DE2216762A1 - Self-pressurizing seal for a rotating shaft - Google Patents

Description

Self-pressurizing seal for a rotating shaft

The invention relates to a shaft seal for a rotating machine and in particular a shaft seal in which a high pressure zone of the fluid to be sealed between a spiral grooved face seal and a common face seal is formed. to an inlet of the contaminating fluid into the machine.

One of the main factors affecting the lifespan of being immersed Limiting engines is pumping water in at the

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Shaft seal created by a slight eccentricity in the face seal which is commonly used for minimal leakage on the shaft. Even when pumping in water can be overcome by reducing the pressure difference between the Sealing fluid such as oil, which is typically contained in the engine, and the surrounding water increases significantly , these higher pressure differentials necessarily require reinforcement of the motor housing and the like essential Modifications in the spring-actuated diaphragm commonly used to create the oil / water pressure differential.

Because of the difficulties associated with increasing the oil pressure within the engine, a number are different Seal configurations have been proposed to allow water to be pumped in despite a small oil / water pressure differential impede. For example, rotor shafts have been sealed to large and small using an external pump To generate pressures in seal chambers, which were arranged at axially offset locations along the shaft. In a similar way has hitherto been proposed that the shaft of a centrifugal pump is sealed by increasing the rotational speed of the Shaft, used to pull oil from an axially outward location pump to an inner seal to limit gas leakage from the pump.

It is therefore an object of the present invention to provide a novel create a self-pressurizing seal characterized by low leakage. Furthermore should A seal can be created which is suitable for monitoring at an external point in order to ensure that the seal is functioning correctly to ensure. The invention also includes a self-pressurizing seal in which the pressure of the sealing liquid is used within the seal to increase the mechanical preload of the seal, thereby the pressure available from the seal without excessive

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- 3 - ■■ '

Maximize use of the seal during start-up. In the end the aim is to create a self-pressurizing seal in which automatic closure of the seal occurs causing a drop in seal pressure.

These and other objects of the invention are provided by a self-pressurizing seal for a rotating machine solved, which is characterized by an inner pump seal with an annular running element, which is on a rotatable shaft is attached next to an annular, coplanar and fixed member. At least one element of the juxtaposed Element is provided with helical grooves extending from the perimeter of the limb along the planar face of the link extending to a web to the sealing liquid from the rotating machine in a substantially closed zone to pump. The pumping action of the inner seal increases the pressure of the sealing liquid within the zone relative to the pressure of the sealing liquid within the machine, and along the shaft at an axially outward location (relative to the inner pump seal) are common face seal means arranged to prevent the flow of sealing liquid from the Restrict the high pressure zone in the surrounding water. That I

a sealing liquid having a relatively high pressure only within a zone between the axially offset end seals prevents water from entering the motor, without the need for structural reinforcement of the entire engine housing and without the necessary flexible oil expansion system is exposed to large pressure differences.

The invention will now be based on further features and advantages the following description and the accompanying drawings of various exemplary embodiments.

Fig. 1 is an enlarged cross-sectional view of a self self pressurizing seal according to the invention,

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Pig. Figure 2 is a top plan view of a portion of the inner face seal showing the placement of the helicals therein Grooves.

Figure 3 is a cross-sectional view of a self beneath Pressurizing seal in which the hydraulic pressure of the oil inside the seal is used to make the available Increase sealing pressure.

Pig. H is an enlarged cross-sectional view showing the force distribution along the spiral grooved seal shown in FIG. 3.

Fig. 5 is another seal configuration showing that of the invention Device for monitoring the seal.

Fig. 6 is a cross-sectional view of a spiral grooved Seal at which the pump-out speed of the seal with a decrease in the external seal pressure is reduced.

Fig. 7 is a top plan view of the fixed part making up the seal; according to Flg. 6 forms.

