EP0653021B1 - Pair of conveyor worms for rotary positive-displacement pumps - Google Patents

Abstract

The invention relates to a pair of conveyor worms for rotary positive-displacement pumps, the conveyor worms of which, which are formed as a rotor and counter-rotor, rotate contactlessly in the drilling and the sides at the same speed (worm-drive pump), form loss gaps between them and have the same throat depth and pitch and symmetrical flank profiles on each side, made up with respect to their land height of a tooth base beneath the flank profile reversal point (q) and a tooth head thereabove. To provide an improvement it is proposed that the height of the loss gap (c) determined by the profile in the axial section on the pitch circle can be kept constant for a given rotor diameter (DK) by shifting the flank profile reversal point (q) as a function of a technically attainable rotor rise (H).

Description

The invention relates to a pair of feed screws for rotating positive displacement pumps, the feed screws of which are designed as runners and counter-rotors rotate in the bore and in the flanks in a contact-free manner at the same speed (screw pumps), form loss gaps between them and have the same pitch depth, the same number of gears and symmetrically formed flank profiles on both sides are composed with respect to their pitch from a tooth base lying below the flank profile reversal point and a tooth head lying above this flank profile reversal point.

It is known to provide the rotating, rotor-named conveying elements with screw-shaped profiles in screw pumps, which are subject to the gearing laws, but are not primarily used for power transmission but for sealing the delivery chamber between the suction and pressure chamber. In the case of the so-called external screw pumps, the rotors run in the bore and in the flanks without contact and are therefore suitable for pumping non-lubricating, viscous media, which are often permeated with dirt and solid particles.

The volumetric and the overall efficiency are essentially influenced by the viscosity of the pumped medium and harmful gaps, while the pressure height is influenced by the bearing distance, the rotor length, the pitch of the rotor, its diameter and the hub ratio Nü (tooth root diameter: tooth tip diameter).

Basically, for the runners, there is a difference between the circumferential gap (this is the gap between the rotor and the surrounding rotor bore), the basic gap (this is the gap between the rotor outer diameter of one rotor and the root diameter of the other rotor) and the gaps between the flank profiles of the rotor in the rotor engagement distinguished. In the flank gap, a distinction is made between a gap to be specified for the required contact-free running and a profile-dependent gap which arises due to the toothing laws. This profile-related loss gap is the subject of the present invention.

If the non-lubricating media to be pumped is permeated with solid particles, the backflow caused by the back pressure through all the games causes jet wear; this increases the games, so that the effective flow rate drops after a short operating time. In practice, this problem is counteracted either by a plurality of closed chambers of the rotor arranged one behind the other, ie by increasing the number of chambers, or by different speeds of the counter rotor. The first proposed solution leads to an extension of the rotor and thus to a large distance between the bearings and to a restriction of the delivery pressure due to the greater deflection. In the second alternative, the runners have different numbers of gears and therefore different speeds, which means that shorter filling times and, above all, larger viscosities prevent complete chamber filling. Another disadvantage is that if one runner has multiple gears, the smallest possible rotor pitch is in any case greater than that of the same-sized rotor pairs, since otherwise the tooth strength or gear strength would be too weak. This disadvantage also leads to a limitation of the suction height.

DE-A-1 553 271 discloses a pair of delivery screws for rotating positive displacement pumps, the delivery screws of which are designed as rotors and counter-rotators rotate in a housing bore and form loss gaps between them. The feed screw, which is referred to as the main rotor, is characterized by cylindrical webs on the tooth base, which are essentially uniform over the rotor length, by cylindrical webs on the tooth head and by inclined webs which extend over the length of an uneven section and are arcuate, the centers of curvature on the rotor axis lie. The inclined lands are widest at the exit end of the rotor. The so-called auxiliary rotor is complementary. The development of rotors with arc rolling and appropriate sealing at the foot of the aisles of the main rotor is achieved. Wide, uninterrupted sealing lines are also created.

