Seal Face Design of Dry Running Seal Assembly to Reduce Gas Leakage

Abstract

Mechanical seal designs have always been a challenge, as there are various working conditions related to the seal. Hydrostatic load, mechanical load, surface roughness, seal film thickness, physical and chemical nature of the working fluid have affected the seal design. The paper examines the development of non-contacting dry gas seal face with rotating groove technology that has a longer life span than mechanical seals. The performance of the seal is predicted by comparing the design from the data used to determine the forces that affect gas seals with the experimental data, and the results are discussed.

1 Introduction

Fluid hydrodynamic pressure operates on a non-contact and steady seal faces. The secondary seal, stator, rotor, rotor seat and elastic component make up the gas seal. For gas seal face, design calculations different seal parameters are selected. The force analysis is studied, and sealing performance is evaluated based on set conditions. For a dry gas face seal, filtering, dehumidifying, and saving the heat flux are required to ensure the long-term continuous functioning of auxiliary system (pumps and compressors). It is also necessary to consider the design of an auxiliary system.

2 Process of Seal Design

As a non-contacting seal, the gas face seal should have a long service life as no wear occurs on its face. Weightage in balancing the force in gas seal design can yield more accurate results than the design of mechanical seals. The working principle, design parameters and the design process of the contact mechanical seal differ from the dry gas seal. To make designs successful, the balance ratios or area ratios are considered. Seal ring balance radius is directly linked to force balance in the design [1].

2.1 Design Conditions

As per the requirement from the customer, the design conditions are listed below:

Material	SiC/C
Sealing medium	Air
Mounting diameter	Ø 114.88 mm
Maximum outside diameter	Ø 174 mm
Seal pressure	0.2281 MPa
Ambient pressure	0.1013 MPa
Spin speed	10,380 RPM
Operating temperature	370 ºK
Dynamics viscosity	0.00002161 kg/ms

2.2 Seal Face Size

Figure 1 shows the gas seal face and forces acting on different parts of the assembly.

Fig. 1

Gas face seal structure [1, 2]

F_o = opening force; F_c = closing force; F_p = sealing medium force; F_sp = spring force; F_fric = friction force; d_b = balancing diameter; D_o = shaft diameter; d_i = inner diameter of the rotor; d_o = outer diameter of the rotor [3].

2.2.1 Seal Clearance

The actual nominal clearance is 0.3–2 μm based on surface roughness and waviness of seal faces. It is necessary for the film thickness to be 2–5 μm in order to prevent any contact wear between seal faces. Leakage rates increase with film thickness. The selection of sealing clearances of 3 μm is set [4].

2.2.2 Seal Face Width

Face width as a macroscopic parameter of a gas seal is difficult to guarantee the flatness and perpendicularity of a large face. Installation errors or central errors can also result in non-parallel faces. A sealing with an excessively large radial dimension would easily deviate from the end face. A narrow seal face will not produce the desired hydrodynamic effect. As the hydrodynamic grooves are processed on the seal face, its face width should be wider than the contact mechanical seal under the same conditions [4].

2.2.3 Seal Face and Shaft Sleeve Clearance

Figure 2 depicts the structural dimensions of the seal friction pair design. Based on the shaft diameter (D_o) and the width, inner diameter and outer diameter of the end face are determined. The shaft sleeve and rotating ring do not move relative to each other, but the clearance between them is e₁ = 0.5−1 mm to ensure floatability. The clearance also compensates for the effect of shaft vibrations, deflections on the shaft sleeve and the stator. The stator ring and shaft sleeve move relative to one another. To reduce the error caused by radial vibration, the stator ring and shaft sleeve should have a gap of e₂ = 1–2 mm. The stator ring and the sealing chamber do not move relative to each other, and the general gap value is e₃ = 0.5−1.5 mm. Smaller gap value should be selected for O-ring to avoid failure on higher seal pressure. The gap value corresponds to an increase in the extrusion failure of the O-ring as the hardness of the chosen O-ring increases. If the selected O-ring has a large diameter, then the gap value corresponding to its extrusion failure will be greater [4].

Fig. 2

Main structural dimensions of the seal friction pair [1]

Sections 2.2.1 and 2.2.2 illustrate selection of face width b = 19.36 mm and seal clearance ℎ_o = 3 μm. As a result, the inner and outer diameters must match the installed inner diameter of 114.88 mm and 112.8 mm, respectively, in the design condition.

