DE2216762B2 - Shaft seal - Google Patents

Description

The invention relates to a shaft seal between an interior to be sealed, in which Liquid is located opposite an outside space, wherein the seal consists of two mechanical seals arranged one behind the other in the sealing path with a between them, filled with the same liquid, under higher pressure than that in the Such a shaft seal is from DE-GM 19 70 853 known

One of the main factors affecting the lifespan of equipment that will operate in liquids, e.g. B. water, are immersed, such as B. Tauchmctoren, limit is the pumping in of water at the shaft seal, which is caused by a slight eccentricity in the mechanical seal which is commonly used for minimal leakage on the shaft. The pumping in of water can be avoided by increasing the pressure of the sealing liquid, e.g. B. that typically in the The oil contained in the engine is set significantly higher than the pressure of the surrounding water. These higher pressure differentials necessarily require reinforcement of the motor housing and the like significant changes in the spring-actuated diaphragm usually used to generate the Oil / water pressure differential is used.

Furthermore, in connection with the pressure increase within such devices, there are a number of different seal configurations that allow the pumping in of Prevent water in spite of a small oil / water pressure difference. For example, rotor shafts are below Using an external sealed pump has been used to generate large and small pressures in sealing spaces to generate, which were arranged at axially offset points along the shaft. Similarly, the The shaft of a centrifugal pump can be sealed by reducing the rotational speed of the shaft is used to pump oil from an axially outer location to an inner seal in order to protect the Limit gas leakage from the pump.

The object on which the invention is based is to provide a shaft seal of the initially mentioned to train said type for a rotating machine so that the required pressure in the barrier space to prevent the penetration of liquid outside the device with simple means can be set up without further effort.

This object is achieved according to the invention in the case of a shaft seal of the type mentioned at the beginning solved that in the sliding surface at least one of the sliding rings of the mechanical seals adjacent to the interior space for conveying liquid from the interior space into the barrier space formed viscosity grooves are arranged, which extend from the circumference of the sliding surface lying against the interior space over a radial part the same extend, with a circumferentially closed sliding surface area remains.

Further advantageous exemplary embodiments emerge from the subclaims.

The advantages that can be achieved with the invention are, in particular, that automatic re-pumping of barrier fluid takes place in the sealing chamber. This is done without any additional equipment easy attachment of grooves on seals that are required anyway. Nevertheless occurs with the Shaft seal according to the invention no pressure drop even at standstill.

The invention will now be explained in more detail with reference to the following description and the drawing of exemplary embodiments.

Fig. 1 is an enlarged cross-sectional view

a shaft seal according to an embodiment of the invention.

F i g. Figure 2 is a plan view of part of the interior Mechanical seal and shows the arrangement of the spiral viscosity grooves in it.

F i g. 3 is a cross-sectional view of a shaft seal according to another embodiment of FIG Invention.

Fig. 4 is an enlarged cross-sectional view and shows the force distribution along the seal according to FIG.

5 shows a further embodiment of the Invention with a device for monitoring the seal.

Fig. 6 is a cross-sectional view of the inner Shaft seal in which the pump-out rate of this seal in the event of a drop in pressure in the barrier chamber severe leakage of the outer seal is reduced.

F i g. 7 is a plan view of the fixed part; which the seal according to FIG. 6 forms

In Fig. 1 shows a shaft seal according to an exemplary embodiment. It generally contains an inner mechanical seal 12 provided with spiral-shaped grooves and an outer mechanical seal 14, which are arranged one behind the other on a shaft 16 of a dynamoelectric machine. Such a dynamoelectric machine can be a pump motor, for example. The pump motor usually contains a barrier liquid in the interior 20, such as. B. transformer oil, which is biased by a flexible membrane to increase the oil pressure by about 0.35 kp / cm ² relative to the water 24, which forms the environment for the motor during operation. The oil inside the motor is with the radial outer surface of the inner seal 12 in connection and is pumped through the spiral grooves of the inner mechanical seal into the substantially closed annular barrier space 26 whereby the oil pressure of the barrier space is increased relative to the oil pressure within the pump motor.

