University of Technology PETRONAS

Failure analysis study on failed fin fan shaft assembly

from Air Cooled Heat Exchangers unit (A case study from Egyptian LNG plant)

Omar Eid Khorshed

Mechanical Engineering Program, University Teknologi PETRONAS

Abstract— This case study is focused in understanding the cause of the repetitive failures happening to machined shafts in the Air Cooled Heat Exchangers units in Egyptian LNG plant. The study was focused on the latest occurred case which was on 30th of September 2012. It involves various aspects to understand the reasons of failure, including fracture face analysis, loads, stress and torque analysis, failure analysis using Distortion Energy Theory for ductile material failure and extensive maintenance history investigation. The study showed that the repetitive failures are due to faulty machining and maintenance activities carried on the failed shafts and returning them back into operation. Taking all the discovered reasons from the extensive analysis, a list of recommendations were put to prevent such problems from occurring again.

Index Terms— Shaft, Fatigue, Distortion Energy Theory, Relief groove, Stress concentration

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1 INTRODUCTION

shaft is a rotating member, usually of circular cross-

sectional area used to transmit power. It’s supported

by bearings and support gears, sprockets, fly wheels and

rotors. It is subjected to torsion, traverse or axial loads,

acting in single or in combination. Generally shafts are not

of uniform diameter but are stepped to provide shoulders

for locating gears, pulleys and bearings. The stress on the

shaft at a particular point varies with rotation of shaft there

by introducing fatigue. Shaft and bearing failures have

been a repetitive and a well expected problem with

rotating equipment machinery in most of the Oil and Gas

plants & facilities equipment. On 30th of September 2012,

a vertical cooler fan shaft (2E – 3201B) had been broken at

the upper part just above the bearing assembly. This failure

was preceded with an exact similar shaft failure but at the

lower bearing of the fan assembly. Accompanying these

reported shaft failures, a long history of bearing failures for

the fans. Egyptian LNG has been facing repetitive failures

and reported problems from cooling fans used in the

plant’s refrigeration cycles. This repetitive failures cost the

plants massive amount of money in maintenance and

renewal, as well as downtime for process production

losing the plant a good amount of revenue and bigger

profit margins.

2 SHAFT MODEL & HISTORY

This shaft has been machined 2 years ago after an upper bearing failure. Upon the bearing failure, the shaft had been sent to workshop to get fixed. To fix, the area of the failed bearing went through welding and lathe operations to return the shaft to its original shape and dimensions. A new bearing assembly was then assembled to the new machined area. Machined and fixed shaft was then returned into operation.

As shown in figure 1, the failure of the shaft happened right above the upper bearing assembly of the shaft

Figure 1. Illustration of place of fracture

A free body diagram resembling the shaft and its features is shown in figure 2.

A

Figure 2. Free body diagram of shaft, bearing & sprockets

assembly

3 FAILED AREA INSPECTION

3.1 Fracture position:

Upon inspecting the failed shaft area shown in figure, we found that the fracture is located exactly above the upper bearing assembly.

Moreover, the fracture is exactly at the shoulder fillet between the bearing assembly seat and the upper shaft part.

Figure 3. Shaft failed area

3.2 Fracture face analysis:

Studying the fracture face provides us with leads and clues to narrow down our study and be able to identify the root cause of the failure. Figure shows the fracture face of the shaft.

Figure 4. Fracture face analysis

Analyzing the fracture face we can find the following:

Instantaneous Zone:

IZ is the roughest part of the fracture face, and it is the last part of the face that the crack reached and where

brittle failure happened, from the shape of the IZ we can understand the following:

The IZ is oval shaped with a part extending towards the circumference of the shaft.

The centered IZ shows that the fracture is caused by torsional fatigue. However, the extended part of the IZ shows that bending stress was also included.

From the size of the IZ, we can understand that the stress causing the failure was relatively low.

Crack Origins:

Crack origins have started from the circumference of the shaft, and then the crack propagated towards the middle.

The IZ part that extended towards the edge of the circumference shows that no cracks originated from this area.

Friction marks:

There are rough friction marks along the direction of rotation of the shaft, showcasing that the shaft kept on rotating even after the failure by some time. Causing the two fractured faces to keep contact and hence the friction marks occurred.

4 FAILURE ANALYSIS

4.1 Force analysis:

Figure 5. Static radial forces on shaft

FB: Bending force resulting from belt tension.

FR1: Reaction force from lower bearing assembly.

FR2: Reaction force from upper bearing assembly.