In Fig. 1, there is a self-pressurizing seal 10 shown according to the invention. It generally includes an inner helical and grooved face seal 12 and an outer end seal 14, which in tandem on a Shaft 16 of a dynamoelectric machine are arranged. Such a dynamoelectric machine can, for example, be a Be a pump motor as described in U.S. Patent 2,790,916 is. The disclosure of this patent specification is incorporated into the present application by this reference. Typically the pump motor contains a sealing liquid such as transformer oil 20, which is biased by a flexible membrane to

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the pressure of the oil by about 0.35 kgf / cm (5 psi) relative to the water 24 to increase the environment for the during operation. Engine forms. The oil inside the engine stands with the radial outer surface of the inner seal 12 in connection and is through the spiral grooves of the inner face seal. in the substantially closed annular oil chamber 26 is pumped, whereby the oil pressure of the chamber relative to the oil pressure within the .Pump motor is increased.

The spiral grooved face seal 12 is generally characterized by an annular race 28 made of carbon. which is mounted on the shaft 16, the flat end face of the race against a flat end face 32 of a stationary one Part 34 is arranged made of ceramic, which on the pump motor housing 36 is rigidly attached. One of the flat end faces the end seal 12, which is shown in Fig. 1 as the end face 32 of the fixed part 34, has spiral-shaped Grooves or grooves 22 in order to supply oil from the engine when the race 28 rotates relative to the stationary part 34 pump. The grooves or grooves shown in Fig. 2 more clearly have a geometric configuration and density that depends on the quantity of pumping desired by the face seal depends, which will be explained in more detail below. Desirably they extend in the radial direction from the outer peripheral edge of the annular fixed part 34 up to an annular ridge 38, which the grooves or grooves of an opening 40 arranged in the center and running in the axial direction through the part 34. If there is an error in the Outer face seal 14 should it be necessary to switch off the motor, the web 38 advantageously acts as a barrier against a Rückströmurtg of the water through the spiral grooved face seal 12.

The end face of the carbon raceway removed from the flat end face 30 28 is provided with an incision to a lower one

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To form shoulder 42, together with a support plate 44 to this serves to form a housing for an O-ring 46, which is the carbon race with respect to the shaft 16 seals. A second shoulder 48 is also along the radially outer face of the carbon race 28 formed to seat a generally L-shaped Locking ring 50 made of brass, held by a spring is pretensioned with respect to the raceway 28. To allow axial movement of the locking ring along the shaft 16 while a Movement of the locking ring is limited in a plane perpendicular to the shaft is an elongated body 54 of the locking ring slidably held within a guide 56 attached to the motor shaft is rigidly attached.

The outer face seal 14 is formed in a conventional manner and is in tandem with the spiral-shaped grooved face seal 12 mounted so that the pressure of the oil within the chamber 26 has the tendency to the coal race 56 against the opposite ceramic, fixed part 58 to press. A preload spring 60 increases the oil pressure that closes the front seal tried by an axial force against the upper extension 62 of a locking ring 64 is delivered to the inwardly protruding To press the support plate 66 against the carbon race 56. The the Stirndi-'Chtung 12 closest edge of the locking ring 64 runs through a guide 68 to limit axial movement of the locking ring while sealing the race against the shaft is brought about by a flexible bellows 70, which is fixed between the shaft and the locking ring above it is appropriate.

To prevent solid contaminants from entering the engine, a sand slinger is attached to the motor shaft at an axial point around the axially outer edge of the outer seal carrier 73 to be enveloped. The seal carrier is on the motor housing 36 rigidly attached by bolts 74, which through suitable openings

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in the outer flange of the seal carrier, while a radially inner recess in the seal carrier serves as a seat for a fixed part 58 of the end seal 14.

During engine operation, the rotation of the coal runner 28 relative to the helically grooved stationary portion J> k pumps oil from the engine housing into the annular oil chamber 26 to increase the oil pressure within the chamber to a predetermined level. This depends mainly on the anticipated water pumping force on the outer face seal Ik , which is caused by an eccentricity in the outer seal. This predetermined pressure level can be determined according to an article entitled "Inward Pumping in Mechanical Paee Seals" by JA Pindlay in Document No. 68 at the Lub 2 ASME-ASIE Lubrication Conference in Atlantic City, New Jersey, from Jan. October 1968 can be calculated using the following formula:

^ 3 £ cos oc '/ u GO (R- R,)

- 5 ö \ J-}

Where: ■ ■■ - the required oil pressure in e 0, ⁷ cm O

for every 2.5 cm (inch) eccentricity (s) of the outer End seal 14,

£! is the maximum tilt that is expected for the face seal m,

CL? is the shaft speed in radians per second,

RR. is the radial distance between the adjacent end faces forming the seal Ik in FIG. 2, j? cm (inch),

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the maximum displacement that is expected for the end seal 14,

/ U is the viscosity of the water assumed by wii
in

will penetrate the sealing interface, 0.455 kg / gec.