The mechanical seal 12, which is provided with spiral grooves, has a sliding ring 28 made of carbon, which is mounted on the shaft 16, the flat end face 30 of the sliding ring opposite a flat End face 32 of a stationary part 34 made of ceramic is arranged, which on the motor housing 36 is rigidly attached. One of the flat end faces of the mechanical seal 12, which is shown in FIG. 1 as the face 32 of the fixed part 34 is shown, has spiral grooves or grooves 22 to at a Rotation of the sliding ring 28 relative to the stationary part 34 to oil from the engine in the sealing space pump. The grooves or grooves shown in FIG. 2 clearer are shown run obliquely in order to achieve a pumping effect. Your geometric configuration and Density depend on the quantity of the pumping action desired by the mechanical seal, what will be explained in more detail below. Desirably they extend in a radial direction from the outer peripheral edge of the annular fixed part 34 to an annular ridge 38, the the grooves or grooves from one arranged in the middle, in the axial direction through the part 34 running opening 40 separates. If there is a fault in the outer mechanical seal 14, the Should make the motor necessary, the web 38, which forms a circumferentially closed sliding surface area, advantageously acts as a barrier to prevent the water from flowing back through the spiral grooved mechanical seal IZ

The end face of the sliding ring 28, which is remote from the flat end face 30, has an incision provided to form a lower shoulder 42 which, together with a support plate 44, serves to support a To form a receiving groove for an O-ring 46, which seals the sliding ring against the shaft 16. A second

ιυ shoulder 48 is also along the radially outer Peripheral surface of slip ring 28 is formed to seat a generally L-shaped locking ring 50 to form from brass, which is biased by a spring 52 relative to the sliding ring 28 To allow axial movement of the locking ring 50 along the shaft 16 while restricting rotational movement of the locking ring, an elongated body 54 of the Lock ring slidably held within a guide 56 which is rigidly attached to the motor shaft

The outer mechanical seal 14 is designed in the usual way and is in relation to the mechanical seal 12 mounted and sized so that the pressure of the oil within the chamber 26 has the tendency that Sliding ring 56 against the opposite ceramic, fixed part 58 to press. A bias spring 60 increases the pressure that the mechanical seal tries to close by exerting an axial force on the upper extension 62 of a locking ring 64 exerts around the inwardly protruding support plate 66 against the

jo slide ring 56 to press. The edge of the locking ring 64 closest to the mechanical seal 12 extends through a guide 68 therethrough to limit the rotation of the locking ring while the sealing of the Sliding ring opposite the shaft by a flexible one Bellows 70 is brought about between the Shaft and the locking ring above is firmly attached.

To prevent solid contaminants from entering the motor, the motor shaft is outside the outer mechanical seal 14 a sand protection seal 71 attached, which the axially outer edge of the outer seal carrier 73 overlaps. The seal carrier is attached to the motor housing 36 by bolts 74 rigidly attached through suitable openings in the

Pass the outer flange of the seal carrier through it,

while a radially inner recess in the

Seal carrier as a seat for a fixed part

58 of the mechanical seal 14 is used.

During operation of the submersible motor is at

a rotational movement of the sliding ring 28 relative to the spirally grooved fixed part 34 oil from the Motor housing is pumped into the annular barrier space 26 to raise the oil pressure within this space to raise a level that prevents the ingress of water by the seal 14 excludes this depends mainly from the Wassereinpumpkraft on the outer mechanical seal 14, which is of a Eccentricity is caused in the outer seal. The resulting pressure level can according to

bo an article titled “Inward Pumping in Mechanical Face Seals «by}. A. Findlay in Paper No. 68 at Lub 2 ASME-ASIE Lubrication Conference in Atlantic City, New) ersey, October 1-10, 1968, calculated using the following formula

hl be:

Ip 3 / cos \ ·

c / r (1

In it is:

- - the required oil pressure in

kp Ευ? '

^Γι1Γ ' ^c -

because I cm eccentricity (e) of the outer mechanical seal 14,
is the maximum tilting movement that is expected for the mechanical seal 14 is the shaft speed in revolutions per second,

is the radial distance between the adjacent end faces forming the seal 14 in cm,

cos \ is the maximum displacement that is expected for the mechanical seal Ϊ4.

R ₁ , -

I ρ. Rj-τ h ³

(2)

In it is:

10

π is the viscosity of the water in - ~ jr-.

believed to penetrate the sealing interface

/ j is the average oil film thickness between the face of the seal in cm.

Usually an oil pressure increase of about 280 kp / cm ^{2 is} necessary to compensate for 2.5 cm of shaft eccentricity in order to ensure that the pumping effect on the outer mechanical seal is zero.

