FI66676C - CENTRIFUGALPUMP - Google Patents

Description

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Jttj »M (11) utlAggningsskkift 6667 6 C (« 9 '; / Γ-C 11 1934 lvw C .. u »» * V.' i Γλ 'tf (51) KvJlk (ht.a.3 F Oi. D 29/44 FINLAND — FINLAND (M) 791090 (22) HlfciwMM-Ai »ata * ifrft« 02.04.79 'Fl' · (2¾ ANa ^ aivt — GlMghätadat 02.04.79 (41) T «M« HUnU — 03.10.80 Pstanttl * and Registry Managed · (4¾ mweittHsewM i> hssIJsliaiwn prm

Pm «i6> oek reglsterstyrelsn '' AmMcan« tf * ded »« oakr * · * piMfcanrf 31.07-84 (32) (33) (31) Prr * "* r mmettum» Hirt piortMc (71) (72) Leonid Fedorovich Kalashnikov , ulitsa 50 let VLKSM, 2, sq. 165,

Kaliningrad, Moskovskaya oblast, Vladimir Nikolaevich Kudeyarov, Leningradskoe shosse 2, kv. 40, Khimki Moskovskaya oblast,

Georgy Monasievich Kushnir, Kronshtadtsky Bulvar 34, building 2, sq. M. 78, Moscow, Anatoly Semenovich Shapiro, Raketny Bulvar 12, kv. 7, Moscow, Ryury Ivanovich Konstantinov, ulitsa 50, let VLKSM, 5/16, kv. 110, Kaliningrad, Moskovskaya oblast,

Vadim Vitalievich Nikolaev, ulitsa Mendeleevskaya 5, Klyazma Moskovskaya oblast, Vladimir Kupriyanovich Kunets, Prospekt Mira 180, kv. 12, Moscow, USSR (SU) (74) Oy Kolster Ab (54) Centrifugal Pump - Centrifugal Pump

The invention relates to a centrifugal pump having an axial impeller mounted on a drive shaft in a housing and a centrifugal impeller and comprising a hub with helical impeller blades and an additional axial impeller of the suction head provided with helical blades and mounted on the drive shaft the helical pitch of the vanes is less than the helical pitch of the axial impeller at the entry point.

The invention can be used in the chemical industry, rock oil refining, land improvement, etc., but is perhaps most useful in power plant construction, shipbuilding and aerospace applications, especially high suction pumps designed for low suction pressure, or high speed pumps.

Important performance values of the pump include its suction capacity, i.e. power, which is expressed as a critical cavitation flow rate C, which is explained in more detail below.

66676

The critical cavitation flow rate C can be represented by the following formula: C = 5.62, (1)

Ah3 / 4 where n is the speed of rotation of the pump drive shaft, rpm, Q is the volume flow supplied by the pump, m3 / s and Z \ h (NPSH) is the positive net suction pressure of the pump, m.

The higher the value of C, the better the suction power of the pump.

The speed of rotation of the pump drive shaft affects the external dimensions and mass of the pump, the feed power of the pump affects the number of pumps required and the suction pressure affects the number of pumps required and the capital cost. Thus, for example, doubling the suction power of the pump while keeping the suction pressure constant makes it possible to double the speed of the pump drive shaft, whereby the pump size and mass can be reduced 3 to 6 times, which means lower pump manufacturing costs while maintaining the same suction power. Today, the aim is to increase the unit capacity of power plants by acquiring pumps with ever higher power and higher suction head. Achieving higher suction pressures in high power pumps is limited by unreasonable costs. Thus, for example, doubling the suction power of a pump makes it possible to cope with one high-power pump instead of four pumps with a total corresponding total power. In addition, the capital investment to achieve the corresponding suction pressure will then be at least three times smaller.

The pump industry therefore has a strong need to acquire pumps with higher suction power.

If the suction power of the pump is not sufficient, the pump will become damaged, which will reduce its head and efficiency, cause cavitation corrosion in the impeller flow sections and fluctuations in pressure and fluid flow in the pump inlet and outlet pipes.

