CA1141232A - Vane pump - Google Patents
Vane pumpInfo
- Publication number
- CA1141232A CA1141232A CA000347916A CA347916A CA1141232A CA 1141232 A CA1141232 A CA 1141232A CA 000347916 A CA000347916 A CA 000347916A CA 347916 A CA347916 A CA 347916A CA 1141232 A CA1141232 A CA 1141232A
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- CA
- Canada
- Prior art keywords
- wheel
- flow
- pump
- suction
- entry
- Prior art date
- Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
- Expired
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Classifications
-
- F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
- F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
- F04D—NON-POSITIVE-DISPLACEMENT PUMPS
- F04D29/00—Details, component parts, or accessories
- F04D29/18—Rotors
- F04D29/22—Rotors specially for centrifugal pumps
- F04D29/2261—Rotors specially for centrifugal pumps with special measures
- F04D29/2277—Rotors specially for centrifugal pumps with special measures for increasing NPSH or dealing with liquids near boiling-point
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- Engineering & Computer Science (AREA)
- Mechanical Engineering (AREA)
- General Engineering & Computer Science (AREA)
- Structures Of Non-Positive Displacement Pumps (AREA)
- Nitrogen And Oxygen Or Sulfur-Condensed Heterocyclic Ring Systems (AREA)
Abstract
"VANE PUMP"
ABSTRACT
A vane pump is disclosed.
The pump comprises a housing in which are installed with a radial clearance an axial-flow suction wheel and an axial--flow impeller wheel, which wheels are mounted on a common drive shaft one after the other in the direction of flow. The inside diameter of the pump housing in the zone of the suction wheel decreases in the direction of flow. The inside diameter of the pump housing at the entry to the sucti on wheel is cal-culated from the formula:
The suction wheel comprises a hub with helical vanes at-tached thereto. The vane setting angle increases in the direc-tion of flow and at the entry to the suction wheel it is:
where D0 = inside diameter of the pump housing at the entry to the suction wheel;
D1 = inside diameter of the pump housing at the entry to the impeller wheel;
D1 = dimensionless coefficient of 0.17 to 0.13;
Ck = predetermined cavitation critical speed coeffi-cient of 5.000 to 11.000;
.DELTA. = radial clearance.
Figure 1.
ABSTRACT
A vane pump is disclosed.
The pump comprises a housing in which are installed with a radial clearance an axial-flow suction wheel and an axial--flow impeller wheel, which wheels are mounted on a common drive shaft one after the other in the direction of flow. The inside diameter of the pump housing in the zone of the suction wheel decreases in the direction of flow. The inside diameter of the pump housing at the entry to the sucti on wheel is cal-culated from the formula:
The suction wheel comprises a hub with helical vanes at-tached thereto. The vane setting angle increases in the direc-tion of flow and at the entry to the suction wheel it is:
where D0 = inside diameter of the pump housing at the entry to the suction wheel;
D1 = inside diameter of the pump housing at the entry to the impeller wheel;
D1 = dimensionless coefficient of 0.17 to 0.13;
Ck = predetermined cavitation critical speed coeffi-cient of 5.000 to 11.000;
.DELTA. = radial clearance.
Figure 1.
Description
lZ3~
~ E PUMP
The invention relates to pump engineering and has parti-cular refere~ce to vane pumps.
~ he invention can be uaed in chemical and petroleum pro~
cessing indu~tries, melioration and other applications.
~ he invention can be used with particular advantage in power engineerin~, shipbuildi~g, and aerospace technology, more specifically, in high-output pumps with low suction head or in high-speed pumps.
One of the most important characteristics of pumps is auction capacity expressed by the cavitation critical speed coefficient: `
Ck = 5.62 n~;~ (1) here n = drive shaft rotational speed, in revolutions per minute;
Q = volumetric rate of flow of the liquid being pumped, i.e. pump output, in cubic meters per second;
a h = net positive suction head, i~ meters.
~ he greater the coef~icient Ck, the greater the pump suction capacity.
~ he rotational speed of the pump drive ~haft determines the size and mass of the pump, whereas tha pump output and the suction capacity determine, respectively, the quantit~ of the pumps required for the given job a~d the capital outlay.
For example, doubling the pump suction capacity, with unchan-ged suction head, enables doubling the rotational apeed of the pump drive shaft, whereby the pump size and ma8s can be ., '''.~
decreased twice or thrice, making for substantial reduction of pump manufacturing costs. The current trend toward incre-ase in the capacity of single power units calls for further increase in pump output with consequent increase in suction head. In high output pumps, increasing suction head is restr-ained by cost considerations. On the other hand~ increasin~
pump suction capacity, for example, twice, makes it possible to uqe one large-output pump instead of four pumps with an equivalent total output and to decrease suction head outlay at least thrice.
Thus, there is a great nsed in pump engineerin~ for in-creasing pump suction capacity.
Insufficient suction capacity of a pump causes cavita-tion with resultant decrease in head and efficiency.
The specific point of the problem is that increase in the suction capacity of a pump is usually accompanied by de-crease in the pump efficiency 1, which causes substantial in-crease in power consumption. ~herefore, as a rule, pumps with high suction capacity have low efficiency, whereas pumps with high efficiency have low suction capacity.
Enown in the art ars pumps with high suction capacity (Ck~ 4,000).
Such a pump comprises an axial-flow impeller wheel which is mounted on a drive shaft and has a hub with helical v~nes attached thereto. ~he vanes are profiled alo~g the wheel ra-dius according to the expression r.tg~ = cost, where r is the current value of the radius of the axial-flow wheel and~ is the van6 setting angle between the plane normal to the pump drive shaft and the plane tangential to the vane,s.
11~ 32 The suction capacity of this pump is increased by vir-tue of increasing the cross-sectional area of the pump flow duct and decreasing ~he vane angle, which results in decrease of t~e wheel entry velocity ratio ~, i.e. the ratio of the axial velocity Cl of the liquid flow to the peripheral velo-city Ul of the pump wheel on the outside diameter thereof.
Increase in the cross-sectional area of the pump flow duct is achieved by increasing the outside diameter of the pumpe wheel and by decreasing the hub diameter as much as possible with respect to strength considerations. This solution ensures de-crease in the axial component of the liquid flow velocity and provides the minimum drop of static pressure in the liquid flow, whereby the suction capacity of the pump is increased.