6755 cm ² »

h is the average oil film thickness between the Seal face in 2.5 cm (inches).

2 Usually an oil pressure increase of around 281 kp / cm (4000 lbs / sq.inch) for every 2.5 cm (1 in.) of shaft eccentricity to be compensated to ensure that the pumping effect on the outer face seal is zero.

Even if the seal eccentricity depends on different Factors can vary, such as the shaft runout under load and speed, the out-of-roundness of the carbon disc, etc., so can but generally the total eccentricity is estimated with a high degree of confidence for any given manufacturing process will. Thus, if manufacturing experience has shown that an eccentricity of about 0.25 mm (0.010 inch) would normally be is not exceeded when the end seals are produced, the pressure required for the chamber 26 to allow pumping in to prevent water, calculated by the maximum

ρ observed eccentricity with pressure per 2.5 cm (square inch) the face seal eccentricity is multiplied, which is calculated from the aforementioned Findlay equation. For example results / results for an empirically determined maximum eccentricity of about 0.25 mm and a calculated oil pressure of 281 kp / cm per 2.5 cm (inch) eccentricity a total pressure of 2.8 l kgf / cm. This pressure must exist in the oil chamber 26 in order to be pumped into it to prevent.

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The pumping out rate at the outer face seal lh must also be considered to ensure that the oil supply to the engine is not depleted in a disproportionately short time in an attempt to prevent water from entering through the face seal. The pumping out speed for the outer seal is therefore calculated, for example, from the following formula:

_q .

/ u

A R

Where: q is the pump-out speed

Difference in pressure across the face seal 14 in 0.07 kp / cm ² ,

h is the average pilm thickness between the opposite end faces of the gasket in 2.5 cm (Customs),

R. is the radius to the inner edge of the sealing web,

αχ is the viscosity of oil in the Dic_Jitungslimit-

ρ
area in 0.07 kp / cm and

R is equal to the radial distance of the opposing end faces forming the seal in 2.5 cm (Customs).

The optimal pressure for the oil chamber 26 is then considered a compromise chosen between the high oil pressure desired to overcome the pumping of water into the engine and the low oil pressure, which is desirable to limit the oil pump-out speed at the outer face seal.

Once the pressure desired for the annular oil chamber 26 Has been chosen, the geometric configuration of the inner spirally grooved face seal 12, which is used to generate

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of this pressure is required, according to an article by E.A. Muijderman to be determined. This article is entitled "Spiral Groove Bearings" and 1st published in 1966 by the Philips Technical Library. A helically grooved face seal configuration, which has been found suitable for a 30 cm submersible motor with a rotating shaft of approximately 5 cm, was characterized by 10 equally spaced grooves extending to a depth of 0.033 mm (0.0013 in.) were cut and ran at a helix angle of 15 ° with a ratio of groove land to groove width of 1. the The inside and outside diameters of the gasket were 4.75 cm and 6 cm (1.87 and 2.37 inches), respectively, while the inside diameter of the groove Was 4.96 cm (1.95 inches). With the seal specified above, which rotates at a speed of 30 revolutions per second

a maximum pressure of 4.75 kg / cm (67.5 psi) in the annular oil chamber 26 is observed.

