the the hydraulic engine torque. Furthermore, when the load of the pump is increased, its volumetric supply is decreased (up to minimal recoil of its pistons), as well as the revolutions of the hydraulic engine with an inversely corresponding increase of the manometric pressure of the pump from the maximum centrifugal force -F- it results into an increase of the hydrauhc engine torque.

Filling of cylinders by an hydrauhc pressure of pistons and a lubrication system of adjacent parts of pistons and cylinders. Such a pump at a cross section view and a swinging button on a centrifugal development, is depicted at figure -21-. While the number of the pistons is simply indicative, the latter are presented at different operational phases.

Following a rotation direction from left to right of both the constant button (E) and the swinging one (47), the piston of the cylinder -Ai-, by displacing the hydrauhc liquid, arrives at «UDC». The piston of the cylinder -Bi- follows which has already displaced part of the corresponding volume. The overall displaced liquid from both cylinders as it completes its cycle and exits from the engine (62) (see figure -19-) through a distributor and an original manometric pressure is driven to the pistons of the cylinders while (due to a lack of connecting rod) they did not have another way to follow the swinging button.

As the backward and forward motion of the pistons is developed by the centrifugal displacement of the swinging buttons deviating from the imaginary center of their rotation axis figure -21-, -Pi- originates which is analyzed into two components -N- and -T- having both the sense and direction of the button rotation (47) resulting into an asymmetric slipping friction of both the cylinder sleeves and piston. In order to decrease the slipping level, by using a channel (38) ending at a cavity (69) at the perimeter of the piston (64) (a cross section view of a piston) oil is supphed from the area of the cylinder (dirring the compressive period) so to decrease the slipping friction. This lubrication provided under pressure is maintained at the cavity because it exists between the 0-rings (70) of the pistons.

Such a pump may bears several piston subsystems of radial lay-out, placed among them at different degree of mclination dividing the circle to smaller parts. This is presented by the depiction of the fifth piston following drawing (21) with dotted lines and placed under an inclination of 45°. In this case the swinging button is properly modified in order to also include the second driving canal such is the button (47a) of figure -22-.

Finally, every button (E) may bear a sheath (X) of a spring (S) whose other end will exercise on to the button (47) a force so that the spring will operate as a damper of radial oscillations, following figure -12-.

D'. Pump with a concentric crankshaft of symmetric rotating masses of variable inertia.

A crankshaft (71) is depicted at figure -24- which concentrically rotates at the seat of the box (33). It bears a transversal passing through hub (76) which by the pin (73) a concentric branch is fixed (72) at the ends of which there exists the load-wheels (74) supported by a pin (75) as also depicted at a side view following figure -25-. When the axis (71) accepts the power of the engine and starts to rotate, the loads (74) through the centrifugal force - FW- displace the arm (72) from the position -A- to the position -B-. The result of such a swinging may be apphed either in cooperation with the radial correcting rod following the one of figure -17-, or by a swinging button following figure -10-.

If the axle (71) cooperates with the swinging button, then the axle and the button have a special design so that the anticipated result to be accomphshed as it happens with the crankshaft (71a) and the swinging button (47b) depicted at figures -26- through -28-. Figure -26- shows the crankshaft (71a) at a top view bearing the gate (77) at the center of which a symmetric crankshaft is fixed (72).

At the ends of the crankshaft the moving balance weights (78) carrying the cylindrical ends (79) are both seated and slipping. The two parallel external surfaces of the framework (81) enter the gate (82) of the swinging button and prevent it from side centrifugal displacement. Figure -27- depicts the swinging button at a cross section view bearing a gate having two cavities (80) to which enter the ends (79) of the balance weights. The swinging button bears at its perimeter a canal on which the piston pinions roll (66), followmg figure -28-.