A seal’s internal diameter is determined by

$$ d_{i} = D_{0} + e_{1} + 1{\text{ mm}} = 112.88 + 0.96 + 1= { }116.84\,{\text{mm,}} $$

Outer diameter of seal face

$$ d_{o} { } = { }116.84{ } + { }\left( {2{ } \times 19.36} \right){\text{ mm }} = { }155.56\,{\text{mm}} $$

2.2.4 Face Groove

The direction of the gas flow is affected by the fluid hydrodynamic groove that is formed on the rotor ring or stator ring during seal operation. When the seal is operating, these grooves may produce fluid hydrodynamic pressure. The spiral groove chosen for the face groove is shown in Fig. 3. Table 1 shows the calculated and analyzed values of the spiral groove geometric parameters based on the given design conditions [4].

Fig. 3

a Schematic dry gas seal stator and b stator [6]

Table 1 Dry gas seal geometry

• a. Opening force

  As a result of the distribution of pressures resulting from the radial pressure flow and the circumferential velocity flow of the gas film, the seal face at contact experiences an opening force as follows [5, 6].

  $${F}_{\text{o}}= {\int }_{0}^{2\pi }{\int }_{0.5{d}_{i}}^{0.5{d}_{o}}pr\mathrm{d}r\mathrm{d}\theta ={\int }_{0}^{2\pi }{\int }_{0.5\times 58.42}^{0.5\times 77.78}0.2281 r\mathrm{d}r\mathrm{d}\theta $$

  (1)
  $$ F_{{\text{o}}} = \, 1889.55{\text{ N}} $$

• b. Closing force

  A sealing friction pair’s closure force comes from the friction between its faces. The sealing fluid’s pressure F_p is exerted on the rotor ring’s sealing face from the back (balance diameter of upper surface force and lower surface force). Spring force Fsp acts on the backside of the rotor ring. Friction force Ffric of the O-ring and inertial force of the friction pair during motion make up the closing force acting on the seal faces. O-ring friction and the inertial force produced by the friction pair during motion approach equilibrium in the stable condition and are not taken into account. Closed force is equal to spring force and medium pressure when designing [5, 7].

  $$ F_{{\text{c}}} = F_{{{\text{sp}}}} + \, F_{{\text{p}}} = F_{{{\text{sp}}}} + \, F_{{{\text{po}}}} + \, F_{{{\text{pi}}}} $$

  (2)

where the pressure F_p of the sealing medium is separated into two parts (equilibrium diameter: the closing force F_po, caused by the action of the sealing medium, and F_pi, caused by the action of ambient pressure) [5, 7. Their expressions are

$${F}_{{\mathrm{p}}_{\mathrm{o}}}= {p}_{o}\frac{\pi \left({d}_{o}^{2}- {d}_{b}^{2}\right)}{4} =1515.97\mathrm{ N}$$

(3)

$${F}_{{p}_{i}}= {p}_{i}\frac{\pi \left({d}_{b}^{2}- {d}_{i}^{2}\right)}{4} =164.58 \mathrm{N}$$

(4)

$${F}_{c} =1889.55 \mathrm{N}$$

The spring force’s main job is to overcome the secondary seal’s friction and inertia and make sure the face of the seal closes when no pressure is applied.

2.2.5 Balance Diameter

The balance diameter is the boundary between the medium pressure and the ambient pressure, which directly influences the closing force [1]. The sealing force and opening force in the design must be balanced to ensure non-contact operation. Choosing a 5.45 mm as O-ring diameter and based on experimental measurements, when the compression ratio is 10%, the friction generated on the unit contact length is 240.72 N/m (according to GB/T 7757.2-2006 (standard for O-Rubber Seal Ring for Mechanical Seal)).

The friction generated by the O-ring can then be estimated as

${F}_{\mathrm{fric}}$ = Unit length friction × length = 240.72 × π × (0.05842 + 0.07778) = 103 N.

Twice as much frictional force is required for the spring force ${F}_{\mathrm{sp}}$ = 2 ${F}_{\mathrm{fric}}$ = 206 N;

The balance diameter calculated according to equation is

$${d}_{b}={\left[\frac{\pi \left({p}_{o}{d}_{o}^{2}-{p}_{i}{d}_{i}^{2}\right)+4({F}_{sp}-{F}_{o})}{\pi ({p}_{o}-{p}_{i})}\right]}^{0.5}= 125.38 \,\mathrm{mm}$$

(5)

2.3 Design of Rotor Dimension

Seal rings typically use soft or hard matching methods for the friction pair. As axial flotation rings, rotors are used, and relatively soft seal materials are used. In essence, the rotor face width is equal to the sealing face width b. Figure 4 shows the seal ring’s geometric structure and main dimensions. Rotor dimensions include axial and radial dimensions [8].