Fig. 3 shows an improved embodiment of this invention, in which the force of the biasing spring of the helically grooved face seal can be reduced without the Pressure available from the face seal is reduced. To achieve this result, a shoulder 72 is on one relative to the O-ring 46 A outside located point in the carbon race 28 A cut, whereby pressurized oil within the annular oil chamber 26 A with the end face 75 in connection stand and the race 28 A hydraulically in the axial direction in mutual contact with the fixed part 34 A can press when the pressure inside the oil chamber rises. Even if a shoulder 76 is provided in the shaft 16 A to seal the cut race 28 A, and the location The support bracket for the spring 60 A has been slightly modified, but the end seal that puts itself under pressure is 3 otherwise essentially identical to that in FIG. 1 seal shown. The increased axial force the race 28 A, which results from the hydraulic pressure on the face

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area 72 results, but reduces the gap or space the spiral-grooved face seal, whereby the from the pressure available to the seal is increased. This increased pressure in turn leads to an increased hydraulic force on the face 75, and the pressure within the annular oil chamber 26 A is increased more or less independently (bootstrap fashion), until an equilibrium pressure is reached.

If the resulting flow occurring at the outer end seal 14 A is assumed to be zero, the pressure generated by the spirally grooved end seal 12 A, which is represented in FIG. 4 by the pressure diagram P, increases approximately linearly from the outer circumference of the fixed part 34A to internal expansion of the grooves in the fixed part, ie from d to d _{2 , the pressure remaining} constant along the non-grooved section of the sealing interface, ie from dp to d ^ j at Ρ ΜΑχ. For the sake of simplicity, the average pressure that is effective over the area between dp and d, can be assumed to be 1/2 P ^ aX. The maximum pressure at equilibrium can therefore be estimated from the approximation formula:

F.
MAX ~. r, "2. _Λ 2. ^KD>

Γ ( V * ^) _ä 2l

L - ₂ % J

In it is:

F _{g is} the axial load exerted on the seal by the spring 52 A in 0.45 kp,

d ₂ is the inner diameter of the helically grooved annular section of the end seal,

d ^ is the outer diameter of the spirally grooved, annular section of the end seal and

djj is the diameter of the hydraulic shoulder 72, the is formed in the carbon race 28A.

A more or less automatic seal (bootstrap seal) with an inner diameter of the seal (ie d ₁ ) of 4.75 cm

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(1.87 inches), a groove inside diameter (ie dp) of 4.95 cm (1.95 inches), an outside diameter of the gasket (ie d,) of 6.03 cm (2.374 inches), and a compensating diameter of the gasket (ie djj) of 5.08 cm (2.0 in) created a hydraulic load of 2.02 kg / cm (28.7 lbs) on the end seal in addition to a preload of 2.95 kg / cm ² (42 lbs), generated by the spring 52A so that a total axial load on the end seal of about 5 kg / cm ² (70.7 lbs) was created.

The helically grooved running surface of the face seal contained 15 equally spaced grooves, which under one Spiral angles of 15 ° were arranged, and the tread was rotated at a speed of about 30 revolutions per second.

If the required pressure is oil between chamber small for the, for example, 1, ¹ Jl kp / cm (20 psi), the outer end seal may back are arranged back to back with the inner helically grooved face seal, as shown in Fig. 5. The pressure within the oil chamber 26B then exerts an axial force on the outer face seal I ¹ IB which attempts to separate the opposing end faces of the seal, thereby requiring a biasing spring 6OB with sufficient axial force to overcome the hydraulic pressure within the chamber 26B, so that the desired opening of the outer face seal is maintained during operation. If an error were to occur in the pressure within chamber 26B, the hydraulic force trying to hold the outer seal open would disappear and the biasing spring 6OB would tend to close the faces of the outer seal, preventing water from entering the engine is prevented. If the back-to-back assembly is used at relatively high seal pressures, for example at pressures of about 4.25 kgf / cm (60 psi), care must be taken to ensure that a bias spring 60B is selected with sufficient force to prevent excessive pumping of oil through the outer face seal

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to prevent through.

A key feature of this invention is that the sealing action at an external location can be monitored by the placement of a pressure transducer 77 within the oil chamber 26B, as shown in FIG. The pressure transducer is connected in series with an alarm device 78 and a voltage source such as a transformer 80, the primary winding 80A of which is connected in parallel with the conductors (not shown) for the motor excitation. This pressure transducer has the function of closing the series circuit in the event of a pressure drop within the oil chamber 26B below a predetermined minimum. The alarm system 78 is then activated and allows the engine to be switched off and removed from the submerged position _} before permanent damage to the interior of the engine is caused by an inflow of water. Should the pressure drop in the chamber 26B be created by failure of the outer face seal 14b, the entry of water through the helically grooved face seal 12B is prevented during shutdown by the annular ridge 38 of the face seal. To while the engine is shut-off, which occurs due to the failure of the outer seal 14B to function effectively as a flow restrictor, the annular web should desirably have a radial extent of at least 1 mm (0.0 ¹ I inch).