A special feature of this problem is that increasing the suction power of the pump usually reduces the efficiency of the pump (<^), which in turn significantly increases the energy consumption. As a result, high suction power pumps usually have low efficiency while high efficiency pumps are known to have low suction power.

3,6676 Pumps with high suction power (C cy 4000) are currently known (see, for example, "Cavitation in vane pumps", Stripling, Tr. ASME Ser. D, No. 3, 1962).

This known pump has an axial impeller arranged on the drive shaft, the hub of which has helical impeller blades, the shape of which along the radius of the impeller can be represented by the formula: r · tgβ = constant.

In this case, r is the running value of the impeller radius and β is the ramp angle of the blade bounded by the plane at right angles to the drive shaft of the pump and the tangential plane of the impeller blades.

The suction power of this pump has been increased by increasing the cross-sectional area of the flow tube and decreasing the rush angle of the impeller blades, resulting in a lower flow factor (Ϋ) at the mouth of the impeller. The flow factor is understood as the relationship between the axial velocity (C) of the liquid flow and the circumferential velocity (U) a of the impeller (measured by the outer diameter of the impeller). In this case, the cross-sectional area of the pump flow pipe has been made larger by increasing the outer diameter of the impeller and decreasing the diameter of the hub as much as is possible from the point of view of the structural strength of the hub. The axial component of the liquid flow rate is thus reduced, and the drop in static pressure in the liquid flow is then minimal, which in turn increases the suction power of the pump.

However, this pump has a low efficiency of 0.5), which is explained by the lower value of the flow factor 0.1).

This, in turn, is due to the larger cross-sectional area of the pump flow tube, the lower axial velocity (C) of the liquid flow, and the differences in the flow rate in the impeller flow tube.

Another already known centrifugal pump again has a high efficiency (^ = 0.75 ... 0.9) (see "Centrifugal and axial flow pumps", AI Stepanov, Mashgiz Publishers, M., 1960, p. 141 ... 164).

This pump has a housing with an impeller mounted on the drive shaft. The blades attached to its hub in the radial direction of the impeller have the shape of the so-called free vortex type. In planar imaging of cylinder sections, the wings form a lattice of aerodynamic profiles at relatively large angles of incidence, defined as an angle of 4 66676 between the profile tendon and the anterior surface of the lattice and corresponding to a higher flow factor (.Ψ> 0.2).

However, the suction power of this pump is low (C <^ 1000), which in turn is due to the relatively high axial velocities (CL) of the liquid flow, because the cross-sectional area of the impeller flow pipe is reduced.

In order to solve a certain opposite problem, i.e. simultaneously aiming for high suction power and high efficiency, a centrifugal pump (DS Patent No. 3,299,821) was developed, the housing of which has an axial impeller mounted on the drive shaft of the pump in front of the centrifugal impeller in the direction of fluid flow. This axial impeller has a hub with impeller blades that form a plurality of wing channels in different directions.

The flow tube of the axial impeller is in two parts, i.e. it has two successive fluid flow compartments, namely a cavitation compartment and a pressure compartment; the rush angles of the wings increase evenly from the mouth of the impeller to the outlet end. In order to keep the axial length of the impeller as small as possible, some theoretical dependencies have been presented to prove the law of the change of the ramp angle of the wing along the impeller, in the direction of fluid flow. These dependencies are intended to meet the requirement of providing a free fluid flow across the width of the vane channels, with the cavitation section of the flow tube guaranteeing a higher suction power and the pressure section guaranteeing a predetermined suction height. With such an axial impeller flow pipe structure, the suction power and efficiency of the pump are simultaneously increased.

Another embodiment of a centrifugal pump is known (see the article "Studies into high-pressure screws having doublerow, blades", by D.N. Contractor and R.I. Atter, in "Hydronautics", Ind., NASA CR-113890, 1969). According to it, an axial impeller with screw vanes is placed on the drive shaft of the pump in front of the axial impeller in the direction of fluid flow, whereby the screw vanes provide high suction power while keeping the pressure to a minimum. 6 6 6 V 6 This arrangement makes it possible to select the designed impeller operating conditions at higher values of the flow factor (> 0.2), which guarantees a high efficiency for the pump.