When the crankshaft (71a) rotates, a centrifugal force -F- is developed at the arm linearly pushing the swinging button. By the slipping balance weight (78) the arm (72) may describe a circular arc independent of the vertical course maintained by the button (47b). Every crankshaft of this type serves two swinging buttons from which every button may cooperate with several pistons in a radial lay-out. This type of crankshaft further to its cooperation with the swinging button may also cooperate with the multi- connecting rod of figure -17- having a high degree of prompt response as to any load variations.

Claims

1. A linear self-adjustable compression pump of variable displacement for power transmission of infinite ratios including a variable geometry crankshaft consisting of reciprocating pistons in a radial layout, coupled with one of the proposed -A-B-C-D- crankshaft systems whose mechanical effects also operate as a pump load torque gauge it is being characterized by the fact that the crankshaft system -A- consists of a supported power axle (1) - which is rolled on the shell (33) of the pump where the other end of this axle bears transversal pins (2) as supports of an mclined support of the parts (8) bearing an asymmetric Watt (W) where the ends of these parts are coupled with the transversal pins (12) of the jacket (11) or (11a) where this jacket is pierced through by the end of a crankshaft (Gi) the other end of which is formed in a sheath (3) of the ball-bearing (4) which is embodied to the shell (33) of the pump, and from the fact that through the participation of the asymmetric Watt on a joint couple with one crankshaft web under a defined variable geometry as the one of the system -A-, a crankshaft system mechanism is created for every load condition in a constant operation condition, and if the crankshaft is of the (G2) type, then its other end is coupled by a forked part (6) ending at an axle (7) seated at a shell (33) from which system -A- the basic accessories (1- Gι-4) or (1- G2- 7) at a still condition are pierced through by a unified imaginary axis through points O-A-A-O, while on a dynamic operation condition where Watt creates a centrifugal force -F- which may create a radial displacement, then from the fact the part (8) which, by the radial displacement will push the end of the controlled (by the Watt) swinging of the crankshaft, will change the geometry of the system through the angular displacement from the imaginary axis from O-A-A-0 to O-A-O by placing the crankshaft in a gyroscopic movement with a precession angle O-A or in a rotation around its axis when the end of the controlled swinging of the crankshaft bears the key (T) according to the type (G2) mcluding the jacket (11a) supported by the corresponding splined coupling and that the crankshafts (Gi) and (G2) bear, the first one an embodied jacket (13), the second one the rolling jacket (13a), a rolling jacket which is being guided by axial baU-bearings (16) at the perimeter of aforementioned jackets there exist either ballbearings (14) or the articulated joints of the angular displacements (36) joining the connecting rods (15) whose other end join the pistons (17) through connecting links of angular displacements, and if the part (8) of the crankshaft (G2) is embodied to the axle (1) having the maximum precession angle where the Watt (W) will operate as a flywheel of a constant angular inertia, then furthermore apart from its use as a pump of constant displacement, it may also contribute to the creation of an internal combustion engine for power production, and if the pump uses the crankshaft system of -B- type, then it is characterized by a pair of crankshafts (Gi) or (G2) on a symmetric layout and covered by a passing through spool (22) which externally bears a toothed wheel (29) with the purpose of power input to the spool which is supported and rolled at the ball-bearings (R) whose constant cornice (21) is found at the jacket of the divided shell (19), and by the fact that the spool across its diameter is pierced by the axle (23) on which the symmetric (24) bar is supported executing a partial rotation pushed by the centrifugal force where the oblong parts of this rod bear two canals (28), each one of which crosses the other canal at 90°, and which facilitate the shding displacing seating of the centrifugal loads (25) including the asymmetric mass (25a) while these loads at a spot of their perimeter and upon a course to the center, bear a hole (32) of a circular cross-section within which the end of the controlled swinging of each crankshaft is based (Gi) the other end of which is supported to the hub of the side plug (34) of the pump through a universal coupling of angular displacements (36) being used as an alternative power input whereas plug (34) ends at the separator (44), and that each crankshaft at a certain position of its body bears the