From belts, sprockets and motor assembly:

Figure 6. Center distance calculation method

Center distance between the two sprockets:

Instantaneous Zone

Friction marks

Crack Origins

FB

FR1

FR2

𝐶 =𝐾+√𝐾2−32(𝐷−𝑑)2

16 , 𝐾 = 4𝑙𝑝 (1)

Pitch Length (lp) = 154.33 in, 𝑲 = 𝟔𝟏𝟕. 𝟑𝟐 , Sub in (1)

𝐶 = 75.34 𝑖𝑛 = 1.9136 𝑚

Belt tensions:

Figure 7. Belt tensions diagram

𝑇𝑇 =(144,067)𝐷𝐻𝑃

(𝑃𝐷)(𝑅𝑃𝑀) (2), 𝑇𝑆 =

(18,008)𝐷𝐻𝑃

(𝑃𝐷)(𝑅𝑃𝑀) (3)

Where: DHP = Horsepower x Service Factor (hp)

PD = Sprocket Pitch Diameter (in) = 6.141 in

RPM = Sprocket Speed (rev/min) = 145 rpm

Service Factor for 50 hp motor = 1.15

HP = 49.6, DHP = 49.6 x 1.15 = 57.5 … Sub in (2) , (3)

𝑇𝑇 = 𝟒𝟏𝟑𝟖. 𝟏𝟖 𝑵

𝑇𝑆 = 𝟓𝟏𝟕. 𝟐𝟑𝟗 𝑵

𝐹𝐵 = (𝑇𝑇 + 𝑇𝑆) ×𝐷 − 𝑑

𝐶= 𝟐𝟎𝟒𝟗 𝑵

Turning force(Causing torque) = 𝑇𝑇 − 𝑇𝑆

= 4138.18 − 517.239 = 𝟑𝟔𝟐𝟎. 𝟗𝟒𝟏 𝑵

In Shaft assembly:

Figure 8. Radial loads on shaft

Using sum of moments around a point is zero equation

𝚺𝑴𝒄 = 𝟎

𝑭𝑹𝟏 = 𝟐𝟑𝟑𝟑. 𝟒𝟓 𝑵

Using sum of forces along a static shaft is zero equation

𝚺𝐅 = 𝟎

𝑭𝑹𝟐 = 2333.45 − 2049 = 𝟐𝟖𝟒. 𝟒𝟓 𝑵

Shaft unbalance:

Permissible residual unbalance: 𝑒𝑡 =𝑝×𝑅

𝑃=

10×𝐺

(𝑛

1000) (4)

Where, G is the balanced grade (mm/s)

P is the weight of the shaft. (Kg)

p is the unbalanced amount (grams)

𝒆𝒕 is eccentricity (μm)

R is the radius of the shaft (mm)

n is the rotational speed of the shaft (rpm)

G for drive shafts = G16 = 16 mm/s

R = 39mm

n = 226.5 rpm

P = 𝜌 × 𝑣 = 𝜌 × (𝜋𝑟2 × ℎ) = 81.718 𝐾𝑔 ≈ 100 𝐾𝑔

Sub in (4) 𝑒𝑡 =10×𝐺

(𝑛

1000)

=10×16

(226.5

1000)

= 706.4 μm

𝒑 =𝑒𝑡 × P

𝑅= 𝟏. 𝟖𝟏𝟏 𝑲𝒈

𝑭 = 𝑈𝑏 × 𝜔2 = 𝟏𝟎𝟏𝟕 𝑵

Force acts along the whole of the shaft dimension.

Torque transmission:

We can calculate the amount of torque transmitted to the

shaft from tracing the amount of torque generated by the

turning force from the motor.

Turning force = 3620.941 𝑁

Shaft sprocket pitch diameter = 39.300 in = 0.99 m ≈ 1 m 𝑻 = 𝐹 × 𝑟 = 𝟏𝟖𝟏𝟎. 𝟓 𝑵𝒎

4.2 Stress analysis:

FB

FR1

FR2 Unbalance Forces

Unbalance Forces

Shear – V Diagram

No shear force at failed

area

Bending Moment – M

Diagram

No bending moment at

failed area

Torque - T Diagram

T = 1810.5 Nm

Shear stress due to Torsion:

𝜏𝑚𝑎𝑥 =𝑇𝑟

𝐽=

16𝑇

𝜋𝑑3 (5)

Where, T is the Torque = 1810.5 Nm

d is shaft diameter = 63.5 mm

Sub in (5): 𝝉𝒎𝒂𝒙 =16𝑇

𝜋𝑑3= 𝟑𝟔. 𝟎𝟏𝟐 𝑴𝑷𝒂

4.3 Stress concentrations:

Figure 9. Stress concentration due to torque, Kt

𝐷

𝑑= 1.1,

𝑟

𝑑= 0.0472

From table, Kt = 1.51

𝝉𝒎𝒂𝒙′ = 𝐾𝑡 × 𝜏𝑚𝑎𝑥 = 𝟓𝟒. 𝟑𝟕𝟖𝟏𝟐 𝑴𝑷𝒂

4.4 Distortion Energy Theory:

From material properties of Carbon Steel 1090:

𝑺𝒀 = 𝟐𝟓𝟎 𝑴𝑷𝒂

From Von Misses Effective stress theory in the case of pure shear stress,

𝝉𝒎𝒂𝒙 = 𝟎. 𝟓𝟕𝟕𝑺𝒀 = 𝟏𝟒𝟒. 𝟐𝟓 𝑴𝑷𝒂

From the above calculations we can see that the calculated 𝝉𝒎𝒂𝒙

′ = 𝟓𝟒. 𝟑𝟕𝟖𝟏𝟐 𝑴𝑷𝒂 is within acceptable region that that of the shear stress causing failure

𝝉𝒎𝒂𝒙 = 𝟏𝟒𝟒. 𝟐𝟓 𝑴𝑷𝒂.