In Figs. 6 and 7 a self-contained motor protection device is shown, in which a spring-operated pressure relief valve is used to control the operation of the inner spiral grooved seal 12C in case of failure of the outer seal and a hard film in ^I F% stkörper solids -Convert touch type. The relief valve 82 acts, during normal operation of the self-pressurizing seal, to block the flow of oil from an annular groove 84 formed at the radially inward end of the spiral groove 22C.

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is arranged to limit a secondary opening 86. If a fault occurs on the outer face seal while the engine is running should, whereby the pressure within the oil chamber 26C formed between the end seals is decreased, also falls the hydraulic pressure on the piston 87 of the valve 82, which communicates with the oil chamber 26C via an axial opening 88 stands, and the relatively higher pressure of the oil within the annular Groove 84 overcomes the preload of the spring 90 to relieve the pressure at the seal interface via the auxiliary port 86.

When the valve 82 is open, the mode of operation shifts of the helically grooved face seal 12C of conventional thick film action, i.e., a film greater than about 1 ^ in (100 microinches), which is usually the case with a usual grooving depth of 10-15 ym (1000-1500 microinches) is produced, too a solid-solid contact mode, i.e., a film thickness of less than about 0.5 µm (50 microinches), which increases the pumpdown rate the spiral-shaped grooved face seal / limitc. Thus, some of that is pumped through the face seal Oil is returned to the suction side of the face seal, causing both the between the faces of the spirally grooved The face seal generated maximum pressure as well as reducing the amount of oil pumped into the oil chamber 26C.

It should be noted that the spirally grooved face seal tries to close itself without the action of the relief valve in the event of failure of the larger seal due to the increased maximum pressure at the sealing interface, which leads to a changed oil distribution at the sealing interface. If the maximum pressure that is required for a leakage for the seal exceeds the maximum production capacity of the _V seal, the seal turns of natural order from a Vollfilm- on a solid-solid-contact type, the pumpdown reduced. Thus, through a

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careful selection of the design of the spiral groove, i.e. the width, Depth and length of the spiral groove, a seal can be made at which the desired pressure is created, with the outer 1 * JC seal works properly with a full film type while with a failure of the outer seal produces a significantly reduced pumping speed.

The above formula (3) for estimating the maximum The increase in pressure clearly shows the influence of the shoulder d ^ 4 on the pressure generated. In formula (3) there was a

average pressure MAX assumed over that portion of the sealing interface 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 d ¹ and d, according to FIG. 11. If the seal diameter OIL is equal to the diameter d ^, there is no shaft shoulder and the formula for calculating the pressure increase assumes the following formula:

^P MAX / TT

In it is:

the maximum pressure in 0.07 / cm created by the

The seal is created with a full-film contact type without leakage,

F _e is the total force in 0.45 kp that is exerted on the seal by the preload spring, dj is the distance from the radially inner end face of the face seal to the shaft axis, dp is the distance from the radially inner end of the spiral grooves to the Shaft axis and d, is the distance from the radially outer circumference of the

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epiral grooves for

Pas ratio of the maximum pressure, which is at the Pumpdiphtung bet a Leekage at the interface of the outer Piehtung wound, can be developed relative to the maximum Pruek, the v§n the processing without leakage _f can be 4ann calculated from the formula:

with leakage

^F MAX, ^onn ® leakage

From this ratio, the majtimaje pressure that can be developed y * m of device 12 without restriction of leakage at (ίφψ outer seal) can be calculated to provide an indication of film thickness resulting from the pressure increase If the leakage exceeds the maximum generating capacity of the seal, which can be calculated from the above-cited publication by Muijderman, the performance of the seal changes from a full film type to a solid-solid contact type in the event of a failure of the outer seal.