However, this pump has a low efficiency ( ~ = 0.5) inas-much as the velocity ratio is low ( y< 0.1) due to increase in the cross-sectional area of the pump flow duct, decrease in the axial velocity Cl of the liquid flow, and breakaway nature of flow through the wheel.
Enown in the art are ~ane pumps with the efficiency ~ as high as 0.75 to 0.9.
Such a pump comprises a housing which accommodates an impeller wheel mounted on a drive shaft and having a hub with vanes attached thereto. ~he developments of the cylindrical sections of said vanes form a cascade of airfoils set at re-latively large angles between the airfoil chord and the cas-cade front, 3aid angles being suitable for an increased velo-city ratio ( ~ ~0.2).
-- 4 _ However? this pump has a substantially low suction capa city (Ck __l,OOO) in connection with relatively high axial velociti ~1 of the liquid flow due to decrease in the cross--sectional area of the impeller wheel flow duct.
Enown in the art are pumps with high suction capacity (Ck is as large as 4,200 to 5,200) and pumps with a relative suction velocity Ss of 40.000 to 60,000, where Ss = 9.19 Ck.
In these pumps, in order to provide high suction capacity, use is made ~f an axial-flow wheel mou~ted on a drive shaft together with an impeller wheel. ~he axial-flow wheel is high-ly immune to cavitation and develops a head sufficient to pro-vide for cavitation-free operation of the impeller whesl.
In the pumps of the prior art use is made of the follow-ing means in order to increase suction capacity:
- a worm with a le~gthwise variable pitch;
- a taper worm mounted in a confu~er;
- a worm with a taper hub, variable diameter and blade pitch, and an entry rake;
- an upstream, axial-flow, converging wheel with a taper shroud;
- a worm in the form of an axially movable helix;
- an upstream taper w~leel with a helical thread on the outer surface;
- an inlet device installed before a centrifugal wheel and comprising several rows of vanes gradually increasing in diameter;
- an upstream, axial-flow wheel the estimated rate of flow through which is three times greater than that throu~h a centrifugal wheel;
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- a conical hub changing into a radial-flow wheel, which hub mounts several circular rows of round-section pins inst-alled at a varying ~ngle to the axis of rotation;
- an upstream, single- or multiple-start worm or a coni-cal, ribbed head;
- an upstream, double-stage, axial-~low wheel wherein the vanes o~ each stage have different diameter and angle of pitch;
- an upstream, axial-flow wheel with a bypass device for recirculating fluid in the zone of a worm.
The pump constructions considered above do not provide for the maximum possible increase in action capacity. Further-more, while improving some parameters, for e~ample, ca~itati~n characteristics, they impair others, for example, pump effici-ency or stability.
~ nown in the art is a vane pump comprising a housing and two axial-flow wheels, viz. a suction wheel and an impeller wheel, which are mounted on a common drive shaft and in~talled with a radial clearance in the housing. The suction wheel has a hub with helical vanes attached thereto, the pitch of the vanes increasing in the direction of flow.
The vane pitch on the suction wheel i8 chosen so as to provide high suction capacity of the pump , whereas the vane pitch on the impeller wheel is chosen so as to provide the re-quired head and increase pump efficiency. The pump operates as follows:
The liquid first enters the axial-flow suction wheel.As the flow passes over the vanes, cavitation originates and de-velops. At the end of the suction wheel the cavitation ceases.
After the suction wheel the liquid, which has acquired some energy, enters the axial-flow impeller wheel which creates, in the main, the required head. ~he pump under consideration provides high suction capacity (Ck = ~.000) and an increased efficiency, but these parameters are not at a maximum ina~-much as the radial clearance of the axial-flow wheels and its relation to the wheel geometry ar~ not stipulated.
~ he technical solutions described above merel~ disclose the level achieved in the prior art in the endeavor to provi-de for a pump to have both high suction capacity and high ef-ficiency, which level is, of course, is not at the utmost.
It is an object of the present invention to device a pump with a specially profiled in- et duct wherein the inside diameter varies according to the geometrical dimensions of the axial-flow wheel involved, thereby increasing the suction capacity o* the pump and also improving the ener6y characte-ristics thereof.
~ he invention provides a vane pump comprising a housing wherein an a~i~l-flow suction wheel and an axial-flow impel-ler wheel are installed with a radial clearance, which suc-tion and impeller whèels are mounted on a common drive shaft one after the other in the direction of flow. The suction wheel comprises a hub with helical vanes of varying pitch at-tached thereto. The inside diameter of the pump housing in the zone of the suction wheel decreases in the direction of flow. ~he vanes of the suction wheel have a varying a~gle of setting at the tip, said angle increasing in the direction of flow. The inside diameter of the pump housing at the entry to the suction wheel is chose~ according to:
Do = Dl El(Ck' 10 + 2~1)2 (2) here Do = inside diameter of the pump housing at the entry to the suction wheel;
Dl = inside diameter of the pump housing at the entry to the impeller wheel;
Kl = dimensionless coefficient of o.~7 to 0.13;
Ck = predetermined cavitation critical speed coeffici-ent of 5.000 to 11.000.
The vane tip setting angle of the suction wheel at the entry thereto is determined by:
~0 = (10 to 33 a /Dl) ~ 1.5~
whe~e ~ O = vane tip setting angle o~ the suction wheel at the entry thereto;
a = radial clearance at the entry to the suction wheel;
Dl = inside diameter of the pump housing at the entry to the impeller wheel.
~ he constructional solution described above provides ~or substantial increase of pump suction capacity. This is attri-buted to the provision of an increased radial clearance bet-ween the outside diameter of the axial-flow suction wheel and the inside diameter of the pump housing, due to which the li-guid flow at the entry to the suction wheel is divided into two flows one ~f which passes through the clearance and the other through said wheel.
By reference to the relation (1) we find that at a given pump drive shaft speed and a given cavitatio~ critical speed coefficient a lesser net positive suction head is required to provide for cavitation-free operation of the axial-flow suc-tion wheel at a decreased volumetric rate of flow t~rough the pump. With respect to the pump as a whole, decrease in the required net positive suction head at a given volumetric rate of flow and a givsn pump drive 6haft speed results in substan-tial increase of pump suction capacity.
The invention will now be more particularly described by wa~ of example with reference to the accompanying drawings, wherein: -Figure 1 i~ a longitudinal sectional view of an embodi-ment of the axial-radial flow vane pump according to the in-vention.
Figure 2 i8 a longitudinal sectional view of an embodi-ment of the axial-diagonal flow vane pump according to the invention.
Figure 3 is a development of the cylindrical section of the axial-flow suction wheel according to the invention.