Fig. 4

Schematic diagram of rotor structure

2.3.1 Design of the Radial and Axial Dimension

The radial dimension design between the rotor and the shaft is shown in Fig. 4, so the main radial dimension design of the rotor can be carried out in accordance with Table 2 [9].

Table 2 Dimensions of the rotor’s geometrical structure

2.3.2 Rotor O-ring

According to GB/T 7757.2-2006 (O-Rubber Seal Ring for Mechanical Seal), the O-ring and its installation structure are selected, as shown in Fig. 5. Diameter of O-ring section ${d}^{\mathrm{^{\prime}}}$= 5.45 ± 0.10 mm. Take groove depth t = 4.92 mm and groove width B = 4.43 mm, then the compression rate of O-ring = (5.45–4.92)/5.45 = 9.8%.

Fig. 5

Mounting dimensions of rotor O-ring

Further, the rotor structure dimensions as shown in Fig. 4 can be determined, as listed in Table 2 [9].

2.4 Stator Structure

Figures 6 and 7 show the schematic diagram of stator stricter.

Fig. 6

Schematic diagram of stator structure size [8]

Fig. 7

Mounting dimensions of stator O-ring

For this stator structure, the primary size design guidelines are as follows:

• (a) Stator axial dimension—The installation of the O-ring is the primary consideration for the static ring’s axial structure dimension. The principle axial dimension design of the stator is given in Table 3.

  Table 3 Stator size design [9]

• (b) Radial dimension design—The main radial dimension design of the stator is given in Table 3 [4, 9].

• (c) Stator O-ring—Section diameter of O-ring ${d}_{0}$ = 5.6 ± 0.10 mm. The groove depth t = 4.76 mm and the groove width B = 7.55 mm, the compression ratio of O-ring = (5.6–4.76)/5.6 = 15%.

3 Gas Film Stiffness

A variation in gas film stiffness per unit changes the opening force of the gas film, which is correlated with the stability of the gas seal. General gas face seal design can make use of a constant gas film stiffness without disturbance frequency as the parameter of groove-type selection and optimization [4, 9].

$${K}_{z}=-\frac{\Delta F}{\Delta h}$$

(6)

where Δh is the change of the gas film thickness values corresponding to the gas film force change ΔF [9]. When calculating, $\Delta h=(0.001-0.01){h}_{o}$,

$${K}_{z}=4.696\times {10}^{12} \mathrm{N}/\mathrm{m}$$

The more stable the seal operates, the greater the stiffness and the smaller the variation of face clearance (signaling the stronger anti-interference ability of the seal).

4 Leakage Rate

Sealing performance is characterized primarily by the leakage rate. Leakage rates are influenced by rotational speed, pressure, temperature, gas viscosity, and the geometry of the seal surface [7, 9]. Volume leakage rate based on isothermal gas lubrication is given by

$$Q= {\int }_{0}^{2\pi }P\left(\frac{{h}^{3}}{12\eta }\frac{\partial p}{\partial r}\right)r\mathrm{d}\theta =8.55\times {10}^{-5} {\mathrm{m}}^{3}/\mathrm{h}$$

(7)

Leakage rate of design seal was $8.55\times {10 }^{-5} {\mathrm{m}}^{3}/\mathrm{h}$.

5 Conclusion

The leakage rate of spiral-shaped bidirectional groove dry gas seal is found to be $8.55\times {10 }^{-5}$ ${\mathrm{m}}^{3}/\mathrm{h}$ at 10,380 rpm and 3.05 μm gap under given operating conditions. By further decreasing the gap, a greater influence of the groove is exerted upon hydrodynamic lift, thereby increasing the load carrying capacity. Moreover, any increase in temperature and speed increases the power loss and leakage through the contact.

The closing force should ideally be determined by the minimum seal gap, thereby reducing leakage and sustaining power losses. However, there is a practical limit in reducing the film thickness, governed by the topography of contacting surfaces. The approach is to reduce the chance of direct boundary interactions, thus reducing friction and wear.