DE-A-594 691 discloses a screw compressor which consists of right-handed and left-handed, intermeshing screws which are inevitably coupled by gearwheels and sealed by a housing. The tooth gap that widens outwards is formed by a basic trapezoid with a tip angle of about 6 ° and a trapezoid of about 12 ° tip angle that overlaps this in the outer part. The thread depth or the pitch or both change according to a geometrical series from the form a, aq ² , aq ^n-1 , where a is half the thread flank, n the number of flanks and q depends on the widening angle of the profile. This should reduce the profile and flank gaps. Depending on the choice of the flank angle, the point of intersection of the two flank angles can shift between the center and the outer edge of the profile, so that small triangular profile gaps arise at the foot and head and at the same time the flank gap is pressed to the smallest possible extent. Here, the highest possible flow rate and delivery pressure is achieved by changing the thread depth or the pitch or both depending on a geometric series of the shape of the thread flank.

GB-A-254 986 also discloses a pair of delivery screws for rotating positive-displacement pumps, each with symmetrically shaped flanks. The flank contour results from a line, the middle section of which is straight, but convex on the outside and concave on the inside of the screw, the imaginary line connecting the points where the said production line connects the outside and the inside of the cylindrical surfaces the feed screw hits, with the screw rotation axis encloses an angle between 30 ° and 75 °. This design should allow the feed screws to be installed without play and with complete sealing.

The invention has for its object to develop a pair of feed screws of the type described above with a short rotor length and a correspondingly small Bearing support spacing, with the smallest possible number of stages, large tooth tip wide, small rotor pitch and small circumferential gap length, in order to create a screw pump with a relatively large flow rate and high delivery pressure with low material consumption.

This object is achieved in that the profile-related loss gap height in the axial section on the pitch circle is kept constant for a specific rotor diameter by shifting the flank profile reversal point as a function of a technically feasible rotor pitch.

It is useful if the flank profile reversal point increases in the ratio of the rotor diameter to the rotor pitch, and if the flank profile reversal point within the series, which is characterized by the ratio of the foot diameter: head diameter = Nü, with a decreasing rotor pitch beyond the pitch circle radius up to a value of maximum r = 1 increases.

It is also advantageous if the rise of the flank profile reversal point starts from a minimum value that is greater than the pitch circle radius, and if the profile-related loss gap height to be kept constant in the axial section on the pitch circle in the range 0.1% to 1.5%, preferably 0.1 % to 0.8% of the rotor diameter.

When pumping liquids with a high gas content, a high local heat of compression often arises after a short running time, particularly on the rotor tooth that is closest to the pressure side. This can lead to local circumferential gap consumption and ultimately to seizure due to material contact (frictional engagement). According to the invention, this problem is to be remedied in that, in the case of multi-stage runners, only the stages closest to the suction side have an optimum loss gap distribution according to the invention as described above. This is in the axial direction of the rotor also compresses gas within the increased loss gap volumes; the resulting compression heat is distributed over a larger surface of the rotor and can be dissipated more cheaply. This prevents local overheating of the tooth head. With these measures, the pump for the conveyance of liquid with a high gas content of more than 95% can be designed with operational efficiency for the first time. This is achieved according to the invention in that, in the case of a multi-stage rotor, the flank profile reversal point changes continuously, stepwise or repeatedly discontinuously towards its optimal position up to the stage closest to the suction side.

The minimum flank profile reversal point is approximately 8/10 of the pitch circle radius plus 0.2. The exact calculation is based on the formula q _min = 0.6258 xe ^0.4886Nü .

The disadvantages mentioned at the beginning of the prior art are thus counteracted according to the invention in that, with the runner and driven runner, the profile-related harmful gap is divided at the same rotational speed, with suitable measures making the foot loss gap up to 24 times larger than the head loss gap depending on the gradient. As a result, a lower leakage current flows at the tooth tip, so that less wear occurs at the tooth tip. In addition, there is an improved volumetric effect, which improves the overall efficiency and leads to a longer service life of the screw pump.

In the solutions according to the invention for multi-stage runners described above, the possibility is created to allow incompressible media to flow back through gaps, while compressible media, on the other hand, is compressed over a longer path in the axial direction. Regardless of how the profile reversal point increases over the length of the rotor, the profile reversal point should always have its optimum at the suction-side rotor end and rise towards the rotor radius towards the pressure-side rotor end. This is to dissipate the local heat generation on the pressure side through the profile-related gaps over the length of the rotor. The profile-related gap is thus reduced towards the suction side. This measure results in a shift in the point of application of the transverse force even at high gas rates; it is shifted from the center on the pressure side to the support bearings located on the suction side, which reduces shaft deflection. Such pumps are particularly suitable for oil production directly at the borehole, where gas-permeable media can be expected (multi-phase production).

