Field of the invention The present invention relates to a method of predicting when an oil film between an axial seal lip and a counterface, being in sliding contact with each other, will reach a critical thickness value that is insufficient to adequately separate the sliding contact. The invention further relates to a condition monitoring system adapted to implement the method and to a seal life estimator that estimates seal life on the basis of calculating when the critical film thickness will be reached.

Background

Bearing seals with multiple contacting seal lips are lubricated with grease in order to reduce friction and wear, and improve sealing performance. The grease provides oil to the sealing contact and an oil film separates the sliding surfaces. The replenishment mechanisms of lubricating greases are very different from those of oils due to the semi-solid behaviour of grease. Unlike oil, grease may not freely flow to the sealing contact, limiting the amount of lubricant supply for lubrication. The supply of oil to the contact is therefore limited by the volume of grease that is actually available for bleeding oil. This volume of grease is typically referred to as the grease reservoir. The supply of oil to the sealing contact is further limited by the oil bleeding characteristics of the grease. It has been found that the oil bleed rate increases with temperature and strongly depends on the grease type. Baart et al. presented a one-dimensional physical model for predicting the oil bleed from the grease - Oil-bleeding model for lubricating grease based on viscous flow through a porous microstructure', Trib. Trans., 2010, 53, p. 340-348 - which is incorporated herein by reference.

When the grease reservoir is no longer able to supply an amount of oil that adequately separates the seal lip and the counterface, friction, wear and loss of sealing function quickly occur. In conventional systems for monitoring sealing function, one or more specific parameters are typically measured and a critical value is defined for the one or more specific parameters. For example, temperature that is indicative of friction may be measured close to the sealing contact. When the measured value equals or exceeds a predefined maximum temperature, a warning might be triggered that the seal needs to be replaced. In other known systems, the seal is fitted with a leakage sensor that detects the presence of oil an air side of the sealing lip. Again, a warning is sent when an amount is detected that exceeds a predefined threshold.

Known systems are based on detecting faults or failures after they have occurred. The risk associated with such systems is that when the fault or failure is detected, damage to the sealed component has already taken place. Consequently, there is a need for a method of predicting when loss of sealing function will occur and for a system that implements the method. Invention summary

The present invention is based on an improved understanding of the available grease reservoir for an axially-oriented sealing lip, in combination with an oil bleed model that is applicable for seals with such an axial contact lip. The inventors have found that it is grease held on a rotating part of the seal, which predominantly acts as the grease reservoir for supplying oil to the contact. Further, the inventors have succeeded in modeling the supply of oil to the contact and loss of oil from the contact, to predict the film thickness as a function of time.

In a first aspect, the present invention provides a method of determining a time, t _c , at which an oil film separating a sliding contact interface between a counterface and an axial lip of a seal will reach a critical film thickness. One of the axial lip and the counterface is provided on a rotational part of the seal and the other of the axial lip and the counterface is provided on a non-rotational part of the seal.

Further, the seal is grease lubricated and the oil film is formed by base oil, which bleeds from a grease reservoir on the rotational part of the seal and which flows towards the contact interface under the action of centrifugal force. According to the invention, the method comprises steps of

• determining an rotational speed of the rotational part of the seal, • determining a viscosity of the base oil,

• determining the time, t _c , based on the rotational speed, the viscosity of the base oil, a cross-sectional area of the grease reservoir and a diameter of the contact interface.

The critical film thickness represents a thickness value at which the oil film is no longer capable of adequately separating the axial seal lip and the counterface, such that direct contact takes place. Without an adequate oil film, the seal lip will soon wear, leading to loss of sealing function. Suitably, the critical film thickness is a thickness equal to a surface roughness of the counterface. The method may thus further comprise a step of measuring a roughness parameter, such as R _a, of the counterface. The method may be advantageously implemented in, for example, a seal condition monitoring system, so that the seal can be replaced before sealing function is lost.

Specifically, the time, t _c , is determined according to the following relationship:

t _c = [C _x A _r + C ₂ ] where

Ci is a first constant dependent on grease microstructure

C ₂ is a second constant dependent on grease microstructure

A _r is the cross-sectional area of the grease reservoir (m ² )

η is the viscosity of the base oil (Pa.s)

n is the rotational speed (rev.min ^"1 )

d _s is the diameter of the contact interface (m).

