Angular contact Bearing failure in OH-2 pump

[bhaskar5150]

Hi!!

We have a single stge centrifugal pump (OH-2) wih Kero (Sp Gr - 0.71, voscocity : < 1 cP) as service in our CDU unit with a pumping temp of 200 degC, capacity : 56.5 m3/hr and a head of 81 mts. Same is the BEP condition.
MOCs : Shaft-AISI 4140, Impeller and casing-CS, Wear rings-12% Cr.

The pump has wear rings both in front & back with balance holes and is equipped with 7309BECBJ / CN (universal matched pair) angular contact bearings with back to back orientation for taking the thrust load.

The pump normally runs at 50-60% of BEP.

Problem outlook is very simple ie in a span of 6-8 months, the outer (ie coupling side) anguar contact bearing (of the pair) fails (cage and balls travels towards the inner ring having wide face, dislodging the balls/cage and thereby, completely jamming the inner ring's rotation). In other words, for this outer bearing, the contact angle of the balls w.r.t outer ring decrases and contact angle of the balls w.r.t inner ring increases. However, the innder angular contact bearing (ie towards pump side) was absolutely normal & free in rotation as observed after dismantling.

This has happened 5 times as per records and there is no early indication. Vibration is normal even before 1-2 hrs of failure and all of a sudden, the pump trips on high current. Even no early signs of fault was reported through ultrasonics / Peakview etc. Also, there is no lube oil / bearing / shaft discolourization attributed to high heat levels.

My understanding & analysis to this problem is as follows :
Since primary thrust direction is unknown but presuming that for an end suction pump with +ve suction head, primary thrust direction shall be towards the coupling side ie inner angular contact bearing should be the active bearing and the outer one being inactive. If that is the case, my active bearing is not failing as it is taking the pump's primary thrust but inactive bearing is being unloaded (relaxed) by the same amount. So, for the inactive bearing, the clearance between balls & raceways increases and because of centrifugal force, bearing balls runs in a skewed direction ie balls/cage rides further up in the raceway. A point comes when the balls along with cage rides up the raceways to such an extent that there is no clearance and therefore, no lubrication passing through that ball-race way contact and ultimately, the bearing fails.
From above, it appears that inactive bearing is not preloaded to the extent desired.

Now, my question for above is that is it possible to put some shims (may be 0.2mm) between the inner rings of the matched pair and then tighening the lock nut so that the outer inactive bearing is preloaded bit higher.

Will further preloading the outer inactive bearing will sove the issue !. Is my presumption regarding the primary thrust direction correct ?

Requested for your views of the above issue. Does any body has come across such an issue. You my please advice upon other perspectives also which I might be missing.

If required, I can share the photo of the damaged bearing.

Thanks in advance,
Bhaskar5150

 

[Artisi]

why are you assuming the thrust is towards the coupling end, for end suction with impeller balance holes the thrust could be towards in the inlet side.
You say normally runs 50-60 % of BEP, what is happening in "not" normal situations?
Have you run this past the pump manufacturer?

It is a capital mistake to theorise before one has data. Insensibly one begins to twist facts to suit theories, instead of theories to suit facts. (Sherlock Holmes - A Scandal in Bohemia.)

 

[bhaskar5150]

Because from the bearing failure of the outer bearing, it appears that the cage and balls have shifted towards the wider flange side of the inner ring. (In back to back configuration, wider flange of the inner ring is always pointing towards the coupling side). Since the the inner rings are clampled, so position of inner rings is fixed; and therefore, balls/cage has to shift.

You are right thrust may be from the other side also (ie from impeller back towards suction), in that case, active bearing shall be the outer one. If that is the case, then the inner ring of the outer bearing shall impose the load on the opposite direction ie balls/cage should move towards the thinner flange side of the inner ring. But again shifting the balls/cage towards the thinner flange side of the inner ring would make the bearing more free/unrestrive (in the sense that it will be best loaded) but that is not happening. Also, in that scenario, the non active bearing ie the inner bearing should run in a skewed manner and same results is expected in line with my original analysis.

Pump always runs at 50-60% of BEP and process is quite satble (no pressure fluctuations).

For bearing BECBJ, when mounted back to back, there should be some clearance between the outer rings and that is given in the bearing charts as axial clearnace (C) for unmounted bearing under no thrust load. But again when inner rings are clamped, there is no clearance between the outer rings; probably as the bearings are face ground flushed.

However, to mitigate the issue, can we think of putting a shim between the inner rings so that there is some gap between the outer rings (so that this shifting of balls/cage is restricted in the inactive bearing) and at the same time, the inactive bearing getting bit more preloaded.

We have raised the issue to the pump manufacturer but they are unable to provide a solution. But again we shall take up.

You may correct me if I am projecting my analysis in a wrong direction.

We will be glad to know the forum members experience on similar issue(s).

----------------
Bhaskar5150

 

[JJPellin]

You can determine the thrust direction and load based on the areas of wear rings and the balance diameter of the mechanical seal. If you show that skidding of the inactive bearing is the problem, you should not attempt any sort of shimming to load the inactive bearing. Better options would include an SKF Pumpac bearing with a 15 degree contact angle on the unloaded bearing. You could consider a bearing with a light preload like an MRC PDU. Based on your description, I am not convinced skidding is the problem. Start with the thrust calculation first.

Johnny Pellin

 

[bhaskar5150]

Sorry for the little delay. Thanks Mr. Johnny for your kind response, advice and time.

Yes although I have explained my analysis, I am also not that convinced that skidding has taken place in the outer bearing. The raceways do not have a polished track. However, fact of the matter is that every time outer bearing is failing (cage, balls gets distorted; pushes up towards the wider flange of the inner ring ie towards the coupling side) and impairs the rotation of the inner ring (shaft) and finally pump trips on high amps.

Although I am not sure regarding the thrust magnitude and direction, but irrespective of the direction, the one of the bearings (provided bearing selection is OK) should take the thrust load since mounted in back to back.

I shall definitely go by your suggestions ie to know the thrust load first. But I have never calculated / approximated the thrust load in a real life scenario although there are lot of books which explains them theoretically.

Can you please advice me any reference books or software tools or even manually (with pump and seal details as available with me) how to calculate it most precisely. Also, I could not understand the role of balance diameter on thrust load calculation.

Shall be looking ahead for your kind response & guidance.

 

[JJPellin]

Unfortunately, I don’t have ready access to my thrust tools right now. I would calculate thrust as follows: Suction pressure acting over the area of the impeller eye ring, differential pressure acting over the area difference between the eye ring and the hub ring, suction pressure acting over the area between the hub ring and the mechanical seal, atmospheric pressure acting over the area of the mechanical seal. Draw it out as a diagram being careful about the direction of each thrust force. Be careful about consistent units. The mechanical seal area is based on the seal size for a pusher seal or the balance diameter for a bellows seal. The seal manufacturer can provide the balance diameter for the seal. Do the calculation based on conditions at dead-head, rated flow and end of curve.

Johnny Pellin