~ igure 4 is a graph showing the coefficient of the pump housing diameter at the entry to the axial-flow suction wheel versus the cavitation critical speed coefficient according to the invention.
Figure 5 is a graph showing the cavitation critical spe-ed coefficient versus specific rate of flow at two vane sett-ing angles on the suction wheel as obtained in testing the pump embodiment of Figure 1.
_ g _ ~ he vane pump comprises a housing 1 (Figure 1) which, in this em~odiment, is made in two pieces, viz. an inlet 2 and a scroll outlet 3. The housing 1 accommodat~s a drive shaft 4 mounted on which one after the other in the direction of flow are an axial-flow suction wheel 5, an axial-flow im-peller wheel 6 and a radial-flow wheel 7. ~he impeller wheel 6 has a hub 8 to which are attached helical vanes 9 forming intervane passages 10 for the ~low of liquid. ~he suction wheel 5 comprises a hub 11 to which are attached helical vanes 12 forming intervane passages 1~. The vanes 12 on the suction wheel 5 have varying pitch which increases in the direction of flow.
The symbols used in Figures 1 and 2 have the following significance:
S = pitch of helix of the vane 12 on the suction wheel 5;
Dl = inside diameter of the housing 1 (in this case of the inlet 2) at the entry to the impeller wheel 6;
Do = inside diameter of the housing 1 at the entry to the suction wheel 5;
= radial clearance at the entry to the suction wheel 5.
In another embodiment of the pump the axial-flow suction wheel 5 is used in conjunction with an axial-diagonal flow wheel 14 (Figure 2) and the drive sha~t 4 is mounted in bear-ings 15 installed in the housing l.The axial-diagonal flow wheel 14 consists of three portions, viz. an inlet axial-flow portion 16 which forms a cavitation part, a diagonal portion 17 which forms a delivery part, and an outlet axial-flow por-tion 18 which forms a straightening part. The inlet axial-flow portion 16 of the axial-diagoral flow wheel 14 fulfils the " 11~1~32 same function as the axial-flow impeller wheel 6 (Figure 1) and the entry to the wheel 14 is considsred the same as the entry to the wheel 6. ~he inside diameter of the pump housing 1 varies in the direction of flow from the maxLmum diameter Do at the entry to the suction wheel 5 to the minimum diamet-er Dl at the entry to the axial-diagonal flow wheel 14 and then to the diameter D2 at the entry to the outlet 3.
The settin$ angle ~ (Figure 3) of the vanes 12 (Figur~I) on the suction wheel 5 is formed between the plane normal to the axis of rotation of the wheel 5 and the plane tangential to the vane 12 of the wheel 5. The direction of the axial ve-locity of the flow is indicated by the arrow Cl (Figure ~).
The direction of the peripheral velocity of the wheel 5 (~ig-ura 1) is indicated by the arrow ~1 (Figure 3).
The graph in Figure 4 shows experimental curves for the coefficient ~D of the pwmp housing diameter at the entry to the axial-flow suction wheel 5 (Figure 1) versus th~ cavita-tion critical speed coefficient Ck for four pumps.
Do 3 , ~ (4) ,~
where Do = inside diameter of the pump housing at the entry to the suction wheel 5;
Q _ rate of flow through the pump;
n - rotational speed of the drive shaft 4.
The graph in Figure 5 shows the cavitation critical speed coefficient Ck versus the specific rate of flow Q .
1:~41Z3Z
The curve 20 is obtained in testing the pump embodiment depicted in Figure 1 and relates to the angle ~0 (Figur~ 3) of the vanes 12 (Figure 1) equal to 5. ~he cur~e 21 is for the angle ~ O (Figure 3) equal to 10.
~ he vane pump operates as follows;`
When the drive shaft 4 (Figure 1) rotates, the liquid being pwnped passes through the inlet 2 of the pump housing 1 into the rotating suction wheel 5. ~ part of the liquid goes through the intervane passa6e~ 13, whilst the other part en-ters the rotating impeller wheel 6 through the clearance~ bet-ween the pump housing 1 and the vanes 12 of the suction wheel 5. The power interaction between the vanes 12 and the liquid causes rise in the liquid pressure. ~he liquid proceeds into the impeller wheel 6 wherein it goes through the intervane passages 10. The power interaction between the vanes 9 and the liqu~d causes further pressure rise and the liquid thereafter enters ths radial-~low wheel 7. ~rom the intervane passages 10 of the impeller wheel 6 the liquid goes into the radial-~low wheel 7 wherein the liquid pressure is raised to the required value. ~he successive build-up of the li~uid pressure ensures cavitation-free operation of each Oe the pump wheels 5, 6 and 7.
From the radial-flow wheel 7 the liquid proceeds into the out-let 3 and thence into a delivery line (not shown).
~ he pump depicted in Figure 2 operates substantially in the same manner as that in Figure 1.
On the basis of theoretical and experimental data obtain-ed for several pwmps constructed according to the embodiments shown in Figures 1 and 2 relationship is found between the geometrical dimensions of the constructional elements determin-lZ3Z
ing the coefficient ~ (Figure 4) of the diameter of thepump housing 1 ~Figure 1) and the cavitation critical spead coefficient Ck (Figure 4) determining the required suction capacit~ of the.pump.
~ or pumps with a superhigh suction capacity, the coef~i-cient KD of the diameter of the housing 1 (Figure 1) should be cho~en according to the experimental curves depicted in Figure 4. ~his graphical relationship can be appro~imately represented in an analytical form:
KD0 - a (Ck ~ 10~ ~ 2.1) (5) where a _ 0.85 to 1.15 according to variation in the curves of Figure 4.
~ o provide for a superhigh suction capacity at the entry to the suction wheel 5 (Figure 1), the diameter Do of the housing 1 at this point should be found from:
- Do = a (Ck 10 4 + 2.1)2 . ~ ~ (6) where a = 0.85 to 1.15.
It is known that the pumps having a high suction capacity have a relatively low efficiency ( ~ = 0.5 to 0.65) because of low velocity ratio ( ~< 0.1) due to increase in the cross-sec-tional area of the flow duct, decrease in the ~low axial velo-city, and a breakaway nature of the ~low through the pump wheel.
According to the invention, the e~ficiency of a pump with a high suction capacity is to be increased by decreasing the inside diameter o~ the pump housing 1 in the direction o~ flow from the value Do calculatsd by the formula (6) to the value Dl found from:
Dl = EDl ~ 7 (7) here KD = 6 to 7 = coefficient of the diameter of the hous-ing 1 at the entry to the axial-~lov~ wheel 7 which provides for increase in efficiency.