splined coupling (39) which accepts the bah-bearing (37) mcluding its jacket (38) bearing radial hinges which by the pins (40) join the connecting rods (41) whose other end is connected through the pins (42) and is connected with the pistons (17) of this crankshaft system -B-, including the fact that it simultaneously makes a symmetric gyroscopic precession angle to both crankshafts, provided that each hydrauhc or mechanical effect which tries to appear to each one of them separately, through the rod (24) is syrnmetrically transmitted to the whole system, and from the fact that the crankshaft system -B- simultaneously operating as mechanical power and load torque gauge, continuously adjusts the precession angle of the crankshafts and as a consequence the course of the pistons related to the angular velocity of, on one hand, the engine and on the other hand, the exit load torque as it also happens with the crankshaft system -A-, and if the rod (24) with the spool (22) at the crankshaft system -B- gets caulked at a position such that it may direct the crankshafts (Gi) to a constant precession, then due to the fact that the pump is transformed to, on one hand, a constant displacement, and, on the other hand, to an internal combustion engine for power production which through rotation exits through a toothed cornice (29) and moves the toothed wheel (30), and if the pump uses the crankshaft type system -C-, then it consists of a crankshaft (C) concentrically rotating whose support ends (49) are rolled on the parts (33) of the pump shell and which crankshaft is modified to constant dilated eccentric buttons (E) which pierce through the gate (48) bearing at its flat surfaces every swinging button (47) of a circular cross section, and due to the fact that the side -A- of the gate (48) is bigger than the corresponding side of the constant button (E), it is permitted to every swinging button to move radially at a predefined direction by the effect of a centrifugal force, and because the button bears a canal (65) at its external perimeter within which a pinion (66) enters which through the support pin (67) is connected with a piston (64), and due to the fact that the piston at a point of its perimeter and among the O-rings (70) bears a cavity (69) which through a hydrauhc channel (68) ending at the piston head, communicates with the cylinder filling volume, and through this small cross section provides lubricants among the friction parts at a certain area of the cylinder's perimeter which is asymmetrically strained by the appearance of a resultant force -Pi- arising from the system rotation which is analyzed into two components - N- and -T- at both the sense and direction of the rotation of the crankshaft (C) including the swinging buttons (47), and because the swinging button may increase in height by accepting more canals (65) so as to be able to push more pistons at a complete rotation as the button (47a) and because every constant button (E) may bear a sheath (X) within which enters the spring (S) whose other end will exercise to the button (47) a force with a centrifugal direction which will operate as a damper of radial osciUations, and from the fact tiiat all pistons without connecting roads (64) which are only moved by the swmging buttons by the use of an appropriate distributor (61) create a closed circuit pump for a linear movement of hydrauhc engines, a pump whose piston withdrawal of on a downward direction is executed by the constant manometric pressure of a closed circuit network, and due to the fact that, by using an asteroid multi- connecting rod (55-53) connected on one hand by the swinging button and on the other by the pistons, the pump of a closed container is modified to a sucking pump of an opened container, and because this crank system bearing a swinging button at a joiriing with a multi-connecting rod, also operates as a power and load torque gauge and, and if the pump uses a crankshaft system of type -D- then it consists of a crankshaft axle (71) concentrically rotating on the shell (33) bearing a transversal pierced hub (76) at which a symmetric double faced arm is supported (72) at the ends of which exist centrifugal loads (74) under the form of small wheels which are supported by pins (75), or by a crankshaft axle (71a) and the appropriate shape of the ends of the arms (72) so as to be able to accept diametricaUy passing through balance weights (78) of a shding variable radius bearing the pins (79) entering the cavities (80) of the swinging buttons (47b) bearing the gate (82) at which the crankshaft (71a) projection (81) enters providing the buttons a vertical position with respect to the crankshaft axle for the avoidance of side centrifugal displacements regardless of the change of the radius in which the button is subjected to due to load changes, and because this crankshaft system operates as a torque gauge.

