1、14FTM14 AGMA Technical Paper Theoretical and Experimental Study of the Frictional Losses of Radial Shaft Seals for Industrial Gearbox By M. Organisciak, P. Baart and S. Barbera, SKF ERC, A. Paykin and M. Schweig, SKF Sealing Solution 2 14FTM14 Theoretical and Experimental Study of the Frictional Los
2、ses of Radial Shaft Seals for Industrial Gearbox Michel Organisciak, Pieter Baart and Stellario Barbera, SKF ERC, Alex Paykin and Matthew Schweig, SKF Sealing Solution The statements and opinions contained herein are those of the author and should not be construed as an official action or opinion of
3、 the American Gear Manufacturers Association. Abstract The improvement of the energy efficiency of industrial gear motors and gearboxes is a common problem for many gear unit manufacturers and users. As other mechanical components, the radial lip seals used in such units generate friction and heat a
4、nd contribute to the energy losses of the mechanical systems. Today, simulation tools are today helping to improve the efficiency of mechanical systems, but accurate models for seal frictional losses need to be developed. In this paper SKF presents an engineering model for radial lip seal friction b
5、ased on a physical approach. The friction model includes the generation of friction due to rubber deformation and lubricant viscous shear between the surfaces of a seal and a shaft. The friction model is coupled with heat generation and seal thermal model. Indeed, seal friction and seal temperature
6、are closely related: the heat generated in the sealing lip is conducted through the seal and shaft and dissipated into the environment. This increases the temperature of the lubricant, changing its viscosity and the temperature of the seal, changing the material properties. The model is verified ste
7、p by step in an extensive experimental study. Measurements of seal friction, seal temperature and lubricant film thickness have been performed for various industrial seals. The analyzed parameters are: surface speed, oil viscosity, seal material, seal size, seal lip type and duty cycles. The compari
8、son between model predictions and experimental friction measurements shows a very good correlation. The model enables the use of the model to predict seal frictional losses in application and identify key parameters to design low friction sealing solutions. Copyright 2014 American Gear Manufacturers
9、 Association 1001 N. Fairfax Street, Suite 500 Alexandria, Virginia 22314 October 2014 ISBN: 978-1-61481-106-0 3 14FTM14 Theoretical and Experimental Study of the Frictional Losses of Radial Shaft Seals for Industrial Gearbox Michel Organisciak, Pieter Baart and Stellario Barbera, SKF ERC, Alex Payk
10、in and Matthew Schweig, SKF Sealing Solution Introduction Industrial gear units are widely used in power transmission systems. They are composed of shafts, gears, rolling elements bearings and dynamic lip seals. The performance of the seals is critical for the proper functioning of the system. The p
11、rimary functions of the seals are to prevent the leakage of oil to the environment and to avoid the ingress of water or other contaminants into the mechanical system. Both can lead to a premature failure of the gear unit. In addition, the seals influence the system by generating friction and heat. T
12、he heat generated by the friction of the seals has an impact on the operational temperature of the gear unit as well as on the viscosity of the lubricant inside the unit. Moreover the seals contribute to the total energy losses of the mechanical system. The improvement of the energy efficiency of in
13、dustrial gear motors and gearboxes is a common challenge for OEMs and end-users. For instance energy efficiency classes are defined for electrical motors and gear motors. Moreover the power losses of gear units and seals can impact the total energy bill of an industrial installation. Therefore under
14、standing seal friction generation and reducing it are essential challenges for seal manufacturers. Simulation tools are commonly used to design mechanical components and systems. For the prediction of specific parameters like seal temperature or friction torque, specific models and calculation tools
15、 need to be developed. In this paper SKF presents an engineering model for the prediction of radial lip seal friction based on a physical approach. The friction model includes the generation of friction due to rubber dynamic deformation and lubricant viscous shear between the surfaces of a seal and
16、a shaft. The friction model is coupled with a heat generation and seal thermal model. Indeed, seal friction and seal temperature are closely related: the heat generated in the sealing lip is conducted through the seal and shaft and dissipated into the environment. This changes for instance the lubri
17、cant viscosity. The model is verified step by step in an extensive experimental study. Measurements of seal friction, seal temperature and lubricant film thickness have been performed for various dynamic lip seals. The analyzed parameters are: surface speed, oil viscosity, seal material, seal size,