~ rom the formulas (6) and (7) we find the relation betwe-en the inside diameters of the pump housing 1 which provides for the maximum suc~ion capacity of the suction wheel 5 and the maximum efficiency of the impeller wheel 6:
= Kl(Ck 10 4 + 2.1)2~ (8) where El = 0.17 to 0.13.
Tn this case increase in the pump suctio~ capacity is attributed firstly to increase in the cross-sectional area of the flow duct and, consequently, decrease in the velocity ra-tio ~ at the entry to the suction wheel 5 ( ~ is the ratio o~
the axial ~low velocity Cl to the peripheral velocity Ul of the wheel 5). This provides for decrease of the axial compo-nent o~ the flow velocity and for the minimum drop of static pressure in the flow, thereby bringir~ about increase in the pump suction capacity.
Secondly, increase in the pump suction capacit~ is attri-buted to Lncrease o~ the radial clearance ~ between the out-side diameter of the axial-flow suction wheel 5 and the inside diameter of the housing 1, due to which the flow at the entry 11~11i~3Z
to the axial-flow suction wheel 5 i~ divided into two ~lows, one of which passes through said clearance ~ and the ot~er through said wheel 5.
It follows from the formula (l) that, with a given rota-tional speed of the pump drive shaft 4 and a given cavitation critical speed coefficient Ck, the net positive suction head ~hk is to be decreased in order to provide cavitation-~ree operation of the axial-flow suction wheel 5 at a decreased volumetric rate of flow. ~s to the pump as a whole, decrease in the required net suction head at a given volumetric rate of flow and a given rotational speed o~ the pump drive shaft 4 brir~s about increase in the pump suction capacity.
It follows from the theory of perfect fluid flow about a cascade of i~finitely thin plates that the smaller the setting angle ~ 0 (~igure 3) of the vane 12, the better the anticavi-tation properties of the suction wheel 5 (Figure 1):
~ hk = 1 Ul tg~o where Cl = axial flow velocity at the entry to the wheel;
~l = wheel peripheral speed;
= vane setting a~gle at the entry;
hk = net positive ~uction head.
~ he experiments of the prior art have shown that in the case of pumps with high anticavitation properties, in which upstream worms have fimall vane setting angles ( ~ 0 ~ 20), the parameter ~ has practically no effect on pump cavitation characteristics if the diameters of the worm ~n~ housing are constant (refer, ~or example, to "Cavitation Characteristics A
of ~igh-speed Worm and Radial ~heel Pwmps" by V.F.Chebotaryov and V.I. Petrov, published in 1973 by "~ashinostroyeniye"
Publishers, ~o~cow, pages 117-118). In these experiments the values of the clearance ~ (Figure 1) between the worm and the housing were small.
Inasmuch as in the pump of the present invention the suction wheel 5 has a constant diameter, i~crease in the dia-meter of the housing 1 results in formation of a relatively large clearance ~ between the suction wheel 5 and the housing 1. In this case, according to the expe~imental data presented in Figure 5, decrease of ~ 0 (Figure 3) makes it possible to substantially increase the cavitation critical speed coeffici-ent Ck. The coefficient Ck thus obtained is 8.000 as compared -with the initial figures of 4.000 to 5.000.
The experiments with various values of the va~e angle ~ o (Figure ~) and the clearance ~ (Figure 1) have provided for finding the optimum relation therebetween with the view of bringing pump suction capacity to a maximum:
~ = (10 to 33 a /Dl) + 1.5 (10) where ~ = vane s~tting angle on the suction wheel 5;
= radial clearance between the outside diameter of the suction wheel 5 and the inside diameter of the pump housing 1 Dl = inside diameter of the pump housing 1 at the entry to the impeller wheel 6.
The experimental curves in Figure 4 are obtained with pump8, wherein, according to the formula (10), the radial clearance ~ (Figure 1) at the entry to the suction wheel 5 is "
1~41;Z3Z
large, whersas the angle ~ (Figure 3) of the vanes 12 (Fig-ure 1) is small.
Dimensioning the housing 1 and the axial-flow suction wheel 5 according to the formulas (8) and (10) provides for obtaining a cavitation coefficient Ck of 6.000 to 10.000 at an efficiency ~ as high as 0~6 to 0.8.
~ E PUMP
The invention relates to pump engineering and has parti-cular refere~ce to vane pumps.
~ he invention can be uaed in chemical and petroleum pro~
cessing indu~tries, melioration and other applications.
~ he invention can be used with particular advantage in power engineerin~, shipbuildi~g, and aerospace technology, more specifically, in high-output pumps with low suction head or in high-speed pumps.
One of the most important characteristics of pumps is auction capacity expressed by the cavitation critical speed coefficient: `
Ck = 5.62 n~;~ (1) here n = drive shaft rotational speed, in revolutions per minute;
Q = volumetric rate of flow of the liquid being pumped, i.e. pump output, in cubic meters per second;
a h = net positive suction head, i~ meters.
~ he greater the coef~icient Ck, the greater the pump suction capacity.
~ he rotational speed of the pump drive ~haft determines the size and mass of the pump, whereas tha pump output and the suction capacity determine, respectively, the quantit~ of the pumps required for the given job a~d the capital outlay.
For example, doubling the pump suction capacity, with unchan-ged suction head, enables doubling the rotational apeed of the pump drive shaft, whereby the pump size and ma8s can be ., '''.~
decreased twice or thrice, making for substantial reduction of pump manufacturing costs. The current trend toward incre-ase in the capacity of single power units calls for further increase in pump output with consequent increase in suction head. In high output pumps, increasing suction head is restr-ained by cost considerations. On the other hand~ increasin~
pump suction capacity, for example, twice, makes it possible to uqe one large-output pump instead of four pumps with an equivalent total output and to decrease suction head outlay at least thrice.
Thus, there is a great nsed in pump engineerin~ for in-creasing pump suction capacity.
Insufficient suction capacity of a pump causes cavita-tion with resultant decrease in head and efficiency.
The specific point of the problem is that increase in the suction capacity of a pump is usually accompanied by de-crease in the pump efficiency 1, which causes substantial in-crease in power consumption. ~herefore, as a rule, pumps with high suction capacity have low efficiency, whereas pumps with high efficiency have low suction capacity.