2. A linear self-adjustable compression pump of variable displacement for power transmission of infinite ratios including a variable geometry crankshaft bearing a pair of crankshafts (Gi) where every crankshaft bears a ball bearing (37) through key (39) and the corresponding splined coupling according to claim -1-, and is characterized by the fact that the use of the key including the splined coupling has as an objective to, on one hand, cease any trend to a local rotation of the roller on the crankshaft and, on the other hand, and to every change of the precession angle occurred to the crankshaft web, permit the axial displacement of the ball bearing (37) over the crankshaft so that the former adjusts its position by following a vertical imaginary axis piercing through the center of the piston and the connecting rod.

3. A linear self-adjustable compression pump of variable displacement for power transmission of infinite ratios including a variable geometry crankshaft bearing a pair of crankshafts (Gi) each one of them bearing a ball bearing (37) which may be axially (adjustably) displaced on the crankshaft (Gi) at every alteration of its precession angle so that the ball bearing is always placed at the center of the imaginary axis piercing through the piston and the connecting rod according to the claims -1- and - 2-, and is characterized by the fact that the measure which defines the l nits of the displacement of the ball bearing including its sheath (38) is the constant separator (44) whose face (46) as it adjoins with the one side of the jacket (38), it also defines the limit of stoppage of any firrther displacement of the roller whose resultant force causing its displacement, has always the same direction.

4. A linear self-adjustable compression pump of variable displacement for power transmission of infinite ratios mcluding a variable geometry crankshaft bearing a pair of crankshaft webs (Gi) each one of them bearing a ball bearing (37) mcluding its sheath (38) where the roller displacement limit on the crankshaft is defined by the separator (44) according to the claims -1-2- and -3-, and is characterized by the fact that the ball-bearing (37) is replaced by an embodied jacket including the crankshaft which all over its perimeter bears articulated joints for angular displacements (36) for the joint to the connecting rods, the other end of which bears joints of the same type for the connection to the pump pistons, and fonowing this method the separator can be placed out of order (44).

5. A linear self-adjustable compression pump of variable displacement for power ttansrnission of infinite ratios including a variable geometry crankshaft bearing a crankshaft with swinging buttons joined to an asteroid multi-connecting rod according to the claim -1-, and is characterized by a sheath (55) at the internal part of which the swinging button (47) is rolling at the flat surfaces of which the asymmetric driving tiles are embodied (58) also operating as complementary centrifugal loads, and because the sheath (55) bears at its perimeter the connecting rods (53) which by the joints (56) are also connected by the pistons (54).

6. A linear self-adjustable compression pump of variable displacement for power transmission of nifinite ratios including a variable geometry crankshaft consisted of swinging pistons in a radial lay-out connected to a crankshaft system of the proposed -A-B-C-D- systems according to the claim -1- and is characterized from the fact that the displacement of the pistons within the cylinders is related to the centrifugal force acting either on the ends of the controlled swing of the crankshaft (Gι)-(G2) of the systems -A-B- which attain under a gyroscopic precession a swinging movement, or on the swinging buttons and the symmetric arms of the crankshaft systems -C-D- thus contributing to a crank rotation of the swinging buttons.

7. A linear self-adjustable compression pump of variable displacement for power transmission of infinite ratios including a variable geometry crankshaft consisted swinging pistons in a radial lay-out connected to a crankshaft system of the proposed -A-B-C-D- ones also operating as a torque gauge of the pump load according to the claim -1-, and is characterized by the fact that each mechanical crankshaft device of the above systems is a function of an active centrifugal force -F- deriving from the engine power, and of a centripetal -C- reaction deriving from the load torque, where the action of these two contrary forces at the mechanical parts of the above systems such as the (Gι-G2-47-72), set those systems to a controlled swinging while, due to this geometrical change become active crankshaft mechanisms with the capacity to mechanical adjustments as to any load changes as well as active mechanical load torque gauges for the load variations in real time including the point of a self-disengaging of every crankshaft system while the engine operates.