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When it comes to the selection of lube oil, engineers and technicians are facing a choice between SAE Viscosity grades and different additive packages. One might think it’s pretty simple, pick the most expensive AW/AF/EP package, and you are set! Sounds easy, right? The reality, however, is somewhere far away. An effective lubricant should not only lubricate but also protect working surfaces from wear, shock, and, at the same time, minimize equipment downtime. As a result, the overall cost of ownership should not go up; it should be down. This is where the proper, cutting-edge, and ready to act additive should be in play.
CASE STUDY: Sugar Mill Bearings
The regime of lubrication in Sugar mill bearings changes due to misalignment, contamination, vibration, and metallic detachment inherent to this type of application. Therefore, conventional lubricants often fail to do their job, even when the “correct” lube was chosen for this specific application. To solve the issue, one must find a balance between lubricant quality, proper use, and… Math! First, we should define all of the variables, including the critical thickness of the protective layer left on the surface by a lube oil.
In this case study, the λ value calculated to define the regime of lubrication combined with the surface roughness of the assembly. The result of the calculation demonstrates the position on the Stribeck curve and shows the extreme pressure EP level of the lubricant additive.
– SKF bearing lubricated by ISO32 // Reference 23152 CACK/CC83W507
– External diameter: 440 mm
– Internal diameter: 260 mm
– Rotation operation: 1000 rpm
– Operating temperature: 53.09 oC
– Working hours: 8000 h/year
– Housing dimensions: H 860 mm, W 245 mm, Thickness 580 mm
- The Calculations
There is a certain beauty behind tribological evaluation of a lubricant that could be the most effective for a specific application. Here is why.
λ = ho / σp
We must find the ho and σp
ho = CD / (Ln)0,74
C = Geometric factor of bearing. In this case it is 8,01×10-4. See Table 1
D = External dimension of the bearing. In this case, it is 0,44 m. Given.
L = Viscosity parameter of the lubricant. In this case, it is 15 mm2/s. See Chart 1
n = Angular speed of the bearing. In this case, it is 1000 rpm. Given.
So ho = (8,01×10-4 x 0,44) x (15 x 1000) 0,74
So ho = 0,43 μm. That is the lubricant film thickness between the average line of surface roughness body 1 and body 2.
Now we calculate σp
σp = (σ12 + σ22)0,5
σ1 = Average surface roughness of body 1. In this case 0,05 μm. See Table 2
σ2 = Average surface roughness of body 2. In this case 0,05 μm. See Table 2
Note: For double ball bearing N=2
So σp = (0,052 + 0,052)0,5
So σp = 0,070 μm. That is the distance between the average line surface roughness of body 1 to body 2.
Now we can calculate the λ factor: .
λ = ho / σp
λ = 0,43 / 0,07
λ = 6,14
- The Conclusion
According to the range of λ in Table #3, the lubricant will work in the regime of operation called Hydrodynamic. Also, we were able to define the EP level at 28<kgf<150, according to Table #4.
Many factors mentioned above trigger the effective regime change from HD to EHL and even kick it to the boundary limit of the Stribeck curve. In this case study, a lubricant that does not meet the required stats, will NOT protect the equipment in a real-life setting. The solution to this lubrication problem should be an effective additive package. This additive package should be adaptable to the working regime changes. This is the only way to protect equipment in harsh conditions.