Enown in the art ars pumps with high suction capacity (Ck~ 4,000).
Such a pump comprises an axial-flow impeller wheel which is mounted on a drive shaft and has a hub with helical v~nes attached thereto. ~he vanes are profiled alo~g the wheel ra-dius according to the expression r.tg~ = cost, where r is the current value of the radius of the axial-flow wheel and~ is the van6 setting angle between the plane normal to the pump drive shaft and the plane tangential to the vane,s.
11~ 32 The suction capacity of this pump is increased by vir-tue of increasing the cross-sectional area of the pump flow duct and decreasing ~he vane angle, which results in decrease of t~e wheel entry velocity ratio ~, i.e. the ratio of the axial velocity Cl of the liquid flow to the peripheral velo-city Ul of the pump wheel on the outside diameter thereof.
Increase in the cross-sectional area of the pump flow duct is achieved by increasing the outside diameter of the pumpe wheel and by decreasing the hub diameter as much as possible with respect to strength considerations. This solution ensures de-crease in the axial component of the liquid flow velocity and provides the minimum drop of static pressure in the liquid flow, whereby the suction capacity of the pump is increased.
However, this pump has a low efficiency ( ~ = 0.5) inas-much as the velocity ratio is low ( y< 0.1) due to increase in the cross-sectional area of the pump flow duct, decrease in the axial velocity Cl of the liquid flow, and breakaway nature of flow through the wheel.
Enown in the art are ~ane pumps with the efficiency ~ as high as 0.75 to 0.9.
Such a pump comprises a housing which accommodates an impeller wheel mounted on a drive shaft and having a hub with vanes attached thereto. ~he developments of the cylindrical sections of said vanes form a cascade of airfoils set at re-latively large angles between the airfoil chord and the cas-cade front, 3aid angles being suitable for an increased velo-city ratio ( ~ ~0.2).
-- 4 _ However? this pump has a substantially low suction capa city (Ck __l,OOO) in connection with relatively high axial velociti ~1 of the liquid flow due to decrease in the cross--sectional area of the impeller wheel flow duct.
Enown in the art are pumps with high suction capacity (Ck is as large as 4,200 to 5,200) and pumps with a relative suction velocity Ss of 40.000 to 60,000, where Ss = 9.19 Ck.
In these pumps, in order to provide high suction capacity, use is made ~f an axial-flow wheel mou~ted on a drive shaft together with an impeller wheel. ~he axial-flow wheel is high-ly immune to cavitation and develops a head sufficient to pro-vide for cavitation-free operation of the impeller whesl.
In the pumps of the prior art use is made of the follow-ing means in order to increase suction capacity:
- a worm with a le~gthwise variable pitch;
- a taper worm mounted in a confu~er;
- a worm with a taper hub, variable diameter and blade pitch, and an entry rake;
- an upstream, axial-flow, converging wheel with a taper shroud;
- a worm in the form of an axially movable helix;
- an upstream taper w~leel with a helical thread on the outer surface;
- an inlet device installed before a centrifugal wheel and comprising several rows of vanes gradually increasing in diameter;
- an upstream, axial-flow wheel the estimated rate of flow through which is three times greater than that throu~h a centrifugal wheel;
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- a conical hub changing into a radial-flow wheel, which hub mounts several circular rows of round-section pins inst-alled at a varying ~ngle to the axis of rotation;
- an upstream, single- or multiple-start worm or a coni-cal, ribbed head;
- an upstream, double-stage, axial-~low wheel wherein the vanes o~ each stage have different diameter and angle of pitch;
- an upstream, axial-flow wheel with a bypass device for recirculating fluid in the zone of a worm.
The pump constructions considered above do not provide for the maximum possible increase in action capacity. Further-more, while improving some parameters, for e~ample, ca~itati~n characteristics, they impair others, for example, pump effici-ency or stability.
~ nown in the art is a vane pump comprising a housing and two axial-flow wheels, viz. a suction wheel and an impeller wheel, which are mounted on a common drive shaft and in~talled with a radial clearance in the housing. The suction wheel has a hub with helical vanes attached thereto, the pitch of the vanes increasing in the direction of flow.
The vane pitch on the suction wheel i8 chosen so as to provide high suction capacity of the pump , whereas the vane pitch on the impeller wheel is chosen so as to provide the re-quired head and increase pump efficiency. The pump operates as follows:
The liquid first enters the axial-flow suction wheel.As the flow passes over the vanes, cavitation originates and de-velops. At the end of the suction wheel the cavitation ceases.
After the suction wheel the liquid, which has acquired some energy, enters the axial-flow impeller wheel which creates, in the main, the required head. ~he pump under consideration provides high suction capacity (Ck = ~.000) and an increased efficiency, but these parameters are not at a maximum ina~-much as the radial clearance of the axial-flow wheels and its relation to the wheel geometry ar~ not stipulated.
~ he technical solutions described above merel~ disclose the level achieved in the prior art in the endeavor to provi-de for a pump to have both high suction capacity and high ef-ficiency, which level is, of course, is not at the utmost.
It is an object of the present invention to device a pump with a specially profiled in- et duct wherein the inside diameter varies according to the geometrical dimensions of the axial-flow wheel involved, thereby increasing the suction capacity o* the pump and also improving the ener6y characte-ristics thereof.
~ he invention provides a vane pump comprising a housing wherein an a~i~l-flow suction wheel and an axial-flow impel-ler wheel are installed with a radial clearance, which suc-tion and impeller whèels are mounted on a common drive shaft one after the other in the direction of flow. The suction wheel comprises a hub with helical vanes of varying pitch at-tached thereto. The inside diameter of the pump housing in the zone of the suction wheel decreases in the direction of flow. ~he vanes of the suction wheel have a varying a~gle of setting at the tip, said angle increasing in the direction of flow. The inside diameter of the pump housing at the entry to the suction wheel is chose~ according to:
Do = Dl El(Ck' 10 + 2~1)2 (2) here Do = inside diameter of the pump housing at the entry to the suction wheel;
Dl = inside diameter of the pump housing at the entry to the impeller wheel;
Kl = dimensionless coefficient of o.~7 to 0.13;
Ck = predetermined cavitation critical speed coeffici-ent of 5.000 to 11.000.
The vane tip setting angle of the suction wheel at the entry thereto is determined by:
~0 = (10 to 33 a /Dl) ~ 1.5~
whe~e ~ O = vane tip setting angle o~ the suction wheel at the entry thereto;
a = radial clearance at the entry to the suction wheel;
Dl = inside diameter of the pump housing at the entry to the impeller wheel.
