US3535062A

US3535062A - Vane pumps

Info

Publication number: US3535062A
Application number: US761084A
Authority: US
Inventors: Raymond Jagger
Original assignee: TOWLER HYDRAULICS Ltd
Current assignee: TOWLER HYDRAULICS Ltd
Priority date: 1967-09-21
Filing date: 1968-09-20
Publication date: 1970-10-20
Anticipated expiration: 1987-10-20
Also published as: GB1246084A; FR1584447A; DE1728279A1

Description

Oct. 20, 1970 R. JAGGER VANE PUMPS 3 Sheets-Sheet 1 Filed Sept. 20, 1968 2 Q E F 0 4 P w J A mm w y W M. 4. G 1 H M s m a A 4 w 3 W m m A R i WV M 2 0 v4 .3 7/ F .DI

Oct. 20, 1970 R. JAGGER $535,062

' VANE PUMPS Filed Sept. 20, 1968 3 Sheets-Sheet 2 x r///// R 1510 14 F/G.5. 10 21,251 1 7812 3 20 l I" I & Q

22 @1' J 74 was luvsm'oR Rmmquo JAGGER A'rrvs.

Oct. 20, 1970 R. JAGGER 3,535,062

VANE PUMPS Filed Sept. 20, 1968 3 Sheets-Sheet 5 INVENTOR RAYMOND JAGGER 17 5 44, JM M ATTYS.

United States Patent M U.S. Cl. 418-268 6 Claims ABSTRACT OF THE DISCLOSURE A rotary vane pump or motor in which the outer end of each pump vane is formed with a longitudinally extending groove which separates lands on opposite edges of the outer end. The pressure drop across the outer end of the vane occurs in two stages, across the two lands, while pressure is substantially uniform in the groove. A bleed passage is provided in each vane for transferring pressure from the high pressure or leading side of the vane to the groove, thereby to ensure that the pressure in the groove is greater than the mean of the pressures acting on opposite land edges. Passages also are provided for transmitting the pressure acting in the groove to the inner end of the vane so that the groove pressure acts across the entire inner end of the vane to provide a resultant force acting radially outwards.

The invention relates to rotary pumps of the kind comprising a rotor having sliding vanes working in cooperation with a cam surface exterior to the rotor to displace fluids under pressure between the inlet and discharge sides of the pump.

It is a requirement of such pumps that the vanes should maintain contact with the cam ring at all times during operation so that the pump runs quietly and also that the forces between the cam ring and the vanes are minimal so as to minimise wear on the cam ring and vanes.

In rotary pumps centrifugal force is not in itself sufficient to maintain the vane tips in contact with the cam ring surface at all times and any device included to ensure, and maintain, contact must be adapted to operate irrespective of any speed variation or pressure variations of the pump.

In a complete cycle of operation there are four parts or stages of the cam ring surface which present different conditions and therefore require different conditions of the operation of the vanes:

1) The ramp between minimum cam ring diameter and maximum cam ring diameter over which the vanes move radially outwardly and the spaces between the vanes increase in volume and are filled with fluid from the suction port of the pump; this position is identified as the suction ramp.

(2) The circular arc portion of maximum cam ring diameter from the end of the suction ramp to the beginning of the discharge ramp (as (3) below).

(3) The ramp between maximum cam ring diameter and minimum cam ring diameter over which the vanes move radially inwardly and the spaces between the vanes reduce in volume thus expelling fluid to the discharge port, this position is identified as the discharge ramp.

(4) The circular arc portion of minimum cam ring di ameter from the end of the discharge ramp to the start of the suction ramp.

The force required to maintain the vanes in contact during (1) (the suction ramp) can be readily provided by means of compression springs inserted in the vanes and reacting against the bottom of the rotor slots. Such springs are cheap, easy to install and can be provided so as to be effective to cover a wide range of rotational 3,535,052 Patented Oct. 20, 1970 speeds and irrespective of discharge pressure variations.

In stages (2), and (4) however hydraulic pressure generated in the pump may upset the hydraulic balance of the vanes due to (a) wear of the vanes and/or cam ring or (b) the disposition of the vanes in the rotor slot, as, for example, if the vane tips in the rotor slot and is not presented accurately to the cam ring are.

It is among the objects of the invention to provide means whereby contact of the vanes with the cam ring is maintained.

According to the invention in a rotary pump of the kind described spring members are provided to exert an outwardly directed radial force on the vanes and means are provided for hydraulically augmenting the outwardly directed radial force of the springs over those portions of the cam ring in which the hydraulic balance may be upset.

Thus the vanes may be formed with a bleed across the leading edge of the vane, the disposition of the bleed being such that it is sealed off at least when the ring is in contact with the portion of the cam ring of minimum diameter.