It should be noted that very shallow (e.g. 1/2 ^ m) or very deep (e.g. 200 µm) grooves for the inner face seal can be used to prevent failure of the outer Seal to reduce the pump-out speed. However, since the prints produced by these forehead seals during a normal Operation is difficult to predict due to changes in fluid viscosity at the seal interface such seals are generally not recommended for the inner face seal.

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Claims (9)

- 17 claims 1. Shaft seal for a machine with rotating shaft for Prevention of the entry of a fluid surrounding the machine into the machine, characterized with an internally arranged pump seal (12) an annular running element (28) which rests on the rotating shaft (16) 'next to an annular stationary element (34) is attached, which is arranged opposite the running element (28), at least one of the juxtaposed Elements (28, 34) viscosity grooves or grooves (22 or 22c) which extend from a peripheral edge of the element along a planar end face of the element and terminate in a web (38) so that a sealant (20) contained in the machine in a relatively self-contained Zone (26) is pumpable, whereby the pressure of the sealant within the zone (26) is relative to the pressure of the sealant can be increased within the machine, and by a face seal device (14) which is mounted on the shaft (16) at a is attached relative to the inner seal (12) axially outward location to prevent the outflow of the sealant from the confined zone (26), wherein the front sealing device (14) comprises a rotating element (56), which in axial direction on the rotating shaft (16) is mounted a fixed element (58) which is in a coplanar position is arranged adjacent to the rotating element (56), and has means (60) to the elements (56, 58) in a mutual Press contact. 2. Shaft seal according to claim 1, characterized in that that the inner pump seal (12) and the end seal device (14) in tandem along the shaft (16) are arranged, wherein the pressure of the sealant in the closed zone (26) has the tendency, the "distance between the fixed and rotating elements (56, 58) 209843/0767 - To reduce the size of the outer headlamping device (14). 3. Shaft seal according to claim 1, characterized in that the inner pump seal (12) and the end seal device (I ¹ O back to back along the shaft (16) are arranged, the pressure of the sealant in the closed zone (26) the inclination has to increase the distance between the fixed and rotating elements of the outer face seal device (14) (Fig. 5). 4. Shaft seal according to claim 1, characterized in that the one element (28A) of the inner pump seal along the shaft (16A) is slidably mounted in the axial direction and further has a shoulder (75) which in that end face of the axially sliding Element (28A) which is located away from the fixed element (34A) forming the seal and which is in communication with the sealant within the high pressure zone (26) so that a hydraulic force can be developed with the inclination which the To press the pumping device forming elements (28A, 3 ¹ IA) into mutual contact. 5. Shaft seal according to claim 4, characterized in that that an O-ring seal (46A) is further provided between the rotating element and the shaft which the rotating element with respect to the shaft / can be sealed. 6. Shaft seal according to claim 1, characterized that means for measuring the pressure within the enclosed zone is provided with a remote signal device are that on the pressure measuring means to display a pressure drop in the closed zone below a predetermined Minimum. 209843/0767 7. Shaft seal according to one or more of claims 1 to 6, characterized in that means (80) for monitoring the pressure within the remote zone (26), a valve (82) ,, which with the area between the spirally grooved face sealing elements \ xnä is in connection with the engine interior »and means j [90) with actuation of the valve (82) are provided if. The bridge dep closed zone (26) falls below a predetermined minimum. 8. Shaft seal according to claim 3 »thereby ge kep n ^ - You can see that the inner pump seal (12) is one Face seal with two opposing elements ?, of which one element has spiral grooves or grooves which bit from the periphery of the element to one annular groove or groove (8 * 1) in the face of the egg ^ Mentes extend, and the valve (82) between the annular groove and the suction side of the Pumpelnriehtung is connected so that the flow of the sealant can be regulated therebetween. 9. Shaft seal according to claim 7, characterized in that that the valve (82) is a spring-operated valve which communicates with the enclosed zone (26) stands, the hydraulic pressure of the sealant within the closed zone (26) a force on the valve in exerts a first direction and the hydraulic force of the sealant within the annular groove or groove (8 * 0 exerts a force in a second direction on the valve. 209843/0767