~ he constructional solution described above provides ~or substantial increase of pump suction capacity. This is attri-buted to the provision of an increased radial clearance bet-ween the outside diameter of the axial-flow suction wheel and the inside diameter of the pump housing, due to which the li-guid flow at the entry to the suction wheel is divided into two flows one ~f which passes through the clearance and the other through said wheel.
By reference to the relation (1) we find that at a given pump drive shaft speed and a given cavitatio~ critical speed coefficient a lesser net positive suction head is required to provide for cavitation-free operation of the axial-flow suc-tion wheel at a decreased volumetric rate of flow t~rough the pump. With respect to the pump as a whole, decrease in the required net positive suction head at a given volumetric rate of flow and a givsn pump drive 6haft speed results in substan-tial increase of pump suction capacity.
The invention will now be more particularly described by wa~ of example with reference to the accompanying drawings, wherein: -Figure 1 i~ a longitudinal sectional view of an embodi-ment of the axial-radial flow vane pump according to the in-vention.
Figure 2 i8 a longitudinal sectional view of an embodi-ment of the axial-diagonal flow vane pump according to the invention.
Figure 3 is a development of the cylindrical section of the axial-flow suction wheel according to the invention.
~ igure 4 is a graph showing the coefficient of the pump housing diameter at the entry to the axial-flow suction wheel versus the cavitation critical speed coefficient according to the invention.
Figure 5 is a graph showing the cavitation critical spe-ed coefficient versus specific rate of flow at two vane sett-ing angles on the suction wheel as obtained in testing the pump embodiment of Figure 1.
_ g _ ~ he vane pump comprises a housing 1 (Figure 1) which, in this em~odiment, is made in two pieces, viz. an inlet 2 and a scroll outlet 3. The housing 1 accommodat~s a drive shaft 4 mounted on which one after the other in the direction of flow are an axial-flow suction wheel 5, an axial-flow im-peller wheel 6 and a radial-flow wheel 7. ~he impeller wheel 6 has a hub 8 to which are attached helical vanes 9 forming intervane passages 10 for the ~low of liquid. ~he suction wheel 5 comprises a hub 11 to which are attached helical vanes 12 forming intervane passages 1~. The vanes 12 on the suction wheel 5 have varying pitch which increases in the direction of flow.
The symbols used in Figures 1 and 2 have the following significance:
S = pitch of helix of the vane 12 on the suction wheel 5;
Dl = inside diameter of the housing 1 (in this case of the inlet 2) at the entry to the impeller wheel 6;
Do = inside diameter of the housing 1 at the entry to the suction wheel 5;
= radial clearance at the entry to the suction wheel 5.
In another embodiment of the pump the axial-flow suction wheel 5 is used in conjunction with an axial-diagonal flow wheel 14 (Figure 2) and the drive sha~t 4 is mounted in bear-ings 15 installed in the housing l.The axial-diagonal flow wheel 14 consists of three portions, viz. an inlet axial-flow portion 16 which forms a cavitation part, a diagonal portion 17 which forms a delivery part, and an outlet axial-flow por-tion 18 which forms a straightening part. The inlet axial-flow portion 16 of the axial-diagoral flow wheel 14 fulfils the " 11~1~32 same function as the axial-flow impeller wheel 6 (Figure 1) and the entry to the wheel 14 is considsred the same as the entry to the wheel 6. ~he inside diameter of the pump housing 1 varies in the direction of flow from the maxLmum diameter Do at the entry to the suction wheel 5 to the minimum diamet-er Dl at the entry to the axial-diagonal flow wheel 14 and then to the diameter D2 at the entry to the outlet 3.
The settin$ angle ~ (Figure 3) of the vanes 12 (Figur~I) on the suction wheel 5 is formed between the plane normal to the axis of rotation of the wheel 5 and the plane tangential to the vane 12 of the wheel 5. The direction of the axial ve-locity of the flow is indicated by the arrow Cl (Figure ~).
The direction of the peripheral velocity of the wheel 5 (~ig-ura 1) is indicated by the arrow ~1 (Figure 3).
The graph in Figure 4 shows experimental curves for the coefficient ~D of the pwmp housing diameter at the entry to the axial-flow suction wheel 5 (Figure 1) versus th~ cavita-tion critical speed coefficient Ck for four pumps.
Do 3 , ~ (4) ,~
where Do = inside diameter of the pump housing at the entry to the suction wheel 5;
Q _ rate of flow through the pump;
n - rotational speed of the drive shaft 4.
The graph in Figure 5 shows the cavitation critical speed coefficient Ck versus the specific rate of flow Q .
1:~41Z3Z
The curve 20 is obtained in testing the pump embodiment depicted in Figure 1 and relates to the angle ~0 (Figur~ 3) of the vanes 12 (Figure 1) equal to 5. ~he cur~e 21 is for the angle ~ O (Figure 3) equal to 10.
~ he vane pump operates as follows;`
When the drive shaft 4 (Figure 1) rotates, the liquid being pwnped passes through the inlet 2 of the pump housing 1 into the rotating suction wheel 5. ~ part of the liquid goes through the intervane passa6e~ 13, whilst the other part en-ters the rotating impeller wheel 6 through the clearance~ bet-ween the pump housing 1 and the vanes 12 of the suction wheel 5. The power interaction between the vanes 12 and the liquid causes rise in the liquid pressure. ~he liquid proceeds into the impeller wheel 6 wherein it goes through the intervane passages 10. The power interaction between the vanes 9 and the liqu~d causes further pressure rise and the liquid thereafter enters ths radial-~low wheel 7. ~rom the intervane passages 10 of the impeller wheel 6 the liquid goes into the radial-~low wheel 7 wherein the liquid pressure is raised to the required value. ~he successive build-up of the li~uid pressure ensures cavitation-free operation of each Oe the pump wheels 5, 6 and 7.
From the radial-flow wheel 7 the liquid proceeds into the out-let 3 and thence into a delivery line (not shown).
~ he pump depicted in Figure 2 operates substantially in the same manner as that in Figure 1.
On the basis of theoretical and experimental data obtain-ed for several pwmps constructed according to the embodiments shown in Figures 1 and 2 relationship is found between the geometrical dimensions of the constructional elements determin-lZ3Z
ing the coefficient ~ (Figure 4) of the diameter of thepump housing 1 ~Figure 1) and the cavitation critical spead coefficient Ck (Figure 4) determining the required suction capacit~ of the.pump.