Depending on the microstructure of the grease used, The constant Ci may have a value of between 1 .10 ¹⁰ and 1 .10 ²⁰ , and the constant C ₂ may have a value of between 0 and 1 .10 ¹⁰ . Suitably, the step of determining viscosity comprises measuring a temperature of the seal under dynamic sealing conditions, whereby temperature is preferably measured near the sliding contact interface. In a second aspect, the present invention provides a condition monitoring system for a grease-lubricated seal with an axial sealing lip as described above. The system is configured for implementing the method of the invention and comprises: · means to determine the rotational speed of the seal;

• means to determine the viscosity of the base oil in the grease;

• processing means programmed to determine the time, t _c , based on the rotational speed, the viscosity of the base oil, a cross-sectional area of the grease reservoir and a diameter of the contact interface.

Suitably, the condition monitoring system is further configured to issue an alarm when the determined time, t _c , falls below a predefined minimum value.

In a further development, the grease-lubricated seal that is monitored by the system comprises a retention surface for holding a predetermined volume of grease in the grease reservoir. The retention surface is provided on the rotating part of the seal, at a location radially inward of the axial sealing lip. The retention surface is designed such that the movement of grease under the action of centrifugal force is prevented, while the movement of base oil, which bleeds from the grease reservoir, is allowed. As a result, the supply of base oil to the sealing contact can take place for a longer period of time, which extends the life of the seal. Also, the cross-sectional area of the grease reservoir can be readily estimated, given that the reservoir has a predetermined volume. In some embodiments, the rotating part of the seal is a slinger comprising a cylindrical part and a radial flange part. In one example, the retention surface is formed by providing a bend in the flange part, which creates an overhang for retaining the volume of grease in a radial direction. Suitably, the retention surface extends in an axial direction at an angle of less than 40 degrees relative to a rotation axis of the seal. The angle of the retention surface may be adapted depending on the operating speed of the rotational part (i.e. the magnitude of the centrifugal force acting on the grease volume). For example, in low-speed applications, an angle of between 20 and 40 degrees may be used. In high-speed applications, an angle of less than 20 degrees is preferable. As a result, an axial component of the centrifugal force acting on the volume of grease is insufficient to cause sideways movement of the grease, but is sufficient to allow side-flow of base oil. In some applications, to prevent movement of the grease, the retention surface may be parallel to the axis of rotation, so that the axial component of the centrifugal force is zero. In such applications, a pressure differential created within the rotating grease is sufficient to cause side flow of the base oil.

In a second example, the reservoir comprises an overhanging lip, so that the volume of grease is retained in a radial direction and in an axial direction. The reservoir may be a separate part that is moulded to or adhesively fixed to the slinger flange. To allow the movement of base oil, the retention surface further comprises channels. The channels may be grooves provided in the lip, which allow side-flow of oil out of the reservoir. The channels may also be through-holes provided in the reservoir which allow base oil to flow in a radially outward direction.

In a still further example, the retention surface is made of a porous material, whereby the pores in the reservoir act as channels for the base oil. Preferably, the channels have a width of less than 1 mm, so that grease cannot escape from the reservoir via the channels. The number of channels and the size of the channels is selected depending on the volume of base oil that is advantageously supplied to the sealing contact.

In other embodiments, the axial lip rotates and bears against a stationary counterface. The retention surface is then suitably moulded into an elastomeric element on which the axial sealing lip is provided. The retention surface may comprise an overhanging lip and further comprise channels, as described above. Alternatively, the retention surface may be formed by a roughened surface on the elastomeric element. Again, the roughened retention surface preferably extends in an axial direction at an angle of less than 40 degrees relative to the rotation axis. In high speed applications, the angle is preferably less than 20 degrees. In a third aspect, the present invention provides a seal life estimator for estimating an expected life of a grease-lubricated seal under dynamic sealing conditions, whereby the seal has an axial lip in sliding contact with a counterface. One of the axial lip and the counterface is provided on a rotational part of the seal and the other of the axial lip and the counterface is provided on a non-rotational part of the seal. Therefore, an oil film forms under dynamic sealing conditions, which oil film separates a sliding contact interface between the axial lip and the counterface. The oil film is formed by base oil, which bleeds from a grease reservoir on the rotational part of the seal and which flows towards the sliding contact interface under the action of centrifugal force. According to the invention, the estimator has program means configured to estimate seal life based on a time, t _c , at which the oil film will reach a critical thickness, where the time, t _c , is calculated based on a rotational speed of the rotational part of the seal, a viscosity of the base oil, a cross-sectional area of the grease reservoir and a diameter of the sliding contact interface.

