TW201833442A - Pump assemblies with stator joint seals - Google Patents

Abstract

A pump assembly is disclosed comprising: two half shell stators defining one or more pumping chambers; end pieces mounted at either end of said two half shell stators; longitudinal seals for sealing between longitudinal contact faces of said two half shell stators on either side of said pumping chamber; and at least one further seal for sealing between one of said end pieces and said stator half shells. The longitudinal seals have end portions that abut against said at least one annular seal; and an aspect ratio of width to height of said longitudinal seals and said further seal is between 1:1 and 2:1.

Description

Pump assembly with stator connection seal

The field of the invention relates to pumps and in particular to seals for a stator of a pump.

By using a half-shell stator, it becomes easier to install a rotor in one of the stators of some pump assemblies. This allows the rotor to be placed in one half-shell and the other half-shell to be mounted on top. A top plate or end piece is then used at either end of the stator to support the bearing and drive system. A seal is needed between the stator half-shells and between the end piece and the stator. The point where the dividing line between the half-shell stators touches the seal that seals between the top plate and the stator is called a T-junction. Especially when the pump assembly is operated at high temperatures, it is difficult to provide an effective or reliable seal at this T-junction. GB2489248 discloses a vacuum pump with a half-shell stator. The vacuum pump includes a longitudinal stator connection seal for sealing between the stator half shells, and a ring-shaped stator seal for sealing between the stator and the end piece. The stator half-shell has a structure for holding the longitudinal seal in place against the annular seal, thereby improving the seal at this T junction.