The invention is explained in more detail with the aid of diagrams. With regard to the characters and terms used, reference is made to the "overview table".

Figur 1a -

eine Eingriffsfläche im Stirnschnitt für q_min;

Figur 1b -

eine Darstellung gemäß Figur 1a für q_max;

Figur 2 -

ein Achsschnittprofil;

Figur 3 -

einen Achsschnittprofilspalt mit einem Vergleich für q = 1, q = m, q = q_min bei gleicher Steigung und gleichem Durchmesser;

Figur 4 -

einen Achsschnittprofilspalt für q_min und q_max;

Figur 5 -

die Verlustflächen in der Eingriffszone im Stirnschnitt dimensionslos dargestellt (R_K/R_K = 1 = r) am Zahnkopf und -fuß über den Profilumkehrpunkt mit Fuß/Kopfverhältnis Nü = 0,4 und Nü = 0,65;

Figur 6 -

ein Achsschnittprofil mit kontinuierlich veränderlichem Profilerzeugungskreis;

Figur 7 -

als Stand der Technik im Längsschnitt eine Schraubenspindelpumpe mit Außenlagerung und

Figur 8 -

in verkleinertem Maßstab die Schraubenspindelpumpe gemäß Figur 7 im Querschnitt.

Show it:

Figure 1a -

an engagement surface in the forehead incision for q _min ;

Figure 1b -

a representation according to Figure 1a for q _max ;

Figure 2 -

an axial section profile;

Figure 3 -

an axial section profile gap with a comparison for q = 1, q = m, q = q _min with the same pitch and the same diameter;

Figure 4 -

an axial section profile gap for q _min and q _max ;

Figure 5 -

the loss areas in the engagement zone in the frontal section are represented dimensionlessly (R _K / R _K = 1 = r) at the tooth head and root over the profile reversal point with foot / head ratio Nü = 0.4 and Nü = 0.65;

Figure 6 -

an axial section profile with a continuously variable profile generation circuit;

Figure 7 -

as a state of the art in longitudinal section a screw pump with external bearings and

Figure 8 -

on a reduced scale the screw pump according to Figure 7 in cross section.

The screw spindle pump shown in FIGS. 7 and 8 has, as the conveying elements, two oppositely opposing pairs of conveying screws which mesh with one another in a contactless manner and each comprise a right-handed conveying screw 1 and a left-handed conveying screw 2. The axial thrust is balanced by this two-flow arrangement. The interlocking conveyor screws, together with the housing surrounding them, form 3 individually closed delivery chambers. When rotating via a drive shaft 4, these chambers move continuously and parallel to the shafts from the suction to the pressure side. The direction of rotation of the drive shaft determines the locomotion of the delivery chambers. The pressure build-up is almost linear over the length of the conveyor elements. The medium flowing in or sucked in through the suction nozzle 5 of the pump is supplied to the two suction chambers in two partial flows in the pump housing 6.

The torque is transmitted from the drive shaft to the driven shaft by means of a gear transmission 7 arranged outside the pump housing 6, the setting of which ensures the contact-free running of the conveying elements. A stuffing box is identified by reference number 8.

The direction of rotation of the drive shaft 4 is identified by the arrow 9 in FIG. Figure 8 shows schematically the pressure port 10.

With regard to the further reference symbols, reference is made to the "overview table".

According to the conveyor screw flanks are so straight designed as possible while avoiding convex or concave shapes. The aim is to minimize the loss gap depending on the profile. The optimum is to be found between the two extreme values q = m and q = 1, the tooth tip being small in relation to the tooth height, the tooth tip width being large and the distance between the meshing tooth flanks being small. These requirements must be met consistently over a range of incline.

If q = 1, the tooth tip height is zero; the tooth tip width is greatest with H / 2, but the flank spacing as well as the tooth root loss area is greatest. This is not practical if the minimum pitch is required and the tooth strength is stable. Even at low back pressure, the backflow losses are very large due to the large foot gap; the effective flow rate is consumed.

At the other extreme q = m, the tooth tip is largest with half the tooth height and the tooth tip width is smallest; there is only one profile-related head gap. The flank distance is zero up to the center of the tooth and also increases to the maximum at the rotor diameter. The reaction forces are greatest, so that efforts should be made to arrange q far from this point.