Specifically, the seal life estimator is configured to calculate the time, t _c , accord to the following relationship: t _c = [C _x A _r + C ₂ ] where

Ci is a first constant dependent on grease microstructure

C ₂ is a second constant dependent on grease microstructure

A _r is the cross-sectional area of the grease reservoir (m ² )

η is the viscosity of the base oil (Pa.s)

n is the rotational speed (min ^"1 )

d _s is the diameter of the contact interface (m).

In one application, the seal life estimator is used as part of seal selection program for selecting an optimal seal for a particular application.

Thus, there are several applications where it is beneficial to estimate when the oil film that separates an axial sealing contact interface will reach a critical thickness as which inadequate separation occurs. Other advantages of the present invention will become apparent from the detailed description and accompanying drawings.

Brief description of the drawings

Figure 1 shows a cross-sectional view of part of a bearing seal having an axial sealing lip in contact with a counterface (a) and a grease reservoir (b).

Figs. 2a, 2b are details showing dimensions of the axial sealing lip contact of

Figure 1 , with oil loss due to seal pumping (Figure 2a) and with ingested meniscus and oil loss from centrifugal body load (Figure 2b).

Figure 3 shows dimensions of the grease reservoir on the bearing seal of

Figure 1 , after the churning phase.

Figure 4 is a flowchart of the calculation scheme for film thickness calculation. Figure 5 is a graph of oil film thickness against time, starting from fully flooded conditions.

Figure 6 is a graph of operating time until critical film thickness is reached against temperature, at different rotational speeds and for A _r = 0.5 mm ² . Points are model calculation results.

Figure 7 is a graph of operating time until critical film thickness is reached against temperature, for different grease reservoir sizes and at n =

1000 rpm. Points are numerical calculations.

Figure 8 is a graph of operating time until critical film thickness is reached based on the /(n ² -d _s ) parameter for A _r = 0.5 mm ² .

Figure 9 shows an example of a condition monitoring system according to the invention, comprising an adapted seal - part of which is shown in cross-section.

Figure 10 is a cross-sectional view of part of a further example of an adapted seal that is suitable for use in a condition monitoring system according to the invention. Detailed description

Figure 1 shows a cross-section of part of a grease-lubricated bearing seal. The seal 100 in the depicted example comprises a metal casing 105 to which an elastomeric seal body 1 10 is bonded, and further comprises a slinger 120. The elastomeric seal body comprises an axial lip 1 15, which is contact with an axial counterface on a radially extending flange part 125 of the slinger. The elastomeric seal body further has a radial lip 1 17, which contacts a radial counterface on an axially extending cylindrical part 127 of the slinger. The present invention is concerned with the axial sealing contact between the axial lip 1 15 and the slinger flange 125, which contact region is highlighted in Figure 1 by the outlined box indicated at reference numeral 130.

In grease lubricated seals, the lubrication condition, or film thickness, is assumed to be determined by the availability of lubricant near the sealing contact. Grease may form a reservoir on the seal body and/or on the counterface, depending on which part is moving. In the example of Figure 1 , the seal 100 is adapted for mounting to a bearing with a rotational inner ring and a non-rotational outer ring. Therefore, in use of the seal, the slinger 120 is the rotating part and most grease will be thrown onto the stationary seal body 1 10. Only a relatively small amount of grease will remain on the slinger 120, which grease acts as a reservoir that will then slowly release oil for lubrication of the sealing contact. The grease reservoir is denoted in Figure 1 by the outlined box indicated at reference number 140.

The oil flow from the grease reservoir 140 is driven by the large centrifugal forces that result from the rotational movement. The stationary grease on the seal will also release oil but on a much longer timescale due to the absence of these forces. The formation of the grease reservoir depends on the amount of grease that is applied and the initial grease distribution in the volume between the seal body 1 10 and the counterfaces of the slinger 120. For the model presented in the following, it is assumed that after a churning phase, the grease is distributed through the volume between the elastomeric body 1 10 and the counterfaces and that a permanent grease reservoir is formed on the slinger 120. No creep flow, dynamics, and grease degradation are assumed to be present. Initially a maximum film thickness in the axial sealing contact 130 is assumed, which is determined by the operating conditions and seal lip force. Subsequently, the film thickness decreases, depending on oil loss, due to seal pumping and centrifugal effects, while there will be replenishment from the grease reservoir.