A first aspect of the present invention provides a vacuum pump assembly including: two half-shell stators, which define one or more vacuum pump chambers; and end pieces, which are installed at either end of the two half-shell stators ; Longitudinal seals on either side of the vacuum pump chamber, which are used to seal between the longitudinal contact surfaces of the two half-shell stators; and at least one further seal, which is used at Sealing is performed between one of the end pieces and the stator half shells; wherein the longitudinal seals have end portions adjacent to the at least one annular seal; and each of the longitudinal seals and the further seal The width-to-height aspect ratio of one of the pieces is between 1: 1 and 2: 1. The inventors of the present invention have recognized that an interface between two different seals will be affected by stress and compression on each seal, and changes in stress and compression on each seal will cause its elastic properties and its deformation. Correspondence changes. In addition, the inventors have recognized that the stress in a seal increases when the width-to-height ratio due to increased deformation increases when a seal is under compression, which changes the seal geometry and causes The risk of interface failure increases. The high stress in a seal also accelerates the onset of compression deformation, which shortens the life of the seal. The geometry of the interface between a seal is particularly affected in cases where the stresses experienced by the seals due to changes in operating conditions, such as a temperature rise, are different in each seal. If the stress on one seal increases from the stress on the other seal, this will result in a stress mismatch that will affect the geometry of the sealing interface and therefore the sealing properties of the interface will change. This problem has been solved by providing a similar, relatively low width to height aspect ratio for each of the seals. This results in a corresponding small change in the interface geometry between the seals and any mismatches in the stresses experienced by the small seals as the operating temperature changes, thereby providing efficient operation across a wide temperature range. In some embodiments, the further seal is an annular seal and includes an O-ring having an aspect ratio of 1: 1 and the longitudinal seal has a rectangular cross section. Due to the geometry of the pump assembly, the seal between the end piece and the stator half shell may have a ring shape and in some cases may include an O-ring. O-rings are effectively sealed and are readily available. The seal generally has a circular cross-section with an aspect ratio of 1: 1. In contrast, the longitudinal seal will have a rectangular cross section that can be a square and will have an aspect ratio between 1: 1 and 2: 1. The longitudinal seal may be in the form of a gasket and the gasket is generally formed with a rectangular cross section having a large width to height aspect ratio. Providing an aspect ratio seal having an aspect ratio similar to an O-ring provides an effective seal across a wide temperature range due to similar stresses experienced within the seal. In addition, compared to a conventional gasket having a width-to-height aspect ratio of 3: 1 or higher, the low width to height-to-height ratio of the longitudinal seal causes a reduction in stress, which results in a longer seal life. In other embodiments, the further seal includes an annular rectangular seal and the longitudinal seal and the annular seal each have a width-to-height aspect ratio between 1: 1 and 2: 1. Although O-rings are readily available and effectively sealed and, therefore, are often used as ring seals, a rectangular cross-section seal is used as a further seal in some embodiments. If a rectangular seal is used, the cross-sections of the two seals can be very closely matched. In some embodiments, the longitudinal seal has a width-to-height aspect ratio between 1.1: 1 and 1.3: 1. One aspect ratio of the longitudinal seal slightly larger than 1 makes it easy to manipulate and place in the groove. However, in the case where the further seal is an O-ring, making the aspect ratio close to 1 provides a closer match, thereby providing more uniform stress across the junction. In addition, although having an aspect ratio slightly larger than 1 makes it easy to manipulate, position and maintain in the groove, making the aspect ratio close to 1 provides a thicker seal, thereby providing a more uniform stress, especially towards the end of the seal. Profile, where the stress profile is important. A lower stress also reduces the chemical susceptibility of the seal as the stress increases. In some embodiments, the further seal also has a width-to-height aspect ratio between 1.1: 1 and 1.3: 1. In some embodiments, the further seal and the longitudinal seal have a cross-sectional area of a similar order of magnitude less than 50%, preferably less than 30%, and in some embodiments less than 10% from each other. The maximum possible area of the interface between the seals is limited by the cross-sectional area of the seal with the smallest cross section. Therefore, providing a similar cross-sectional area prevents the interface area from being excessively limited by a particularly low cross-sectional area. In addition, in the case where there is a good stress match in two different seals, this provides a pressure balance within each seal, thereby allowing the interface to maintain its shape, and in the case of a good match, a flat sealing surface. In some embodiments, the longitudinal seal, when uncompressed, includes a flat end surface for abutting the further seal. Although the end surface of the longitudinal seal may have a contoured shape for mating with, for example, an O-ring, in some embodiments it includes a flat surface. A flat surface is easier to manufacture and easier to manipulate. In this regard, when the pump is assembled, the end of the longitudinal seal may extend farther than the end of the stator half shell and pushed back using a tool to align with the end surface of the stator half shell. If the end surface is flat, this procedure and tool manufacturing become simpler. In addition, if the O-ring is compressed, this will provide a relatively flat surface on which the flat surface of the longitudinal seal can effectively fit and seal. In some embodiments, the longitudinal seal is manufactured to have a height above 2 mm with a tolerance of 0.07 mm and the groove is manufactured to have a height of 20% smaller than the height of the gasket with a tolerance of 0.05 mm A depth due to the compression is less than 7% due to the compression. In some embodiments, the longitudinal seal is manufactured to have a height above 2.5 mm with a tolerance of 0.07 mm and the groove is manufactured to have a height of 20% smaller than the height of the gasket with a tolerance of 0.05 mm A depth attributable to the compression change due to these tolerances is less than 5.5%. As described above, if the minimum height of the seal is set to 2 mm and the tolerance is 0.07 mm and the groove is manufactured to have a depth of 20% smaller than the height of the gasket with a tolerance of 0.05 mm, then Compression changes due to these tolerances are limited to less than 7%. However, if the longitudinal seal is manufactured to have a height above 2.5 mm with one of the similar tolerances, the compression variation due to these tolerances is reduced to less than 5.5%. In addition, if the height is further increased to 3 mm with the same tolerance, the compression variation due to these tolerances is reduced to less than 4.7%. In this way, I can understand that increasing the height of the seals (where the tolerances remain the same) reduces the compression variation due to these tolerances and this results in increased predictability and matching The ability to sense stress and strain at other interfaces is better. For the aforementioned seals, a width of preferably 3 mm between 2.5 mm and 3.5 mm provides the desired aspect ratio. When selecting the relative width of the grooves compared to the other widths of the seals, the relative width should be selected to be wide enough so that the seals do not run out of space or overfill the grooves when compressed or expanded due to increased temperature. However, the relative width should not be much wider than the requirements specified above, otherwise the seal may deviate or become out of alignment. Vacuum pumps have particularly high pressure differentials and require particularly effective sealing. Therefore, the embodiments of the present invention are particularly applicable to such vacuum pumps. Further specific and preferred aspects are stated in the accompanying independent technical solutions and dependent technical solutions. The characteristics of the dependent technical solutions may be combined with the features of the independent technical solutions, as appropriate, and in combinations other than those explicitly stated in the technical solutions. Where a device feature is described as being operable to provide a function, it will be understood that this includes providing a device feature that is either adapted or structurally designed to provide the function.