The inventive division of head and foot gaps, taking into account the surface friction between the screw flanks at the same differential pressure, can significantly reduce the backflow loss, especially in the case of low-viscosity and high-gas media. This leads to an improvement in efficiency and less beam wear.

In the case of gas-containing media, the measure of compression heat is distributed in a targeted manner by the measure according to the invention and thereby counteracts an exhaustion of the circumferential gap on the tooth head and running noise.

Overview table

L _o :

Distance between the runner's support bearings

F _oil :

Runner length

A:

Center distance of the support bearings

H:

Runner pitch

K:

Number of runners

G:

Number of gears of the runner

S:

Tooth tip width

A _u :

Loss area on the rotor circumference and enclosing housing

A _k :

Loss area on the tooth head at the tooth mesh (in the face cut)

A _f :

Loss area at the tooth base at the tooth mesh (in the face cut)

A _V :

Sum of the loss areas of the tooth head and tooth root at the tooth mesh (in the face cut)

D _K :

Tooth tip diameter = rotor diameter (R _K = D _K / 2)

D _F :

Tooth root diameter (R _F = D _F / 2)

D _T :

Pitch circle diameter = pitch circle diameter (R _T = D _T / 2)

D _q :

Profile circle diameter (R _q = D _q / 2)

or Nü:

Hub ratio D _F / D _K

m:

Dimensionless pitch circle radius (R _T / R _K )

q:

dimensionless profile circle radius (R _q / R _K )

q _min :

minimal profile reversal point

q _max :

maximum profile reversal point

q = 1:

no head gap

q = m:

no foot gap

r = 1:

dimensionless tooth tip radius (R _K / R _K = 1)

c:

Relatively constant distance in the axial section on the pitch circle of the loss surface due to the profile

α _u :

Wrap angle of the rotor from the rotor housing

α _ö :

Opening angle of the enclosing rotor housing

α _m :

Profile angle in the face cut on the pitch circle radius

α _SK :

Profile angle in the face cut at any head radius

α _SF :

Profile angle in the face cut at any foot radius

α _FL :

Profile angle on the axis cut at the tip circle radius

β:

Profile angle in the face cut at the rotor radius

Claims (7)

1. Pair of conveyor worms for rotary positive-displacement pumps of which the conveyor worms, designed as rotor and counter-rotor, revolve at identical speeds without contact in the bore and in the flanks (screw pumps), form leakage gaps between themselves and have identical thread depths, identical numbers of starts and flank profiles which are symmetrical in design on either side and are composed, with respect to their lead, of a tooth foot located beneath the cusp point (q) of the flank profile and a tooth head located above this cusp point (q) of the flank profile respectively, characterized in that the profile-dependent leakage gap height (c) is kept constant in the axial section on the pitch circle for a specific rotor diameter (D_K) by displacement of the cusp point (q) of the flank profile as a function of an industrially viable rotor pitch (H).
2. Pair of conveyor worms according to claim 1, characterized in that the cusp point (q) of the flank profile rises in the ratio of the rotor diameter (D_K) to the rotor pitch (H).
3. Pair of conveyor worms according to claim 1 or 2, characterized in that the cusp point (q) of the flank profile increases - within the line of products characterized by the ratio of foot diameter (D_F): head diameter (D_K) = Nü - with decreasing rotor pitch (H) beyond the pitch circle radius (m) to a value (q_max) which is at most r = 1.
4. Pair of conveyor worms according to claims 1, 2 or 3, characterized in that the rise of the cusp point (q) of the flank profile begins from a minimum value which is greater than the pitch circle radius (m).
5. Pair of conveyor worms according to one of the preceding claims, characterized in that the profile-dependent leakage gap height (c), which is to be kept constant, in the axial section on the pitch circle lies in the range of o.1%. to 1.5%, preferably 0.1% to 0.8% of the rotor diameter (D_K).
6. Pair of conveyor worms according to one of the preceding claims, characterized in that, with multi-stage rotors, only the stages located closest to the suction side have optimum leakage gap distribution according to one of the preceding claims.
7. Pair of conveyor worms according to claim 6, characterized in that with a multistage rotor the cusp point (q) of the flank profile changes continuously, stepwise or often unsteadily in direction to its optimum position, up to the stage located closest to the suction side.

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