18、seal lip style and duty cycles. The correlation between model predictions and experimental friction measurements can therefore be verified. This unique modelling capability allows selecting or developing shaft seals which would meet and exceed the demands of modern gearbox applications. It also enab
19、les gearbox manufacturers to bring to the market better performing and more reliable gearboxes. Seal friction modeling Physical phenomena influencing seal friction The friction force, FT, is the force resisting the relative motion of two bodies when a normal force, FN, is applied to the contact betw
20、een these bodies. The coefficient of friction, , can be defined as: TNFF (1) The coefficient of friction is not constant for radial shaft seals, which makes the prediction of seal frictional torque much more complicated. This has been demonstrated in various studies. In 2005, Plath 1 developed a sea
21、l friction model based on finite element analysis. They assumed initially a constant coefficient of friction for the seal-shaft contact. However this led to inaccurate results and they demonstrated that it was necessary to take into account the variation of temperature of the seal due to the generat
22、ed frictional heat to accurately predict seal friction. More recently, the studies from Haas 2, 3 have revealed the influence of surface roughness and of the duty parameter G (representing the lubricant viscosity, angular speed and contact pressure) on the 4 14FTM14 friction coefficient. Their paper
23、s show that the friction coefficient follows a Stribeck-like curve (see Figure 1). A transition between mixed and fully lubricated regime is clearly shown in the evolution of the friction coefficient. The variations of coefficient of friction in a radial lip seal contact can be attributed to three p
24、henomena: - The variation of lubricant viscosity as a function of temperature. Typical curves for standard gearbox oils are shown in Figure 2. - The variation of the coefficient of friction between rubber and steel. As shown by Grosch 4 and Hermann 5, the coefficient of friction varies significantly
25、, between 0.1 and 3 in extreme cases, as a function of temperature, sliding speed and pressure in dry and lubricated conditions. This is due to the fact that for rubbery material, friction is essentially governed by the dissipation of energy during the dynamic deformation of the rubbery material on
26、the counter-face. - The variation of rubber modulus with temperature. A typical curve is shown in Figure 3. Therefore the prediction of seal friction is a complex task and requires a model being able to predict the temperature in the seal and in the contact and to take into account the variations me
27、ntioned in the previous paragraph. Friction model The friction between the seal and shaft is considered to be generated by two main governing phenomena: - Lubricant viscous shearing. This takes place in the contact between the lip and the shaft surface. The frictional force produced in this manner i
28、s defined as Flub. - Viscoelastic losses. This is due to dissipation in the rubber as its surface is dynamically deformed by the shaft roughness asperities. The frictional force generated by the rubber material is referred as Fmaterial. Figure 1. Stribeck curve: coefficient of friction as a function
29、 of contact speed and lubricant viscosity Figure 2. Lubricant viscosity as a function of temperature for VG 32, VG68 and VG220 oils 5 14FTM14 Figure 3. Modulus as a function of temperature of a typical NBR material Taking both these effects into account, the total frictional torque TTorquecan be exp
30、ressed as: 2shaftTorque Lub materialDTFF (2) where TTorqueis seal frictional torque, Nm; Flubis seal lip force, N; Fmaterialis contribution of the material to the seal frictional force, N; Dshaftis shaft diameter, m. The material contribution is calculated following the relation below: material dry
31、tip cF FfA (3) where dryis the coefficient of friction between the rubber and steel surface; Ftipis seal lip force, N; f is a function of given variables; Ac is real contact area at the surface roughness level, which is calculated from contact mechanics, m2. The lubricant contribution can be written
32、 as: lub contacteuFSh (4) where is lubricant viscosity in the contact, Pa s; u is surface speed, m/s; he is effective film thickness depending on the lip tip style (i.e., wave or plain), m; Scontactis surface area where the lubricant is sheared, m2. The effective film thickness is based on elastohyd
33、rodynamic lubrication theory 6 and can be written as: 0.66ehfu (5) The combination of these equations allows the calculation of the seal friction torque at any given speed and temperature. 6 14FTM14 Thermal dissipation model The friction between a rotating shaft and a seal lip generates heat that is