The oil feed to and oil loss from the sealing contact has been simulated to predict the film thickness inside the sealing contact 130. The oil loss from the contact takes place due to the natural pumping action of the seal and centrifugal forces acting on the oil film in a gap between the axial lip 1 15 and the slinger flange 125, as indicated in Figure 2a and Figure 2b respectively. Figure 3 schematically presents the oil feed due to oil bleed from the grease reservoir 140 on the rotating slinger 120. A balance between the oil feed and oil loss mechanisms determines the amount of oil in the sealing contact as where ,,· _/ is the volume of oil in the contact 130 as a function of time t and Qf _eed and Qioss are the flow rates of the oil feed and oil loss respectively. The initial volume of oil in the contact is V ₀ and results from the churning phase. Assuming a uniform oil film in the contact, the oil film thickness is defined as

\ ₀ '(Q _feed - Q _loss )dt + v _o

(2)

InR b where R _c is the radial position of the contact and b is the width in the contact that contains the oil as defined in Figure 2a and Figure 2b.

Oil feed

The specific oil feed from the grease reservoir is based on the oil bleed model presented in Oil-bleeding model for lubricating grease based on viscous flow through a porous microstructure', Trib. Trans., 2010, 53, p. 340-348 by Baart et al, which is incorporated herein by reference. In this model the grease is considered as a porous medium where the grease thickener forms a solid microstructure that holds the base oil inside. Due to external body forces, the oil will flow through the microstructure and bleed out at the outer radius. This approach has been based on Darcy's law, which gives the fluid velocity in its general form as

u =— Vp (3)

η where k is the permeability, η the base oil viscosity, and vp the pressure gradient. The model referred to above, by Baart et al, shows how the permeability, or ability for the oil to flow through the thickener microstructure, is a function of the oil volume fraction in the grease and the orientation of the microstructure. The pressure gradient vp in Darcy's law can be replaced by a body force, e.g. gravity or centrifugal force. Consequently the oil flow rate or oil loss is a function of the grease microstructure, temperature, and rotational speed and in its most simple form reads

Q _bleed = p ² r , (4)

where R ₀ the outer radius of the grease reservoir, W is the width of the grease reservoir, f is the thickener volume fraction in the grease, η is the oil viscosity from the Walther equation, and the body force is given by the oil density p, the angular velocity ω, and the radius r. Characteristic dimensions of the grease reservoir are given in Figure 3. The dimensions may be estimated, based on the grease shear stress and on measurements performed on sample seals, after the churning phase. Preferably, the rotating part of the seal is designed to hold a predetermined volume of grease, which will be explained later on.

The oil bleeding, or oil supply to the sealing contact, has a maximum at the beginning and reduces as a function of time. Oil loss

The oil loss from the sealing contact results from body forces on the oil film in the contact and the natural pumping action of the seal such that (5) where Qp _ump is the oil loss from the natural pumping of the seal and Qbody is the oil loss from body forces. Oil loss from natural pumping of the seal is predicted using empirical equations derived from Horve in 'Shaft seals for dynamic applications' [Marcel Dekker Inc. 1996, textbook], which is incorporated herein by reference. These equations have been derived from extensive experimental work on oil lubricated radial shaft seals of which both sides are fully flooded with oil. It is assumed that the equations can also be applied to axial contacting seals. Since the seal geometry, i.e. lip angles, and seal material are not included in this model, it is assumed that the axial sealing contact in Figures 2a and 2b is equal to the seal geometry and material that Horve used. From the Horve equations, an equation for the seal pump rate is derived assuming fully flooded conditions

= 1.04 - 10 ⁸ ? ³ —G ^1/3 , (6)

where R _c is the radial position of the contact, n the shaft speed in rpm, and G the duty parameter given as n η

G = 2nb

60 F, (7)

where Fn _p is the specific lip force. The film thickness can also be calculated from the Horve equations. This is the film thickness at fully flooded conditions and consequently the maximum film thickness possible as: /V _X = 0.01 U? _C G ⁵ ' (8)

The natural pumping equation and maximum film thickness equation from Horve are only valid when the seal is fully flooded. Other work has shown that as soon as all the oil is pumped from the air side to the oil side of the seal, the oil meniscus will ingest into the sealing contact as indicated in Figure 2b. When this happens, pumping due to the asymmetric pressure distribution will no longer take place. This hypothesis has been proven through numerical simulation, demonstrating that even multiple equilibrium positions for the meniscus exist. Consequently the seal will not continue pumping oil. The volume of oil that is subsequently present in the sealing contact is given as