The invention is illustrated by way of example in the accompanying diagrammatic drawings in which:

FIG. 1 is a pressure distribution force diagram for stage (2) operation of a pump of known type;

FIGS. 2 and 2a show alternative constructions for providing a bleed across a leading edge of a vane;

FIG. 3 shows a vane in a retracted position with the bleed sealed offi;

FIG. 4 is a part plan and part longitudinal cross sectional view of a vane pump;

FIG. 5 is a crosssection along the line II in FIG. 4; and

FIGS. 6 and 7 are respectively sections along the line II-II in FIG. 5 showing the arrangement of the pump for clockwise and anticlockwise rotation.

With reference first to FIGS. 4 to 7, a vane pump comprises an outer casing 10 containing a body member 12 which latter is formed in three portions. A first portion 14 is located at the left hand end of the casing interior as viewed in FIGS. 4 and 5 and abuts against abutment members 16 integral with the casing. A second portion of the body member is located at the right hand end of the casing interior as viewed in FIGS. 4 and 5. The second portion is formed with a neck 13 which is received within an annular collar 20 integral with the casing 19. The third portion of the body member is carried by the second portion. The body member divides the interior of the casing into the inlet and outlet side of the pump. The inlet or suction port is indicated by the numeral 21 and the outlet or pressure port is indicated by the numeral 23.

The first and second portions of the body member are centrally bored to receive a shaft 22 carrying a rotor 24. The rotor is splined to the shaft and is situated between the first and second body member portions. The shaft is supported by a ball bearing assembly 26 and the inwardly facing side of the bearing assembly is sealed by a seal means 28 which is urged against the bearing assembly by a spring 30 housed in a counterbored portion of the second body member portion.

A cam ring 32 is arranged about the rotor 24. A plu rality of radial vanes 34 are mounted in slots 36 arranged about the circumference of the rotor. In the illustrated example the rotor carried ten vanes, each vane extending the full width of the rotor. The vanes are urged radially outwards into engagement with the cam ring by springs 38. Three individual springs are associated with each vane. During a half revolution of the rotor 24, a vane moves over four stages of the cam ring, namely a suction ramp at, an arcuate path being the maximum cam ring diameter {3, a delivery ramp 7, and a second arcuate path being the minimum cam ring diameter 6.

Each vane is formed with a groove 4-0 extending radial- 1y along the side faces thereof bearing against the first and second body member portions and transversely therebetween across the outer face of the vane bearing against the cam ring. The leading face of each vane as determined by the direction of rotation of the rotor is provided with a bleed slot or bore 42 which communicates with the peripheral groove 40.

The conditions affecting operation of the pump in the four stages referred to above are as follows:

(1) On the suction ramp (at in FIGS. 6 and 7) a force must be provided to urge a vane to move radially outwards. The construction of the pump is such that the top and bottom of the vane is connected to the suction port over this portion and by virtue of this the vane is in hydraulic balance radially. The force required cannot be provided hydraulically from the discharge port because the pressure in the discharge port can be substantially zero when the pump is off-loaded and it would be necessary to ensure that some pressure remained at the discharge port (by restricting free flow discharge to tank with consequent power wastage) to enable any type of hydraulic piston disposed beneath or within the vane to be effective.

Compression springs 38 are therefore inserted in the vanes reacting against the bottom of the rotor slots to provide the necessary force to cover a wide range of rotational speeds and irrespective of discharge pressure conditions.

(2) Over the maximum cam ring diameter ([3 in FIGS. 6 and 7) the vane must prevent excessive leakage of fluid from the pressure zone it is approaching (i.e., on its leading face) to the suction zone it is leaving (i.e., on its trailing face). Fluid can leak across the tips of the vane where they are being urged into contact with the cam ring and across the side of the vane which projects out of the rotor. The spaces at the top of the vane and the bottom of the vane are at this stage not in direct contact with the suction and discharge ports and measures must be taken to ensure that the pressure build up over the width of the vane does not unbalance the vane hydraulically. To accomplish this the vane is provided with two equal width edges or lands which effect the sealing along the cam surface. Groove 40 communicates with the space below the vane. If the vane has a sharp edge along its top face, the vane leading and trailing edge widths are equal and the vane is presented truly radially to the ring are then pressure in the vane groove is substantially half the pressure on its leading face and the vane is substantially in hydraulic balance radially, the force between vane and cam ring surface being that due to springs and centrifugal force as there is no radial ac celeration of the vane over this cam are. In practice however the vane edge is not sharp and can, due to wear become substantially chamfered, the chamfer not necessarily being uniform along the edge. The vane can also tip in the rotor slot so that it is not presented truly radially to its arc ring. The redistribution of the pressure forces on top and bottom of the vane can, in these cases, give rise to an unbalanced hydraulic force tending to push the vane radially inwards and away from the cam surface with consequent noise and increased leakage.