34、 dissipated by the different components of the system. The power dissipated qdispby the sliding contact can be written as: 2disp TorqueshaftqT uD (6) where qdispis power dissipated in the sealing contact, W. The generated heat flux in the seal - shaft contact is integrated into the heat conservation
35、 equation for the lip contact. The heat is then diffused in the shaft and seal according to the energy equation: ()222PdispC TTTTfqkt X Y Z (7) where is material density, kg/m3; CPis heat capacity, J/K; k is heat conductivity, W/(mK); T is temperature, K; t is time, s; The complete computational alg
36、orithm is indicated in Figure 4. Here, the friction model is combined with the thermal model. The effects of temperature change on lip force and oil viscosity are also included. The algorithm is transient, allowing computations for different speed cycles. Experimental techniques used for model valid
37、ation The validation of the model is conducted for three parameters: - the lubricant film thickness in the contact (to validate equation 5); - the frictional torque; - seal temperature. Figure 4. Calculation algorithm 7 14FTM14 Film thickness measurements The measurement of an absolute value of film
38、 thickness in the sealing contact has always been a challenge. For instance in 1992, Poll and Gabelli 7 developed a method where they use magnetic fluid as a lubricant and measure the magnetic resistance through the lubricant film in the sealing contact. In the same period, Poll 8 used the fluoresce
39、nt technique: a fluorescent dye is added to the oil and is excited with a laser. The intensity of the light can be related to the film thickness in the contact. However, both techniques require complex calibration and specific equipment. In this work, a capacitance technique using the SKF Lubcheck s
40、et-up is applied to measure the evolutions of lubricant film thickness in a radial lip seal - shaft contact. Seals molded from a special conductive rubber compound have to be used to realize the experiments. This compound is part of the SKF compound portfolio and has similar mechanical proprieties a
41、s standard sealing materials. Figure 5 shows the electric schematic of the measurement system. Vmxis the maximum voltage applied to the system; Crefis a reference capacitance added to the system; Cmis the capacitance of the sealing contact; and Rmis the electrical resistance of the seal itself. Afte
42、r calibration using lubricants with different viscosities and simultaneous friction torque measurements, the measured voltage Vcapcan be related to the capacitance of the sealing contact and therefore to the lubricant film thickness. The system is implemented on the test rig shown in Figure 6. Frict
43、ion torque and seal temperature measurement Seal friction measurements are performed on a specialized SKF test rig shown in Figure 6. The shaft is driven by an electrical motor allowing a very wide, programmable, range of rotational speed. The central part of the test rig is the air bearing spindle,
44、 onto which the stationary seal specimen is mounted and the friction torque sensing unit is connected. The air bearing ensures that the measured friction is only due to the seal. The seal is lubricated with an oil bath and different oil sump volumes are possible. Figure 5. Electrical schematic for L
45、ubcheck measurement Figure 6. Seal friction measurement test rig 8 14FTM14 In addition to the frictional torque, seal temperature is constantly recorded during the tests. Thermal measurements are made using a thermocouple placed in the spring groove of the seal. The analysis of temperature changes i
46、s used in combination with frictional torque to validate the model. Model validation: correlation between the model and experimental results Film thickness Using the set-up described earlier, the film thickness is measured for a seal with different oils having different viscosities and for different
47、 rotating speeds. Figure 7 shows the film thickness as a function of the product sliding speed u times lubricant viscosity at the running temperature. The results can be fitted with a power law function: 3.450.68film thickness u (8) displaying a R2value of more than 0.95. The result from the equatio
48、n 5 used in the model is added to the figure (in red). Equation 5 assumes a power 0.66 applied to the product (u ), which is very close to the numerical fit (equation 8). This shows a very good agreement qualitative between the theoretical formula and the measured film thickness, validating the appr
49、oach in the model. Model validation: seal friction and temperature Measurements and seal friction and temperature calculations are performed for molded wave seals and trimmed plain lip seals (HMS 5 RG and V seals, see Figure 8). The two seal types are standard seals used in industrial applications such as in gearboxes. Figure 7. Measured film thickness for different oils with different viscosities and sliding speeds (points). Power fit of the experimental results (in black) and prediction by equation (5) (in red