2nR _c b 'max ' (9) where b is the width of the oil volume when the meniscus is ingested and h _max is the maximum film thickness as in fully flooded conditions. This volume V _max is considered as the maximum volume of oil in the contact below which the seal no longer pumps oil according to Eq. (6). Oil loss from seal pumping is dominating when the oil volume in the contact is ,,· _/ > V _max . When ,,· _/ < V _max then seal pumping is no longer present and oil loss only takes place due to the centrifugal forces on the oil film in the sealing contact. Figure 2a and 2b graphically define these lubrication conditions and corresponding contact dimensions.

Assuming a linear velocity profile over the film thickness in circumferential direction, the centrifugal forces on the oil film vary from zero on the stationary seal surface, where z = 0, to the maximum value on the rotating flinger surface, where z = h. The specific body force then becomes where p the oil density, ω the angular shaft velocity and r the radial position. Considering a thin layer, the body force induces shear flow defined by where u _r is the flow velocity in radial direction, Because the width b is small compared to the contact radius R _c , it can be assumed that the radial flow velocity is constant through the contact width such that r in Eq. (1 1 ) can be replaced by R _c . Eq. (1 1 ) can be integrated twice and after applying the boundary conditions where u _r = 0 at the walls, the radial velocity u _r is then found as

u _r (z) - (12)

\2 ²

Integrating Eq. (12) over the gap height and multiplying with the contact circumference gives the oil loss from body forces as

a* - ² ^-* ¹ < ¹³ >

The oil loss C _y will result in a reduction of the lubricant film thickness in the seal contact keeping the width b constant.

To summarize the theoretical model, three phases of operating conditions can be identified.

• Phase 1 : In the churning phase the grease is being churned and the grease reservoirs are formed. The film thickness in the contact is equal to the maximum film thickness h _max .

• Phase 2: Oil bleed from the grease reservoir supplies oil to the sealing contact.

At the same time oil is lost from the contact due to centrifugal forces and seal pumping as long as fully flooded conditions remain as in Figure 2a. • Phase 3: The maximum film thickness cannot be maintained due to insufficient replenishment of the contact. The oil meniscus is ingested into the sealing contact and seal pumping stops as in Figure 2b. Oil loss from the contact is continued due to the body forces and the film thickness decreases slowly in time.

To predict the film thickness in time, Eq. (2) is numerically integrated according to the calculation scheme in Figure 4, where the different phases are included. Here the two conditions as described in phase 2 and phase 3, and shown in Figures 2a and 2b, can be identified.

Results

A prediction of the oil film thickness in the axial sealing contact of a specific bearing seal and grease type at different operating conditions is made using the parameters presented in Table 1 . Parameters related to the grease properties are taken from Baart et al, which apply for a lithium complex grease with mineral base oil, as used in their study. Other parameters are related to the seal design in Figure 1 and Figures 2a and 2b and the seal pumping parameters from Horve.

Table 1 Parameter values for the grease and seal design. To evaluate the influence of the operating temperature, rotational speed and grease reservoir, the model is used to calculate the time until a critical film thickness h _crit = 0.25 μιτι, equal to the counterface surface roughness (R _a ), is reached. The critical film thickness may be used as an indication for the transition from the full film lubrication to the mixed lubrication regime where some direct contact between the seal and counterface may occur.