Before discussing any more details of the embodiments, an overview will first be provided. Embodiments propose longitudinal seals or gaskets that are suitable for use across a wide temperature range and have a low width-to-height aspect ratio to provide low seal deformation and low seal stress. This low aspect ratio is provided by providing a longitudinal seal with an increased height or thickness compared to conventional stator case gaskets, which not only reduces the stress in the gasket but also requires a deeper groove, and This leads to easier placement of the seal in the groove. Reduced compression variations make it easier to predict the surface pressure on the end surface that abuts or mates with the end piece seal, allowing for better stress matching and a sealing interface geometry that changes less with changes in operating conditions, such as temperature variations shape. The reduction in internal stress due to the decrease in width-to-height aspect ratio also reduces chemical susceptibility and increases the life of the seal, because permanent deformation (ie, irreversible deformation) occurs more rapidly with increasing stress. By providing similar aspect ratios to the two seals, a certain degree of stress matching is achieved for these seals, and the variation in interface geometry with temperature changes is reduced. Where the cross-sectional areas of the two seals are similar, this also provides improved stress matching and reduced interface geometry variations, resulting in improved sealing effectiveness over a wider temperature range. In addition, a similar size cross section of one of the two seals provides an increased interface area because the maximum size of this area is limited by the cross section of a seal with a smaller cross section. Compared to many conventional seals, a seal having a low width-to-height aspect ratio results in a longitudinal seal height increase. An increase in the height of a seal causes an increase in the depth of a corresponding groove, and this not only makes it easier to place and hold the seal in the groove, but its decrease is due to variations in manufacturing tolerances. For example, if the tolerance is 0.05 mm, for a 1 mm height seal, this is related to a 5% change, and for a 3 mm height seal, the change is 1/3 of the 5% change. A smaller percentage change in the size of the seal results in a smaller change in the compression range and makes it easier to predict the surface pressure applied by the end surface of the seal. Figure 1 shows diagrammatically a section through a gasket according to the prior art with an uncompressed height of 1 mm and a width of 3 mm. As can be seen, if these longitudinal washers are compressed when installed between two stator shells of a pump assembly, they will expand within the width of the groove and undergo relatively high deformation and stress. In particular, as can be seen, the edge portion of the seal has a relatively sharp cross section. The end portion of the seal adjoining the annular seal will deform in a similar manner and this sharp section can reach the corresponding annular end seal, especially if the end seal is not subjected to similar compression and stress, and this results in The interface is deformed, which in turn can cause seals to leak. Figure 2 shows a cross section through a gasket for a pump assembly according to an embodiment of the invention. In this case, at rest, the width of the washer is 3 mm and the height is 2.5 mm. Compared to the washer of Fig. 1, this shape provides a much lower width-to-height aspect ratio, and this results in considerably lower deformation and stress when the washer is compressed. In addition, the stress is more uniform across the gasket, resulting in flatter and less sharp edges and end surfaces. Figure 3 provides a table showing the effect of manufacturing tolerances on changes in compression ratio as the thickness of the gasket increases. In particular, one of the prior art gaskets with a width of 3 mm and a thickness of 1 mm in one of the gasket grooves with a depth of 0.8 mm was shown to have a compression ratio variation that was attributed to 0.07 of the gasket thickness when the stator was assembled 0.05 mm manufacturing tolerance and groove depth. The mm manufacturing tolerance is compressed as a percentage between 7.5% and 34.4%. Then consider the effect of tolerance on the compression ratio of the gasket according to various embodiments. These washers have the same width of 3 mm but have an increased height or thickness. The manufacturing tolerances of washers and grooves should be considered to be the same as the prior art, that is, the manufacturing tolerance of the thickness of the gasket is 0.07 mm and the manufacturing tolerance of the groove depth is 0.05 mm. First, consider a washer that is 2 mm thick and in a 1.60 mm groove. In this example, the tolerance causes a change in the compression ratio between 6.9 and 6.5 to be less than 7%. Increasing the thickness of the washer to 2.5 mm reduces the compression ratio variation to 5.5% or less. Increasing the thickness further to 3 mm, which provides a washer with a square cross section, reduces the compression ratio variation to 4.6% or less. The table of Fig. 3 clearly shows that the variation of the compression ratio caused by the manufacturing tolerance decreases significantly with the increase in the thickness of the gasket. This reduction in variation results in a more predictable stress in the gasket, making it possible to manufacture the pump more consistently, with better stress matching between the seals. In summary, a low thickness of 1 mm results in a wide variation in one of the possible compression ranges, and the high compression ratio at the upper end of this range accelerates the onset of