Although the seal eccentricity may vary depending on various factors such as e.g. B. the shaft runout under load and speed, the out-of-roundness of the carbon disc, etc., the total eccentricity can generally be estimated with a high degree of reliability for a given construction. If experience _{J5 has} shown that an eccentricity of about 0.25 mm is normally not exceeded in the case of the mechanical seal, the pressure required for the blocking space 26 is calculated accordingly according to known equations , 25 mm and a calculated total oil pressure of 2.80 kp / cm ² . This pressure must exist in the sealing space 26 in order to prevent it from being pumped into it.

The pumping out speed must be at the., ■ -, outer mechanical seal 14 should be considered to ensure that the oil supply in the engine is at the attempt to prevent the entry of water through the mechanical seal, not in one The pumping speed for the outer seal is exhausted for a relatively short time therefore calculated, for example, from the following formula:

55

60

q is the pump-out speed,
I ρ is the pressure difference across the mechanical seal 14 in kp / cm ² .
/ 1 is the average film thickness between the opposite end faces of the seal in cm,

Ri is the radius to the inner edge of the sealing web.

η is the viscosity of the oil in the seal

interface in kpsec / cm ² and
I R is equal to the radial distance in cm of the opposite end faces that form the seal.

The optimum pressure for the barrier space 26 is then chosen as a compromise between the higher one Oil pressure desired to overcome the pumping of water into the engine and the lower Oil pressure that is desirable to limit the oil pump-out speed at the outer mechanical seal is.

Once the pressure desired for the annular barrier space 2 (has been selected, the geometrical configuration of the inner spiral-shaped grooved mechanical seal 12, which is required to generate this pressure, can be determined according to an article by EA Muijderman "Spiral Groove Bearings", published in 1966 by the For example, a spiral grooved mechanical seal for a 30 cm submersible motor with a rotating shaft of approximately 5 cm has ten equally spaced grooves cut to a depth of 0.033 mm and at a helix angle of 15 ° with a ratio in front of groove land to groove width of 1. The inside and outside diameters of the seal were 4.75 cm and 6 cm, respectively, while the inside diameter of the groove was 4.96 cm a speed of 30 revolutions per second, a maximum pressure was 4.75 kg / cm ² in the annular Lock space 26 determined.

F i g. 3 shows an improved embodiment in which the force of the pretensioning spring of the spiral-shaped grooved mechanical seal can be reduced without reducing the available pressure. Urr to achieve this result, a shoulder cut 72 on Einei relative to the O-ring 46A located outward position in the sliding ring 28/4, whereby pressure oil inside stop of the annular Sperraums 26A with the end surface 7i is connected and the slide ring 28.4 hydrauliscr in the axial direction in a mutual contact mi the fixed part 34Λ can press when dei pressure within the Sperraumes increases also who is a shoulder provided in the shaft 16Λ 76 un seal the cut slide ring 284, the location unc the supporting bracket for the spring GOA has been slightly modified, the mechanical seal, which exerts a pumping action, is shown in FIG. 3 otherwise the one shown in FIG. 1 is essentially the same. The increased axial force on the sliding ring 2BA, which results from the hydraulic pressure on the end face 75, however, reduces the gap or the interspace of the spiral-shaped grooved mechanical seal, which increases the pressure available from the seal. This increases! Pressure in turn leads to an increased hydraulic! Force on the end face 75, and the pressure within the locking space 26.4 is increased more or less selbstandij until an equilibrium pressure is reached

If the resulting flow occurring at the outer mechanical seal 14 / is assumed to be zero, the pressure generated by the mechanical seal 12 /, which is represented in FIG. 4 by the pressure diagram P , increases approximately linearly from the outer circumference of the stationary part 34A to internal expansion of the grooves in the fixed part

ie from di to d ₂ , the pressure remaining constant along the non-grooved section of the sealing interface, ie from d ₂ to d \ at Pmax. For the sake of simplicity, the average pressure that is effective over the area between d ₂ and _{d 3 can} be assumed to be 1/2 Pmax. The maximum pressure at equilibrium can therefore be estimated from the approximation formula:

"max ~~ r- / j2 ι J2v -i ■ (3)

In it is:

Fj is the axial load in kp exerted on the seal by the spring 52 A,

d ₂ is the inside diameter of the grooved section of the mechanical seal,

d $ is the outer diameter of the grooved annular section of the mechanical seal and

i / ₄ is the diameter of the hydraulic shoulder 72, which is formed in the sliding ring 28 Λ.