Figure 5 shows the predicted film thickness as a function of time and operating conditions including and excluding a grease reservoir with a cross-sectional area A _r . The cross-sectional area is obtained by multiplying the width W of the grease reservoir 140 by the height H ₀ (refer Figure 3). The maximum film thickness depends on the rotational speed and temperature, i.e. the base oil viscosity. In the case that a grease reservoir is present, i.e. A- = 0.5 mm ² , for a limited amount of time, the maximum film thickness remains constant before it decreases due to the limited replenishment of the sealing contact. The curves for the two speeds at 70 °C show that at lower speeds the maximum film thickness is present for a much longer time since at high speed the oil loss from the sealing contact is much higher. Figure 5 also shows the predicted film thickness without the presence of a grease reservoir (A _r = 0). Here the film thickness starts to decrease instantaneously and it takes significantly shorter time to reach the critical film thickness. For example at 70 °C and 2000 rpm it takes with the grease reservoir about 80 h to reach the critical film thickness of 0.25 μιτι and without the grease reservoir only 0.5 h. In Figure 5 the cases that include a grease reservoir all start with the same grease reservoir size. In reality the formation of the grease reservoir during the churning phase will also depend on the operating conditions and it is likely that a smaller grease reservoir is formed at higher temperatures and higher speeds. Consequently, the difference between the two curves at a constant temperature of 70 °C will be even more pronounced. The initial grease fill and the distribution of the grease in the sealing system are therefore of crucial importance for the formation of the grease reservoir and the total amount of oil that can be available for lubrication. In the model, a temperature increase only results in a reduction of the base oil viscosity. Consequently, equal model results, i.e. equal film thickness, can be obtained when at higher temperature grease with higher viscosity base oil is used such that the effective base oil viscosity in unchanged. Figure 6 shows the time until the critical film thickness is reached for several operating conditions. Here the symbols represent the model results from Eq. (2). The results show how the critical time is a function of temperature for different rotational speeds and a fixed grease reservoir size. Figure 7 shows that this critical time can be significantly extended when a larger initial grease reservoir is present, provided that HJW = 2, at fixed rotational speed.

The lines in Figure 6 and Figure 7 represent a fit to the results of the physical model in Eq. (2). This fit can be used as a simple engineering model which is based on some characteristic parameters. In both the oil bleeding model in Eq. (4) and the oil loss model in Eq. (13) the influence of the oil viscosity, rotational speed and seal contact radius on the flow rate is given as (ω-Rcf/n provided that R ₀ ~ R _c - Here the u)-R _c is very similar to the n-d _m term, where n is the rotational speed in rpm and d _m is the mean diameter as commonly used in the modelling of rolling element bearings. Similar parameters are used to define a characteristic term /(n ² -d _s ² ) where d _s = 2 R _C is the diameter of the sealing contact. According to the film thickness equation in Eq. (2) this term has to be divided by the contact radius giving /(n ² -d _s ) and is used as de basis for the engineering model. Here also the influence of the size of the grease reservoir is included which has been presented in Figure 7. Including two fitting constants the engineering model reads where for this particular grease type C ₁ = 2-10 ¹⁵ and C ₂ = 6-10 ⁶ and A _r = WH ₀ provided that H ₀ /W = 2. The seal geometry is included in d _s , the base oil viscosity and temperature in η and the rotational speed in n. The results of Eq. (14) are plotted as continuous lines in Figure 6 and Figure 7. Figure 8 shows how the results from the physical model in Figure 6 fall on one line when scaled with the characteristic term /(n ² -d _s ).

A physical model to predict the lubricating conditions and film thickness in an axial sealing contact has been presented. Based on this model a simplified engineering model has been developed that can be used to predict the time until the mixed lubrication regime is reached. It is possible to include a more advanced pumping model including the actual seal lip geometry. However, when the pumping phase (phase 2) is short compared to the phase where the film thickness decays (phase 3) and the maximum film thickness is at least a few times higher than the critical film thickness, the pumping rate has little effect on the predicted time to mixed lubrication. Under these conditions the present engineering model is also valid for other axial seal materials and geometries. In addition, the model can be extended with e.g. a prediction of the grease reservoir formation and percolation theory for calculations below the critical film thickness, i.e. in the mixed lubrication regime.

In a further aspect, the present invention defines a condition monitoring system for a grease-lubricated seal, which is adapted to implement the engineering model expressed in Equation (14).

An example of a system according to the invention is shown in Figure 9. The system comprises a seal, which is suitable for use in a wheel bearing unit adapted for inner ring rotation . The inner ring rotates about an axis of rotation 950. The seal has an elastomeric body 910 bonded to a metal casing 905, which casing is e.g. press-fitted into a bore of the outer ring. The elastomeric body 910 has an axial sealing lip 915. The seal further comprises a slinger 920, which has a radial flange part 925 and a cylindrical part 927. The cylindrical part 927 is mounted on the bearing inner ring and the slinger 920 is thus rotational about the axis 250. In use of the bearing, a dynamic sealing contact 930 is defined between the axial sealing lip 915 and an axial counterface on the radial flange part 925. The seal is further provided with a grease reservoir 940 for supplying base oil to the sealing contact 930 under the action of centrifugal force. The reservoir 940 is provided on the rotating part of the seal, i.e. the slinger 920. In the method of the invention, the cross-sectional area of the grease reservoir 940 is one of the key parameters for calculating the time at which the oil film separating the sealing contact 940 will reach a critical (inadequate) film thickness. For the slinger 120 depicted in Figure 1 , the cross-sectional area may be estimated, based on the grease shear stress and based on measurements performed on sample seals after the churning phase. In a further development, as depicted in the example of Figure 9, the slinger may be designed to hold a predetermined volume of grease, to function as the grease reservoir 940. This not only enables a more accurate estimation of the cross-sectional area of the reservoir; the amount of grease in the reservoir can be increased. As explained above, this lengthens the time to critical film thickness and extends seal life.