compression deformation in a gasket, which results in early seal failure. High gasket compression also results in gaskets being cut into O-rings at the interface. In a preferred embodiment, the gasket is 3 mm wide and 2.5 mm high and has a cross-sectional area that is substantially the same as the annular seal. A low aspect ratio of 1.2: 1 results in low seal distortion and low seal stress. The 2.5 mm thickness results in a narrow compression range and a lower compression ratio at the upper end of this range slows the beginning of compression deformation in the gasket, thereby extending seal life. Finite element analysis has been used to optimize the balance between compression on the O-ring and compression on the gasket so that the extremes of compression and temperature do not cause an interface pressure loss or O-ring destruction. In addition, a 2.5 mm thick washer engages more actively in a housing than a 1 mm thick washer and is less likely to disengage from the groove. This helps the washer squeeze as it is compressed by the compression screw, which has been occasionally observed in a 1 mm thick washer. 4a and 4b schematically show how the shape of the interface between the end pieces in the longitudinal gasket 20 between the annular seal 40 and the stator half shell changes as the thickness of the gasket increases. Figure 4a shows the interface between a gasket 20 of increased thickness and an O-ring 40 under compression according to one embodiment. As can be seen, the gasket 20 and the O-ring apply exactly equal pressure to each other, resulting in a substantially flat surface across most of the area of the interface. The substantially equal pressure from each side is due to the substantially equal stresses experienced by each seal. In addition, this match is maintained across a wide temperature range. Figure 4b shows a thinner gasket and in this case the stress and strain experienced by the gasket is significantly higher than the stress and strain experienced by the O-ring, and therefore the end of the gasket is pushed into the O-ring, causing the interface to deform and seal the The possibility of leakage between them increases. In summary, for a seal to work well, the contact pressure between the gasket and the O-ring should be maintained across the operating temperature. The interface between the gasket and the O-ring will be shaped according to the relative compression on the gasket and the O-ring. In the case of compression on the balance washer and the O-ring, the interface is a substantially straight line (as shown in Figure 4a). Assuming there is sufficient compression on the gasket and O-ring, there will be no problems with sealing under the equilibrium conditions shown here. However, if the O-ring compression and the gasket compression do not match, one will expand into the other. If, for example, the gasket compression is greater than the O-ring compression (possibly due to its shape), the gasket will expand into the O-ring, as shown in Figure 4b. 5 to 7 show specific examples of the longitudinal seal and the annular seal and a vacuum pump assembly using the same. These seals benefit from the low width to height aspect ratio of embodiments of the present invention. The seal has additional features for centering the longitudinal seal to improve the interface with the annular seal and for providing the axial flexibility of the longitudinal seal. Some embodiments also include features for offsetting the longitudinal seal against the inner surface of the groove closest to the pump chamber to prevent fluid from leaking along the groove. Fig. 5 schematically shows a multi-chamber rotary vacuum pump assembly with two stator half-shells and end pieces, which assembly can advantageously be sealed using a seal according to an embodiment. The pump assembly is formed by two stator half-shells 104 and 102 with a rotor (not shown) installed therebetween. The two shells are fixed together to form a pump chamber. Each of the chambers 106, 108, 110, 112, 114, and 116 is separated by a pump chamber wall 134. The end pieces 122 and 124 are installed on the half shell of the stator to complete the pump assembly. FIG. 6 shows an isometric view of one of the seals arranged between the stator half shells and between the end faces. In this embodiment, there are longitudinal seals 20 and 22 on either side of the pump chamber 30. O-ring seals 40 and 42 also exist between the end pieces and the stator half-shell. In this embodiment, the ring seals 40, 42 are standard O-rings and are installed in a rectangular groove of a constant cross section that can be machined using a common cylindrical tool. It may be difficult to maintain an effective seal between one of the longitudinal seals 20, 22 and the O-rings 40, 42. In particular, the longitudinal seals 20, 22 are installed in a groove wider than one of the seals to provide freedom for certain lateral movements and the ability of the gasket to expand under compression and at an elevated temperature. However, this provides a degree of freedom for the end surface of the longitudinal seal, which means that the position of the end surface has not been accurately determined, and the end surface cannot accurately fit with the O-ring seal. FIG. 7 shows the feature of providing the centering of the longitudinal seal in the groove to solve the above problems. In particular, the washer 20 includes a one-end loop 25 that is within the groove 50 and has a curved longitudinal arm forming a bow shape. Due to the symmetrical nature of the loop, the bow shape reacts on the outer surface of the groove on either side of the loop, and each side generates a force of substantially the same value but in the opposite direction, thereby centering the seal, and In particular, the end portion 29 