GB patent publication 1,523,893 discloses a centrifugal pump having an axial impeller with two vane groups mounted before the centrifugal wheel. In this pump, the thread pitch of the first group of vanes in the flow direction is smaller than the thread pitch of the second group of impeller vanes. Such an axial impeller structure makes it possible to take the axial return flow from the impeller into the zone between the first and second wing groups without hydraulic shock in the suction line and to prevent the so-called "hydrulic shock".

However, the above-mentioned structural arrangements are characteristic only of the development that has prevailed so far when it comes to providing a high suction power and at the same time a high efficiency to the pump. However, this problem cannot be considered definitively solved. Namely, as the suction power of the pump increases, the cavitation corrosion associated with the pump flow pipe decreases, and the level of fluid pressure fluctuations and flow rate remains lower in the pump inlet and outlet pipes.

The object of the present invention is therefore to develop a pump structure with a higher suction power, then using an axial impeller at the suction head whose cavitation properties are close to the limit values.

According to the invention, this is achieved in that the outer diameter of the additional axial impeller is smaller than the outer diameter of the axial impeller, and that the ratio of the outer diameters of the additional axial impeller and axial impeller high suction capacity, i.e. suction power.

With this structure, the suction power of the pump can be significantly increased, which is due to the fact that a larger annular space is formed between the outer diameter of the additional impeller and the inner diameter of the pump housing. Therefore, the fluid flow is divided into two flows at the inlet end of the additional impeller. One liquid stream passes through the above-mentioned t 666 and the other through the impeller. It follows from the analysis of clause (1) above that by reducing the volume flow in order for the auxiliary axial impeller to operate without a cavitation corrosion, a lower positive net suction pressure is required at known, predetermined values for the drive shaft speed and critical cavitation flow rate. In the case of the entire pump, when the net suction pressure decreases - using pre-set values for the volume flow of the liquid to be treated and also with a pre-set speed of the pump shaft - the suction power of the pump increases considerably. With simple calculations it is possible to prove that the increase in suction power of a pump can be calculated from the expression: C = C —gr, (2) where C is the critical cavitation flow rate of a pump with an additional axial impeller, C is the external cavitation flow rate D is the outer diameter of the actual axial impeller.

It is a well-known fact that the optimal pitch of the impeller in each axial impeller depends on the outer diameter of the impeller to ensure maximum suction power.

Therefore, starting from the principle of geometric uniformity, the pitch of the auxiliary impeller blades according to the outer diameter of the impeller must be determined to be compatible with po. with a larger outer diameter of the additional impeller.

In addition, the additional impeller creates such a pressure that no cavitation occurs in the impeller, so that cavitation corrosion of the impeller flow tube is less. In this case, no significant fluctuations in fluid pressure and flow rate can occur in the pump.

It is preferred that the outer diameter of the additional axial impeller of the suction head has a constant length along the entire length of the impeller and this outer diameter is 10 to 50% smaller than the outer diameter of the axial impeller, and the thread pitch of the additional axial impeller blades at the inlet end.

Em. the ratios have been obtained experimentally and have been found to be optimum values with the outer diameter of the additional impeller remaining constant at 66676. When the outer diameter of the additional impeller is reduced by less than 10% compared to the outer diameter of the actual impeller, the effect of increasing the suction power of the pump is then much smaller. The fact that the reduction of the additional impeller diameter to 50% is again due to the fact that the additional impeller must have a higher pressure before the axial impeller so that the axial impeller operates without cavitation formation. This pressure drops significantly if the outer diameter of the additional impeller is reduced by more than 50%, which again causes cavitation to form in the pump.

It is important that along the entire length of the additional axial impeller of the suction head, its outer diameter and the pitch of its rails decrease as the fluid flows against the flow. In this case, it must be taken into account that - as shown in expression (2) - the suction power of the pump is maximum when the outer diameter of the additional axial impeller is as small as possible.