~ or pumps with a superhigh suction capacity, the coef~i-cient KD of the diameter of the housing 1 (Figure 1) should be cho~en according to the experimental curves depicted in Figure 4. ~his graphical relationship can be appro~imately represented in an analytical form:
KD0 - a (Ck ~ 10~ ~ 2.1) (5) where a _ 0.85 to 1.15 according to variation in the curves of Figure 4.
~ o provide for a superhigh suction capacity at the entry to the suction wheel 5 (Figure 1), the diameter Do of the housing 1 at this point should be found from:
- Do = a (Ck 10 4 + 2.1)2 . ~ ~ (6) where a = 0.85 to 1.15.
It is known that the pumps having a high suction capacity have a relatively low efficiency ( ~ = 0.5 to 0.65) because of low velocity ratio ( ~< 0.1) due to increase in the cross-sec-tional area of the flow duct, decrease in the ~low axial velo-city, and a breakaway nature of the ~low through the pump wheel.
According to the invention, the e~ficiency of a pump with a high suction capacity is to be increased by decreasing the inside diameter o~ the pump housing 1 in the direction o~ flow from the value Do calculatsd by the formula (6) to the value Dl found from:
Dl = EDl ~ 7 (7) here KD = 6 to 7 = coefficient of the diameter of the hous-ing 1 at the entry to the axial-~lov~ wheel 7 which provides for increase in efficiency.
~ rom the formulas (6) and (7) we find the relation betwe-en the inside diameters of the pump housing 1 which provides for the maximum suc~ion capacity of the suction wheel 5 and the maximum efficiency of the impeller wheel 6:
= Kl(Ck 10 4 + 2.1)2~ (8) where El = 0.17 to 0.13.
Tn this case increase in the pump suctio~ capacity is attributed firstly to increase in the cross-sectional area of the flow duct and, consequently, decrease in the velocity ra-tio ~ at the entry to the suction wheel 5 ( ~ is the ratio o~
the axial ~low velocity Cl to the peripheral velocity Ul of the wheel 5). This provides for decrease of the axial compo-nent o~ the flow velocity and for the minimum drop of static pressure in the flow, thereby bringir~ about increase in the pump suction capacity.
Secondly, increase in the pump suction capacit~ is attri-buted to Lncrease o~ the radial clearance ~ between the out-side diameter of the axial-flow suction wheel 5 and the inside diameter of the housing 1, due to which the flow at the entry 11~11i~3Z
to the axial-flow suction wheel 5 i~ divided into two ~lows, one of which passes through said clearance ~ and the ot~er through said wheel 5.
It follows from the formula (l) that, with a given rota-tional speed of the pump drive shaft 4 and a given cavitation critical speed coefficient Ck, the net positive suction head ~hk is to be decreased in order to provide cavitation-~ree operation of the axial-flow suction wheel 5 at a decreased volumetric rate of flow. ~s to the pump as a whole, decrease in the required net suction head at a given volumetric rate of flow and a given rotational speed o~ the pump drive shaft 4 brir~s about increase in the pump suction capacity.
It follows from the theory of perfect fluid flow about a cascade of i~finitely thin plates that the smaller the setting angle ~ 0 (~igure 3) of the vane 12, the better the anticavi-tation properties of the suction wheel 5 (Figure 1):
~ hk = 1 Ul tg~o where Cl = axial flow velocity at the entry to the wheel;
~l = wheel peripheral speed;
= vane setting a~gle at the entry;
hk = net positive ~uction head.
~ he experiments of the prior art have shown that in the case of pumps with high anticavitation properties, in which upstream worms have fimall vane setting angles ( ~ 0 ~ 20), the parameter ~ has practically no effect on pump cavitation characteristics if the diameters of the worm ~n~ housing are constant (refer, ~or example, to "Cavitation Characteristics A
of ~igh-speed Worm and Radial ~heel Pwmps" by V.F.Chebotaryov and V.I. Petrov, published in 1973 by "~ashinostroyeniye"
Publishers, ~o~cow, pages 117-118). In these experiments the values of the clearance ~ (Figure 1) between the worm and the housing were small.
Inasmuch as in the pump of the present invention the suction wheel 5 has a constant diameter, i~crease in the dia-meter of the housing 1 results in formation of a relatively large clearance ~ between the suction wheel 5 and the housing 1. In this case, according to the expe~imental data presented in Figure 5, decrease of ~ 0 (Figure 3) makes it possible to substantially increase the cavitation critical speed coeffici-ent Ck. The coefficient Ck thus obtained is 8.000 as compared -with the initial figures of 4.000 to 5.000.
The experiments with various values of the va~e angle ~ o (Figure ~) and the clearance ~ (Figure 1) have provided for finding the optimum relation therebetween with the view of bringing pump suction capacity to a maximum:
~ = (10 to 33 a /Dl) + 1.5 (10) where ~ = vane s~tting angle on the suction wheel 5;
= radial clearance between the outside diameter of the suction wheel 5 and the inside diameter of the pump housing 1 Dl = inside diameter of the pump housing 1 at the entry to the impeller wheel 6.
The experimental curves in Figure 4 are obtained with pump8, wherein, according to the formula (10), the radial clearance ~ (Figure 1) at the entry to the suction wheel 5 is "
1~41;Z3Z
large, whersas the angle ~ (Figure 3) of the vanes 12 (Fig-ure 1) is small.
Dimensioning the housing 1 and the axial-flow suction wheel 5 according to the formulas (8) and (10) provides for obtaining a cavitation coefficient Ck of 6.000 to 10.000 at an efficiency ~ as high as 0~6 to 0.8.
Claims
1. A vane pump comprising:
- a housing with a varying inside diameter;
- a drive shaft;
- an axial-flow impeller wheel with an entry thereto;
- an axial-flow suction wheel with an entry thereto, said axial-flow suction and impeller wheels setting up the direction of flow, which wheels are installed in the pump housing with a radial clearance and are mounted on said drive shaft one after the other in the direction of flow;
- a hub of said suction wheel;
- helical vanes of said suction wheel, which vanes are attached to said hub and have varying pitch, the vane tip setting angle increasing in the direction of flow, said sett-ing angle at the entry to said suction wheel being calculated from the formula:
where ?0 = vane tip setting angle on said suction wheel at the entry thereto;
.DELTA. = radial clearance at the entry to said suction wheel;
D1 = inside diameter of said pump housing at the entry to said impeller wheel;
- said housing accommodating said drive shaft with said axial-flow suction and impeller wheels, which housing has in-side diameter at the entry to said suction wheel being chosen according to:
, where D0 = inside diameter of said pump housing at the entry to said suction wheel;
D1 = inside diameter of said pump housing at the entry to said impeller wheel;
K1 = dimensionless coefficient of 0.17 to 0.13;
Ck = predetermined cavitation critical speed coeffici-ent of 5.000 to 11.000.