of the seal extending from the loop and mating with the O-ring is centered. The bow shape of the loop retains axial rigidity but allows lateral flexibility to ensure that an interference fit is possible. The cross section / width of the washer is maintained in the bow-shaped centering feature to avoid excessive expansion at high temperatures. Space is provided in the arch side and in the straight member to provide freedom of extension during compression and expansion at higher temperatures. In addition to providing this centering feature, the loop also provides some axial stability and axial flexibility. Axial stability is provided by an axial alignment surface 52 of a groove forming an outer surface of the arm extending from the longitudinal groove in either direction. The loop portion of the seal contacts the axial alignment surface and this holds the longitudinal seal axially in place. The length of the grooves between the axial alignment faces at the ends of the stator is structurally designed so that the longitudinal seal is installed under a slight tension and is therefore held more firmly in the grooves. Axial flexibility is provided by the deflection of the transverse arm to allow a certain axial movement of the seal thereby inhibiting it from becoming overtightened, where this may trigger the thinning of the corresponding seal. In the embodiment of FIG. 7, the loop 25 has a small protrusion 23 on the outer surface adjacent to the axial alignment surface 52 of the stator half-shell. These small protrusions 23 provide a known position on the loop to contact the axial alignment surface 52 of the washer against axial tension. The small protrusions are positioned toward the outside of one of the loops to allow an increase in axial deflection due to bending of the loop portion between the small protrusions 23. In this embodiment, further axial deflection is provided by the longitudinal divergences or bumps 27, which also provides the offset of the gasket against the inner surface of the groove. This bump can be contracted and expanded to allow axial deflection. Each end of the washer includes an end portion 29 extending from the loop 25. This end portion 29 is aligned with the end of the stator half-shell and contacts the O-ring. In this embodiment, the end portion 29 is centered in the shape of a bow of the loop. When the pump assembly is assembled, the lower stator has two washers mounted on either side of the pump in the groove 50. The two washers are installed at either end of the stator against the axial alignment surface 52 under tension. Then, the protruding end 29 is abutted against the end of the stator half shell by a flat end tool, and the upper stator half shell is lowered onto the lower stator half shell and the upper stator half shell and the lower stator half shell are fixed together so that the washer Hold in place under compression. Then, the end pieces 122 and 124 containing the O-ring seals 40 and 42 can be installed against the end face of the stator. As can be seen, the groove 50 has a constant width and can therefore be processed at one time using one tool. This provides an advantage over grooves containing small protrusions to maintain the seal in place. In addition, the absence of such small protrusions reduces the chance that the gasket will have a pinch point when the gasket expands under compression and temperature changes. During assembly, the protrusion 29 is pushed back to be flush with the end face of the stator. This is made possible by the flexibility of the loop 25 and, in particular, by the transverse arms of the loop, which can flex inward and outward and provide this axial flexibility. The groove 50 diverges towards the pump chamber as it approaches the axial alignment surface 52. This divergence is provided so that the offset of the gasket toward the pump chamber (s) does not offset the gasket against the inside of the groove close to its end. The centering of the washer within the groove allows the washer 20 to move laterally toward the loop 25 in response to a lateral force. This helps to center the tip 29 when opposite lateral forces are applied from either side. In summary, the washer design described above uses a simplified geometry. The end zone has a "single box" shape that provides one of the desired functions. According to the invention, a flat end washer having a width-to-height aspect ratio as stated above is formed to interface to one end seal, in this case a standard circular cross-section O-ring. Axial alignment of washers and O-ring grooves is important for a good quality seal. This centering is achieved by ensuring that when the gasket is placed in the housing, they protrude beyond the O-ring groove and then by using a custom tool to push the gasket back to the O-ring groove. An axial flexibility is required in the end of the washer to provide a protruding end that can be pushed back. This flexibility is provided by a lateral or lateral member of the support end sealing surface, see FIG. 7. The axial flexibility in the center of the washer helps ensure that it is under tension and does not bend. The washer is stretched and is positioned on the alignment protrusion 23 against the alignment surface 52 in some embodiments. Flexibility is provided by a laterally flexible member that supports the small protrusions. The bump 27 in the central region of the washer provides additional axial flexibility. Although the illustrative embodiments of the present invention have been disclosed in detail herein with reference to the accompanying drawings, it should be understood that the present invention is not limited to the precise embodiments and those skilled in the art may Equivalents define various changes and modifications within the scope of the present invention.