The additional impeller can be formed by arranging several basic axial impellers according to the invention in succession. In the direction of fluid flow, each basic impeller is in a way an additional impeller for the next axial basic impeller. Thus, the first basic impeller is required to have the lowest possible NPSH value in order to operate without cavitation formation, while the corresponding operation of the next basic impeller is guaranteed by both the NPSH value and the pressure generated by the first basic impeller, etc.

Cavitation-free pump operation is generally possible with a significantly lower NPSH value determined by the operating conditions of the first basic impeller in the direction of fluid flow.

It is desirable that the pitch of the blades of the axial auxiliary impeller of the suction head be selected on the basis of the following expression: D! + d.

SJ = (0.75 ... 1.25) - p— · S, (3) where S |, D | and d | are the running values of the thread pitch, outer diameter and hub diameter of the axial additional impeller blades of the suction head, and S, D and d are the values of the thread pitch, outer diameter and hub diameter of the axial impeller impeller at the inlet end, respectively.

8 66676

Equation (3) above is essentially a mathematical expression for the geometric equality of all axial basic impellers po. the impellers then together forming an additional impeller. The average diameter of each basic impeller corresponds to its linear dimension. The values of the constant factor, ie the constant (0.75 ... 1.25), have been obtained on the basis of the experiments performed. Said range allows a small deviation from the maximum suction power of the pump corresponding to the constant characteristic number of the unit.

The invention will now be described in more detail in the form of a structural example with reference to the accompanying drawing, in which Figure 1 is a schematic longitudinal section of a centrifugal pump according to the invention with a centrifugal impeller. is an additional suction step, shown in conjunction with a centrifugal impeller.

The pump shown in Fig. 1 has a housing 1 with a liquid inlet pipe socket 2 and a liquid discharge unit formed as a helical chamber 3. In the housing 1, a drive shaft 5 is arranged on the bearings 4. An axial impeller 6 and a centrifugal impeller 7 are arranged on the drive shaft in the direction of fluid flow. The axial impeller 6 has a hub 8 to which are mounted vanes 9 which form wing channels 10 for the flow of liquid. The outer diameter of the axial impeller 6 is D and the pitch of the vane thread line S at the suction end at the outer diameter D. The axial impeller 6 has an additional axial impeller 11 at the suction end in the shaft 5. The additional impeller 11 comprises a hub 12 and screw vanes 13 attached thereto, which form wing channels 14. The outer diameter of the additional impeller 11 and D 'is smaller than the outer diameter 13 of the axial impeller 6. than the helical pitch of the vanes 9 at the inlet end of the axial impeller 6 at the outer diameter D. The outer diameters D 'and D as well as the thread pitches S' and S of the blades (additional impeller 11 and axial impeller 6) are chosen so that a high suction power is obtained for the pump.

The pump shown in the accompanying figures has a ratio of D 'to D and a ratio of S' to S of approximately 0.64 with the outer diameter of the additional axial impeller 11 remaining constant. The following performance values have been obtained in experiments with this type of 9 66676 type pump:

Pump parameters D '/ D C' C C / C

Pump No. 1 0.72 6200 ... 700 4700 0.76 ... 0.675

Pump No. 2 0.64 7000 ... 9000 5200 0.74 ... 0.58

Pump No. 3 0.63 6000 ... 8500 4500 ... 5000 0.75 ... 0.59

Pump No. 4 0.73 5500 ... 7400 4500 ... 5000 0.82 ... 0.68

The results obtained confirm clause (2).

As the drive shaft 5 rotates, the liquid enters the pump through the inlet pipe support 2 and passes to the rotating suction head impeller 11. Part of the liquid passes along the vane channels 14, while part passes to the rotating axial impeller 6 between the housing 1 and the impeller blades 13. The dynamic interaction of the vanes 13 and the fluid increases the pressure of the fluid entering the axial impeller 6. In the running wheel 6, the liquid flows along the wing channels 10. The dynamic interaction of the vanes 9 and the fluid causes an even greater pressure rise of the fluid. The liquid then passes to the centrifugal impeller 7. The liquid coming from the wing channels 10 of the axial impeller 6 also enters the centrifugal impeller 7, where the liquid pressure rises to the desired level. Such a successive increase in fluid pressure ensures that no cavitation is formed in any of the impellers of the pump. The liquid then passes from the impeller 7 to the outlet socket 3 and from there to the pressure piping.