A more or less automatic seal with an inner diameter of the seal of 4.75 cm, an inner groove diameter d ₂ of 4.95 cm, an outer diameter of the seal ds of 6.03 cm and a compensating diameter of the seal dt of 5.08 cm produced a hydraulic load of 2 kp / cm ² on the mechanical seal in addition to a preload of 2.95 kp / cm ² , which was generated by the spring 52A , so that a total axial load on the mechanical seal of about 5 kp / cm ² arose.

The spiral-shaped grooved running surface of the mechanical seal contained 15 equally spaced Grooves arranged at a helix angle of 15 ° and the tread was rotated at a speed of about 30 revolutions per second

If the required pressure for the barrier space is small, for example 1.4 kgf / cm ² , the outer mechanical seal can be arranged back-to-back with the inner grooved mechanical seal, as shown in FIG. 5 is The pressure within the barrier space 265 then exerts an axial force on the outer mechanical seal 145, by means of which it is more heavily loaded in the opening direction, whereby a preloading spring 605 with a sufficient axial force is required to overcome the hydraulic pressure within the space 265 so that the desired opening of the outer mechanical seal is maintained during operation. If an error were to occur in the pressure within the space 265, the hydraulic force trying to keep the outer seal open would disappear and the bias of the 60S would attempt to close the faces of the outer seal, causing water to enter the Engine is prevented. If such an arrangement is used at relatively high sealing pressures, for example at pressures of about 4.25 kgf / cm ² , care must be taken that a biasing spring 60.5 is selected with sufficient force to prevent excessive pumping of oil through the to prevent external mechanical seal through.

According to a further exemplary embodiment, the sealing effect at an external location can be achieved by the Arrangement of a pressure switch 77 within the blocking space 265 can be monitored, as shown in FIG. 5 is shown. The pressure switch is provided with an alarm device 78 and a voltage source such as for example a transformer 80, connected in series. The primary winding SOA of the transformer is connected in parallel to the conductors (not shown) for the motor excitation. This pressure switch has the function of being connected in series in the event of a pressure drop

ίο within the barrier space 265 under a predetermined one Minimum close. The alarm device 78 is then activated and allows a shutdown and removing the engine from the diving position before permanent damage to the interior of the engine occurs causing water to flow in. Should the pressure drop in the space 265 by a Failure of the outer mechanical seal 145 will be generated, the ingress of water through the grooved mechanical seal 125 through during the Shutdown by the annular web 38 of the Mechanical seal prevents to be effective as a flow restriction during engine shutdown function, the annular ridge should have a radial extension of at least 1 mm.

In Figures 6 and 7 is an automatically acting Apparatus for reducing the pumping action of the inner face seal shown, in which a spring valve 82 is used to control the operation of the inner spiral grooved seal 12C in the case of a Failure of the outer seal from the film contact to convert it into a body contact. The spring valve 82 has the effect of the oil flow from a shaft seal during normal pumping operation Annular groove 84 at the radially inner end of the spiral grooves 22C is arranged to one Shut off bypass opening 86. If on the outer mechanical seal during engine operation an error should occur, thereby reducing the pressure within that formed between the mechanical seals Locking space 26C is reduced, the hydraulic pressure on piston 87 of valve 82 also drops, that with the blocking space 26C via an axial opening 88 is in communication, and the relatively higher pressure of the oil within the annular groove 84 overcomes the Biasing the spring 90 to relieve the pressure at the seal interface via the bypass port 86 to relax.

When valve 82 is open, it spirals grooved mechanical seal 12C from a common one Thick film mode of action, d. H. a movie of more than about 1 μίτι, which is usually at a groove depth of 10-15 μΐη is generated to a solid-solid contact mode, i.e. to a film thickness of less than about 0.5μΐτι, whereby the Auspumpge speed of the mechanical seal used essential is restricted. Thus, part of the oil pumped through the seal becomes the suction side of the Seal returned, thereby reducing both the between the end faces of the grooved mechanical seal generated maximum pressure as well as the amount of oil pumped into the barrier space 26C can be reduced.