- a housing with a varying inside diameter;
- a drive shaft;
- an axial-flow impeller wheel with an entry thereto;
- an axial-flow suction wheel with an entry thereto, said axial-flow suction and impeller wheels setting up the direction of flow, which wheels are installed in the pump housing with a radial clearance and are mounted on said drive shaft one after the other in the direction of flow;
- a hub of said suction wheel;
- helical vanes of said suction wheel, which vanes are attached to said hub and have varying pitch, the vane tip setting angle increasing in the direction of flow, said sett-ing angle at the entry to said suction wheel being calculated from the formula:
where ?0 = vane tip setting angle on said suction wheel at the entry thereto;
.DELTA. = radial clearance at the entry to said suction wheel;
D1 = inside diameter of said pump housing at the entry to said impeller wheel;
- said housing accommodating said drive shaft with said axial-flow suction and impeller wheels, which housing has in-side diameter at the entry to said suction wheel being chosen according to:
, where D0 = inside diameter of said pump housing at the entry to said suction wheel;
D1 = inside diameter of said pump housing at the entry to said impeller wheel;
K1 = dimensionless coefficient of 0.17 to 0.13;
Ck = predetermined cavitation critical speed coeffici-ent of 5.000 to 11.000.
Applications Claiming Priority (2)
| Application Number | Priority Date | Filing Date | Title |
|---|---|---|---|
| SU792739411A SU1023138A1 (en) | 1979-03-19 | 1979-03-19 | Impeller pump |
| SU2739411 | 1979-03-19 |
Publications (1)
| Publication Number | Publication Date |
|---|---|
| CA1141232A true CA1141232A (en) | 1983-02-15 |
Family
ID=20816403
Family Applications (1)
| Application Number | Title | Priority Date | Filing Date |
|---|---|---|---|
| CA000347916A Expired CA1141232A (en) | 1979-03-19 | 1980-03-18 | Vane pump |
Country Status (10)
| Country | Link |
|---|---|
| JP (1) | JPS6046276B2 (en) |
| CA (1) | CA1141232A (en) |
| DE (1) | DE3038735C2 (en) |
| DK (1) | DK479080A (en) |
| FI (1) | FI67435C (en) |
| FR (1) | FR2452016A1 (en) |
| GB (1) | GB2058923B (en) |
| SE (1) | SE444351B (en) |
| SU (1) | SU1023138A1 (en) |
| WO (1) | WO1980002058A1 (en) |
Families Citing this family (3)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US4884943A (en) * | 1987-06-25 | 1989-12-05 | A. Ahlstrom Corporation | Method and apparatus for pumping high-consistency fiber suspension |
| DE4314478A1 (en) * | 1993-05-03 | 1994-11-10 | Klein Schanzlin & Becker Ag | Inlet housing for centrifugal pumps |
| DE19517900A1 (en) * | 1995-05-16 | 1996-11-21 | Itt Flygt Pumpen Gmbh | Submersible motor-driven pump for viscous media, e.g. liquid manure |
Family Cites Families (8)
| Publication number | Priority date | Publication date | Assignee | Title |
|---|---|---|---|---|
| US3299821A (en) * | 1964-08-21 | 1967-01-24 | Sundstrand Corp | Pump inducer |
| GB1153993A (en) * | 1965-06-16 | 1969-06-04 | Rolls Royce | Rotary Impeller Pumps |
| DE1931527U (en) * | 1965-11-26 | 1966-01-27 | Steinebrunner & Co Maschinenfa | DEVICE FOR TRIMMING BRUSHES WITH A CENTRAL FIELD DETACHED FROM THE OUTSIDE FIELD. |
| JPS4331495Y1 (en) * | 1966-01-24 | 1968-12-20 | ||
| US3522997A (en) * | 1968-07-01 | 1970-08-04 | Rylewski Eugeniusz | Inducer |
| AT313066B (en) * | 1969-07-31 | 1974-01-25 | Worthington Corp | centrifugal pump |
| NL7111948A (en) * | 1971-08-31 | 1973-03-02 | ||
| US4275988A (en) * | 1978-12-18 | 1981-06-30 | Kalashnikov L F | Axial or worm-type centrifugal impeller pump |
-
1979
- 1979-03-19 SU SU792739411A patent/SU1023138A1/en active
-
1980
- 1980-03-07 JP JP55500898A patent/JPS6046276B2/en not_active Expired
- 1980-03-07 DE DE3038735T patent/DE3038735C2/en not_active Expired
- 1980-03-07 WO PCT/SU1980/000042 patent/WO1980002058A1/en active Application Filing
- 1980-03-07 GB GB8035247A patent/GB2058923B/en not_active Expired
- 1980-03-18 CA CA000347916A patent/CA1141232A/en not_active Expired
- 1980-03-19 FI FI800848A patent/FI67435C/en not_active IP Right Cessation
- 1980-03-19 FR FR8006114A patent/FR2452016A1/en active Granted
- 1980-11-06 SE SE8007808A patent/SE444351B/en not_active IP Right Cessation
- 1980-11-11 DK DK479080A patent/DK479080A/en not_active Application Discontinuation
Also Published As
| Publication number | Publication date |
|---|---|
| SU1023138A1 (en) | 1983-06-15 |
| GB2058923A (en) | 1981-04-15 |
| FI67435C (en) | 1985-03-11 |
| FI67435B (en) | 1984-11-30 |
| GB2058923B (en) | 1983-04-20 |
| DE3038735T1 (en) | 1981-04-23 |
| SE8007808L (en) | 1980-11-06 |
| WO1980002058A1 (en) | 1980-10-02 |
| DE3038735C2 (en) | 1986-02-27 |
| JPS6046276B2 (en) | 1985-10-15 |
| FR2452016B1 (en) | 1985-01-18 |
| JPS56501852A (en) | 1981-12-17 |
| SE444351B (en) | 1986-04-07 |
| FI800848A (en) | 1980-09-20 |
| FR2452016A1 (en) | 1980-10-17 |
| DK479080A (en) | 1980-11-11 |
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