Figure 2 shows yet another pump structure according to the invention. It has the outer diameter of the suction head axial impeller 11 and po. thread pitch of impeller blades 13 S | decrease against the direction of fluid flow. According to the principle of geometric uniformity, the helical pitch S | is selected on the basis of expression (3) so that it fits the outer diameter D of the auxiliary impeller 11 | and po. to the normal values of the diameter of the impeller hub 12.

The operation of the pump in this case corresponds to the operation of the pump shown in Figure 1. The only exception is that the required suction pressure is now lower due to the smaller diameter of the auxiliary impeller 11 at the suction end of the pump and the pressure is higher at the outlet end due to the larger diameter of the auxiliary impeller there.

66676 Thus, due to the meridian section shape of the auxiliary impeller 11, the suction power of the pump is improved and the operation of the pump becomes more reliable because cavitation formation does not occur in the axial impeller 6, the centrifugal impeller 7 or the entire pump.

Figure 3 shows a centrifugal or vane pump in which the additional axial impeller 11 of the suction head is utilized in an additional step. The impeller 11 is fixed to a bearing 15 on a supported rotating drive shaft 5. The bearing 15 is located in the discharge control device 16 between the impellers 11 and 6. The dimensioning of the suction head impeller 11 corresponds to expression (3): D! + d! S! = (0.75 ... 1.25 - ^ · S.

D + d

The operation of the pump otherwise corresponds to the pump of Fig. 2, but the flow rate is now lower because expansions have been made in the vane channels of the device 16. Instead, the static pressure of the liquid increases, which improves the operating conditions of the axial impeller 6 without the formation of cavitation.

The use of an additional stage is particularly advantageous when the pumps already in use are structurally modified in order to improve their suction efficiency.

Claims (6)

1 D + d jossa S |, D | yes d | ovat imupään aksiaalisen lisäjuoksupyörän (11) 12 66676 siipien (13) kierteennousun, ulkohalkaisijän ja navan (12) halkaisijan juoksivia arvoja ja S, D ja d ovat vastaavasti aksiaalisen juoksupyörän (6) siipien (9) kierteennousun arvoja sisääntulopäässä. 13 Patent Claims: 66676 A centrifugal pump having an axial impeller (6) and a centrifugal impeller (7) arranged on the drive shaft (5) in the housing (1) and comprising a hub (9) provided with helical impeller blades (9) and an additional axially aspirated impeller (11). with helical blades (13) and arranged on the drive shaft (5) in the flow direction before the axial impeller (6), the thread pitch (S ') of the blades of the extra axial suction impeller (11) being less than the thread pitch (S) of the blades ( 9) of the axial impeller (6) at the entrance, characterized in that the outer diameter (D ') of the extra axial intake impeller (11) is smaller than the outer diameter (D) of the axial impeller (6), and that the the land between the outer diameters (D1) and (D) of the extra axial intake impeller (11) and the axial impeller (6), and the relationship between the thread pitches (S ') and (S) of the extra axial intake impeller (11) and the axial impeller (6) and at the external diameters of the impeller are selected such that a Eye suction capacity is achieved at the pump. Pump according to claim 1, characterized in that the outer diameter (D1) of the extra axial suction impeller (11) has a constant length in the meridian plane and this outer diameter (D *) is 10 to 50% smaller than the outer diameter (D). of the axial impeller (6), and the pitch (S ') of the blades (13) of the extra axial suction impeller (11) is 10 to 50% less than the pitch (S) of the blades (9) of the axial impeller (6) ) at the input end. 3. Pump according to claim 1, characterized in that over the entire length of the extra axial suction impeller (11) its outer diameter (D 1) and the thread pitch (S 2) of its blade (13) are reduced against the flow of the liquid. Pump according to claim 3, characterized in that the thread pitch (S) of the blades (13) of the extra axial intake impeller (11) is selected in accordance with the following condition: D! + d! S! = (0.75 to 1.25) ——— · S, 1 D + d