It should be noted that the grooved mechanical seal will self-assemble without the action of the spring valve 82 if the outer seal fails tries to close due to the increased relative pressure at the sealing interface, resulting in a changes in the oil distribution at the sealing interface

Seal is required that exceeds the maximum pumping capacity of the seal, the seal switches naturally from film contact to body contact, so that the pump-out speed is reduced will. Thus, by carefully selecting the design of the spiral groove, i. H. the width, depth and Length of the spiral groove, a seal can be made with proper operation of the outer seal 14C generates the desired pressure in the event of film contact, while in the event of failure of the outer Seal a significantly reduced pump-out speed is generated.

The formula (3) given above for estimating the maximum pressure increase clearly shows the influence of the shoulder dt according to FIG. 4 on the generated pressure. In formula (3), there was an average pressure

-ψ- above that section of the sealing limit-

assumed area where the pressure changes. Another theoretically exact formula for calculating the maximum pressure increase can be obtained by integrating the assumed linear pressure increase over the area between the diameters di and cfe according to FIG. If the seal diameter dt is equal to the diameter d \ , there is no shaft shoulder and the formula for calculating the pressure increase takes the following form:

'(JIiI .V -'

10

F ₅ is the total force in kp exerted on the seal by the preload spring,

di is the distance from the radially inner face of the sliding seal to the shaft axis,

</ ₂ is the distance from the radially inner end of the spiral grooves to the shaft axis and

d ₃ is the distance from the radially outer circumference of the spiral grooves to the shaft axis.

The ratio of the maximum pressure at the pumping shaft seal in the event of a leak in the The interface of the outer seal 14 is developed relative to the maximum pressure exerted by the seal can be developed without a leak can then be calculated from the following formula:

In it is:

P ", _{ax is} the maximum pressure in cm ² that is generated by the seal in the event of full film contact without leakage,

P, "". _v , with leakage
P _11111x , without leakage

I +

1 +

^ Λ + ΤΛ

<! ι Λ _ d ₃ V

Ί
d

From this ratio, the maximum pressure that can be exerted by the seal 12 without restricting the leakage developed on the outer seal can be calculated to give an indication of the film thickness supply that arises from the pressure increase If the calculated maximum film pressure in the event of a leak is the maximum pumping capacity of the seal exceeds that specified in the aforementioned publication by Muijderman can be calculated, changes the operation of the seal in the event of failure of the outer seal from film contact to body contact.

For this purpose 4 sheets of drawings

Claims (7)

Patent claims: 1. Shaft seal between an interior space to be sealed in which there is liquid, with respect to an outside space, the seal consisting of two arranged one behind the other in the sealing path Mechanical seals with one between them, filled with the same liquid, exists under higher pressure than the blocking space held in the interior, characterized in that in the sliding surface (31) at least one of the slide rings (28, 34) of the slide ring seals adjacent to the interior space (20) (12) for pumping liquid from the interior into the barrier space (26) designed viscosity grooves (22) are arranged, which extend from the circumference of the sliding surface lying against the interior via a radial part of the same, with a circumferentially closed sliding surface area (38) remains 2. Shaft seal according to claim 1, characterized in that the viscosity grooves (22) run in a spiral. 3. Shaft seal according to claim 1, characterized in that the mechanical seal (14) adjacent to the outer space is designed such that that they are stronger in the closing direction or in the opening direction with a pressure increase in the blocking space (26) is burdened 4. Shaft seal according to claim 1, characterized in that the middle! (78,80) for measuring the Pressure within the locking space (26) and remote warning are provided on a Address pressure drop in the barrier space below a predetermined minimum. 5. Shaft seal according to one or more of claims 1 to 4, characterized in that the radially inner region of the Viskosilätsnuten via a bypass line (86) with the interior space (20) is in communication and in the bypass line a by a spring (90) and the pressure in the barrier chamber (26c) is applied in the closing direction and is arranged by the pressure in the viscosity grooves (22c ^ in the opening direction and dimensioned in such a way that at a certain pressure drop in the barrier chamber (26c) the liquid pumped by the viscosity grooves in the interior is returned. 6. Shaft seal according to claim 5, characterized in that an annular groove (84) is formed in the sliding surface area, into which the viscosity grooves ( 22c) run and the valve (82) is in communication on the input side with this annular groove. 7. Shaft seal according to claim 6, characterized in that the valve (82) in the non-rotatable The mating ring of the mechanical seal adjacent to the interior is attached to the machine housing (12c,} is arranged.