For a given width of vane the amount of radial outward force that can be obtained from springs is limited and if more force is required the vane width must be increased to house larger springs or the radius of rotation of the vanes must be increased to obtain greater centrifugal force. Both methods involve an increase in the physical size of the pump.

As the radially inward unbalanced hydraulic force set up as described above is directly proportional to the pressure on the leading face of the vane, viz., the discharge pressure, if this pressure is used to provide extra force radially outwards, the force thus obtained will always be great enough irrespective of discharge pressure.

As indicated in FIG. 1 if the pressure in the vane groove is half the pressure on the leading face of the vane, the sealing edges are the same width, and the suc tion pressure is zero, the vane is in hydraulic balance radially if the vane is assumed to be a true geometrically rectangular vane.

If the pressure in the vane groove exceeds /2p there is a net force radially outwards equal to:

where t width of vane sealing edges p pressure on vane leading face r ratio of pressure inside vane groove to pressure on vane leading face.

The pressure in the vane groove 40 can be increased by reducing the resistance to flow across the vane leading edge. This is achieved by providing a bleed 42 across the edge. As shown in FIG. 2b a bore through the leading face of the vane communicates with the vane groove. Alternatively, in FIG. 2a a notch 44 of any desired shape is provided across the leading edge of the vane. The size of the bleed is controllable and any desired pressure greater than /2p can be obtained at will. The bleed size is selected to obtain the minimum outward force to maintain contact with due regard to unbalanced pressure forces and friction of the vane in the rotor slots and keep wear on the cam ring and vane to a minimum.

(3) On the delivery ramp (7 in FIGS. 6 and 7) the vane is being urged radially inwards by the cam ring, the top and bottom of the vane are open to discharge pressure and the vane is in radial hydraulic balance the bleed being ineffective. The force between vane and ring is now the sum of spring force and centrifugal force and the vane inertia force brought into play by the ring accelerating the vane radially inwards and this force will maintain the vane in contact with the ring.

(4) On the ring circular arc portion of minimum diameter (6 in FIGS. 6 and 7) discharge pressure is now on the trailing face of the vane and if the resistance to flow of the leading edge of the vane is less than that of the trailing edge, pressure in the vane groove will be less than /2p and there will be a net force on the top of the vane to push the vane radially inwards away from the cam ring. Consequently the bleed introduced to accommodate stage (2) may now be removed and this is effected as shown in FIG. 3 by positioning the bleed in such a position on the side of the vane or in its leading face that it is closed by entering the rotor slot Whenever the vane is on the discharge ramp. As the vane protrudes little beyond the outside diameter of the rotor, pressure force on the trailing edge is small and consequent friction of the vane opposing radial movement in its rotor slot is minimal. As the pressure in the vane groove is now /2 p due to the bleed being cut oif the vane is in radial hydraulic balance and the forces urging its radially outwards are the spring force and centrifugal force and this combination is sufficient to maintain contact between vane and cam ring.

I claim:

1. A fluid pressure energy translating device comprising a housing, a cam wall within said housing, a rotor eccentrically supported in said housing for rotation, a plurality of radially slidable vanes carried by said rotor, said vanes each having an inner end disposed within said rotor and an outer end Cooperating with said cam wall, said outer end of each vane being formed with lands extending longitudinally along opposite edges and being separated by a groove, a bleed passage formed in each vane for communicating pressure acting at the leading face of said vane to said groove, and means for transmitting pressure acting in said groove to the inner end of said vane so that said groove pressure acts across the entire width of the inner vane end.

2. The fluid pressure energy translating device of claim 1 in which each groove extends longitudinally the entire length of the outer vane end.

3. The fluid energy translating device of claim 1 in which each said means for transmitting pressure from said groove to said inner vane end is a groove formed in the longitudinal end of the vane extending between the outer and inner ends.

4. The fluid energy translating device of claim 1 in which each said bleed passage is a bore extending through the leading face of the vane and communicating with said groove.

5. The fluid energy translating device of claim 4 in which said bleed passage is sealed off by said rotor over portions of the travel of said vane.

6. The fluid energy translating device of claim 1 in which said bleed passage is a notch formed in the land at the leading face of said vane.

References Cited UNITED STATES PATENTS 1,366,139 1/1921 Traudt. 1,805,063 5/1931 Wrona. 2,632,398 3/1953 Ferris. 2,688,924 9/1954 Links 103136 X 3,014,431 12/1961 Van Den Bussche. 3,291,384 12/1966 Garland et a1 230152 3,451,346 6/1969 Pettibone 103-136 ROBERT M. WALKER, Primary Examiner W. I. KRAUSS, Assistant Examiner US. Cl. X.R. 123-8; 230-152