5 FINITE ELEMENT ANALYSIS:

Figure 10. FEA typical example

From the figure, we can understand that the stresses concentrate and multiply around the circumference of the shaft shoulder. (Stress concentration area).

6 SHAFT FAILURE FISHBONE DIAGRAM

7 CONCLUSION

Shaft failure was mainly due to torsional fatigue, the torsional fatigue for ductile materials cause a cut off fracture face feature.

This torsional stress is mainly from the torque originating from the driving motor. This is further proved by the fracture face analysis.

Failure analysis showed that design conditions is correct, therefore, an off design reason caused the failure.

Since failure occurred at the exact same area that’s been machined before, this proves that the machining and improper assembly/maintenance was the main cause of the shaft failure.

Machined area undergone welding and lathe operations, causing the area to be affected with heat treatment and inhomogeneity that differs from that of the original shaft properties, causing design conditions to differ at this area. (Higher stress concentrations, different material strength properties, etc.)

The traces of the bending loads on the fracture face is mainly due to the radial forces generated by the shaft’s unbalance.

8 RECOMMENDATIONS

8.1 Wireless vibration sensors:

The structure of a cooling tower makes collecting vibration data on the gearbox difficult and dangerous without permanently installed sensors. Because the gearbox is a typical failure point, lack of feedback on the machine’s health puts you at risk for unexpected failure.

As an example, Emerson offers the service to empower companies to:

Achieve optimal health of cooling tower fan

Prevent catastrophic failure and unplanned shutdown

Determine the best time to schedule maintenance to overhaul the asset

Shift from reactive and preventive to predictive maintenance

Diagnose the root cause of degradation and reoccurring problems

Safely monitor inaccessible cooling tower fans to keep people out of hazardous areas

Receive advanced notification of a developing problems, such as rolling element bearing defects, imbalance, and misalignment

8.2 Anti Rotation Lock:

Anti-rotation devices like the Gates Draftguard unit provide an economical solution to the two major problems created by wind milling ACHE fans. From a safety standpoint, they secure fan drives and prevent them from rotating freely when not receiving power, allowing maintenance technician’s access to the fan cage without risk of injury. Secondly, they prevent hard starts by allowing fan drives to power up from a neutral, standstill position, minimizing damage to drive components caused by shock loading.

8.3 Not to repair failed shaft:

Clearly the repetitive failure of the machined shafts poses a clear threat towards the reliability of the Air Cooled Heat Exchanger units, especially if the plant is going to be under heavy loads and production requirements increase.

Therefore, a strict decision to stop repairing failed shafts and returning them back to duty has to be made, in order to increase the lifespans of the cooling tower fans.

8.4 Redesign & selection of shaft and bearing assembly:

Even though we stood upon the root cause of the case’s shaft failure, but from a deeper perspective, the shaft was initially machined because it has failed, and it has failed because of bearing failure at the beginning. Meaning that bearing failures is the main case at hands and a full redesign and selection study needs to be performed to stand upon the optimum design and bearing selection to prevent bearings from failing repetitively.

9 REFERENCES

[1] The University of Oklahoma, Refrigeration basics and LNG course notes, 2009

[2] Deepan Marudachalam M.G, K.Kanthavel, R.Krishnaraj, “Optimization of shaft design under fatigue loading using Goodman method” International Journal of Scientific & Engineering Research Volume 2, Issue 8, August-2011

[3] Hudson Products Corp, Air Cooled Heat Exchangers technical specifications manual, 2009

[4] Neville W Sachs, “Practical plant failure analysis: A guide to understand material deterioration and improving equipment reliability”, Taylor & Francis group,

[5] Gates Corp, Gates Poly Chain GT2 Belt System Specifications manual.

[6] Hayward Gordon Ltd., Shaft Design, Section TG8, July 2000.

[7] Flygt ITT Industries, Shaft and Bearing calculation, February 2003.

[8] Kruger, Permissible Residual Balance, Technical Bulletin, 2001

[9] Deepan Marudachalam M.G, K.Kanthavel, R.Krishnaraj, Optimization of shaft design under fatigue using Goodman method, International Journal of Scientific & Engineering Research Volume 2, Issue 8, August 2011

[10] Bary Dupen, Notes for Strengths of Materials, ET 200, 2011

[11] Pilkey, Water D. Peterson's Stress Concentration Factors. Second Edition. New York:John Wiley & Sons, Inc., 1997.

[12] Emerson Process Management, Cooling Tower Monitoring: Wireless Vibration Monitoring for Motor and Gearbox Combinations, 2012

[13] Gates Corp., Solving wind milling problems on belt-driven ACHE fan systems: How to improve worker safety and reduce maintenance.