20‧‧‧longitudinal gasket / longitudinal seal

22‧‧‧longitudinal seal

23‧‧‧ small protrusion

25‧‧‧ end loop

27‧‧‧ Longitudinal divergence or bump

29‧‧‧ tip part / protruding end / tip

30‧‧‧pump room

40‧‧‧Ring seal / O-ring / O-ring seal

42‧‧‧O-ring seal / ring seal / O-ring

50‧‧‧ groove

52‧‧‧Axial alignment surface / Axial alignment surface

102‧‧‧Stator half shell

104‧‧‧Stator half shell

106‧‧‧pump room

108‧‧‧pump room

110‧‧‧pump room

112‧‧‧pump room

114‧‧‧pump room

116‧‧‧pump room

122‧‧‧End pieces

124‧‧‧End pieces

134‧‧‧Pump chamber wall

An embodiment of the present invention will now be further described with reference to the accompanying drawings, in which: FIG. 1 illustrates a cross section through a prior art gasket having a width of 3 mm and a height of 1 mm in a free state and a compressed state; Fig. 2 shows a cross section through a gasket in a free state and a compressed state according to an embodiment; Fig. 3 shows the influence of tolerance on a gasket of the prior art and a gasket according to an embodiment; Figs. 4a and 4b show a gasket and A cross-sectional view of an interface between O-rings of longitudinal seals of different aspect ratios; Figure 5 shows a pump assembly with a horizontal dividing line; Figure 6 shows a pump assembly with a seal according to an embodiment An isometric view; and Figure 7 shows the outline of a longitudinal seal according to an embodiment.

Claims (11)

A vacuum pump assembly includes: two half-shell stators that define one or more vacuum pump chambers; end pieces that are installed at either end of the two half-shell stators; either side of the vacuum pump chamber Longitudinal seals for sealing between the longitudinal contact surfaces of the two half-shell stators; and at least one further seal for sealing between one of the end pieces and the stator half-shells A seal; wherein the longitudinal seals have end portions adjacent to the at least one annular seal; and a width-to-height aspect ratio of the longitudinal seals and the further seals is between 1: 1 and 2: 1 . The vacuum pump assembly of claim 1, wherein the further seal is an annular seal and includes an O-ring having an aspect ratio of 1: 1, and the longitudinal seal has a rectangular cross section. The vacuum pump assembly of claim 1, wherein the further seal comprises an annular rectangular seal, and the longitudinal seal and the annular seal each have a ratio between 1: 1 and 2: 1, preferably each Width to height aspect ratio between one of 1.1: 1 and 1.3: 1. The vacuum pump assembly of any preceding claim, wherein the longitudinal seal has a width-to-height aspect ratio between 1.1: 1 and 1.3: 1. As in the vacuum pump assembly of claim 1, wherein the further seal and the longitudinal seal have a cross-sectional area of a similar order of magnitude less than 50% from each other. As in the vacuum pump assembly of claim 5, wherein the further seal and the longitudinal seal have a cross-sectional area that is less than 30% from each other. The vacuum pump assembly of claim 1, wherein the longitudinal seal, when uncompressed, includes a flat end surface for abutting the further seal. If the vacuum pump assembly of claim 1, wherein the longitudinal seal is manufactured to have a height of 2 mm or more with a tolerance of 0.07 mm, and the groove is manufactured to have a ratio of the gasket with a tolerance of 0.05 mm The height is 20% less than the depth, and the compression variation due to these tolerances is less than 7%. For example, the vacuum pump assembly of claim 1, wherein the longitudinal seal is manufactured to have a height of more than 2.5 mm with a tolerance of 0.07 mm, and the groove is manufactured to have a ratio of the gasket with a tolerance of 0.05 mm. The height is 20% less than the depth, and the compression variation due to these tolerances is less than 5.5%. The vacuum pump assembly of claim 1, wherein the longitudinal seal is manufactured to have a height of 3 mm or more with a tolerance of 0.07 mm, and the groove is manufactured to have a ratio of the gasket with a tolerance of 0.05 mm. The height is 20% less than the depth, and the compression variation due to these tolerances is less than 4.7%. The vacuum pump assembly according to any one of claims 8 to 10, wherein the longitudinal seal includes a width between 2.5 mm and 3.5 mm, preferably one of 3 mm.