In this example, the radial flange part 925 of the slinger has a bend located between the sealing contact 930 and the cylindrical slinger part 927. The flange part 925 therefore has an axially extending surface 926. This surface, which will be referred to as a retention surface, acts as an overhang for radially retaining the grease reservoir 940. Suitably, the retention surface 926 extends at an angle a of less than 40 degrees relative to the rotational axis 250. In the example of Figure 9, the angle is approximately 25 degrees. As a result, the centrifugal force acting on the grease reservoir 940 has a relatively small axial component, which is insufficient to allow sideways movement of the grease, but which allows side flow of base oil from the grease. Upon reaching the edge of the retention surface 926, the base oil will then flow in a radially outward direction towards the sealing contact 930.

Other designs are possible. For example, an overhanging part may be molded to the slinger flange. To enhance the flow of oil from the grease reservoir retained by the overhang, radially extending channels may be provided in the moulding. Equivalent features may be moulded into the elastomeric body of a seal, when it is this part of the seal that is rotational in use. An example of such a seal is shown in Figure 10. The retention surface 1016 is formed on an axially extending part of the seal body 1010. The grease reservoir 1040 is further retained by an overhanging lip 1017, which is provided with grooves 1018 to allow the flow of base oil.

In the method of the invention, it is further necessary to determine the rotational speed of the rotating part and to determine the viscosity of the base oil that bleeds from the grease. In the example, depicted in Figure 9, an axially outer side of the slinger flange is provided with a magnetized rubber moulding 960. As the slinger rotates, the changing magnetic field is detected by e.g. a Hall sensor (not shown) that transmits its signal to a processing unit 980 (shown schematically). The rotational speed of the slinger is obtained from the Hall sensor signal.

To determine viscosity, the system further comprises a temperature sensor 970, which transmits a temperature signal to the processing unit 980. Viscosity is then calculated based on, for example, a look-up table of viscosity values at various temperatures for the base oil in the grease used. The temperature sensor 970 may be a thermocouple that is embedded in the elastomeric body 910, close to the axial sealing lip 815. An optical fiber sensor or any other suitable sensor may also be used. The processing unit 980 is programmed to implement Equation (14) and calculate the time at which critical film thickness will be reached, based on the predetermined volume of the grease reservoir, the measured rotational speed and the oil viscosity determined from the measured temperature. Suitably, the processing unit is further programmed to issue a warning when the calculated time (t _c ) falls below a minimum threshold.

Equation (14) may also be used to estimate the life of a particular seal, based on the expected operating conditions, such that the optimal seal may be selected for a particular application.

The invention is not to be regarded as being limited to the examples described above, a number of additional variants and modifications being possible within the scope of the subsequent patent claims. Nomenclature

A _r Cross sectional area of grease reservoir m b Seal contact width m c, Constant - c ₂ Constant -

D _f Soap fiber diameter m ds Sealing contact diameter m

Fbody Specific body force N/m ³

Flip Specific lip force N/m fo Initial soap mass fraction -

G Duty parameter -

Ho Initial height of grease reservoir m h Film thickness m hmax Maximum film thickness m k Permeability m n Rotational speed rev.min ^"1 vp Pressure gradient Pa/m

Qbody Flow rate oil loss - body force m ³ /s

Qfeed Flow rate oil feed m ³ /s

Q/oss Flow rate oil loss m ³ /s

Qpump Flow rate oil loss - pumping m ³ /s

Rc Radius axial contact m

Ro Radius grease reservoir m r Radius m t Time s

U _r Radial velocity m/s

Vo Initial volume oil in contact m v _oil Volume oil in contact m

W Width of grease reservoir m η Oil viscosity Pa-s

Base oil viscosity at 40°C Pa-s

Base oil viscosity at 100°C Pa-s

P Grease density kg/m ³ ω Angular velocity rad/s