International Journal of Research and Innovation (IJRI)
International Journal of Research and Innovation (IJRI) 1401-1402
A REVIEW ON CONTACT STRESS ANALYSIS OF HELICAL GEAR BY USING FINITE ELEMENT ANALYSIS
Dr. Kareem Dakhil Jasym1, 1 Research Scholar, Department Of Mechanical Engineering, AL-Qadissiya University college of engineering, Iraq.
Abstract
Gears are mainly used to transmit the power in mechanical power transmission systems. These gears play a most predominant role in many automobile and micro electro mechanical systems. Generally gear transmission failures due wear and tare of gear which occurs because of the contact between two gears while power transmission process is started, Due to meshing between two gears contact stresses are developed, these can be evaluated using FEM Based calculation’s or software’s like Ansys. Finding stresses has become most popular in research on gears to minimize the vibrations, bending stresses and also reducing the mass percentage in gears. These stresses are used to find the optimum design in the gears which reduces the chances of failure. The model is generated by using CATIAV5 and ANSYS is used for numerical analysis. Contact pair analysis will be conducted by varying the geometrical profile of the teeth and to find the change in contact stresses between gears. Calculated results and ANSYS results will be compared to provide variation percentage and better solution to reduce contact stresses.
*Corresponding Author: Dr. Kareem Dakhil Jasym, Research Scholar, Department Of Mechanical Engineering, AL-Qadissiya University college of engineering, Iraq. Published: September 01, 2015 Review Type: peer reviewed Volume: II, Issue : V Citation: Dr. Kareem Dakhil Jasym , Research Scholar (2015) A REVIEW ON CONTACT STRESS ANALYSIS OF HELICAL GEAR BY USING FINITE ELEMENT ANALYSIS
INTRODUCTION Gears are most commonly used for power transmission in all the modern devices. The toothed wheels are used to change the speed or power between input and output. They have gained wide range of acceptance in all kinds of applications and have been used extensively in the highspeed marine engines. In the present era of sophisticated technology, gear design has evolved to a high degree of perfection. The design and manufacture of precision cut gears, made from materials of high strength, have made it possible to produce gears which are capable of transmitting extremely large loads at extremely high circumferential speeds with very little noise, vibration and other undesirable aspects of gear drives. A gear is a toothed wheel having a special tooth space of profile enabling it to mesh smoothly with other gears and
power transmission takes place from one shaft to other by means of successive engagement of teeth. Gears operate in pairs, the smallest of the pair being called “pinion” and the larger one “gear”. Usually the pinion drives the gear and the system acts as a speed reducer and torque converter. ADVANTAGES OF GEAR DRIVES: The following are the advantages of the gear drives compared to other drives. • Gear drives are more compact in construction due to relatively small centre distance. • Gears are used where the constant velocity ratio is desired. • Gears can be operated at higher speeds. • It has higher efficiency. • Reliability in service. • It has wide transmitted power range due to gear shifting facility. • Gear offers lighter loads on the shafts and bearings. • Gear can change the direction of axis of rotation. DISADVANTAGES OF GEAR DRIVES: • Not suitable for the shafts which are at longer center distance. • Manufacturing is complex. It needs special tools and equipment. • Require perfect alignment of shafts. • Requires more attention to lubrication. • The error in cutting teeth may cause vibration and noise during operation. 178
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HELICAL GEARS
Helical Gear geometrical proportions
Helical gears offer a refinement over spur gears. The leading edges of the teeth are not parallel to the axis of rotation, but are set at an angle. Since the gear is curved, this angling causes the tooth shape to be a segment of a helix. Helical gears can be meshed in a parallel or crossed orientations. The former refers to when the shafts are parallel to each other; this is the most common orientation. In the latter, the shafts are non-parallel.
• • • • • • • • • • • •
The angled teeth engage more gradually than do spur gear teeth causing them to run more smoothly and quietly. With parallel helical gears, each pair of teeth first make contact at a single point at one side of the gear wheel; a moving curve of contact then grows gradually across the tooth face to a maximum then recedes until the teeth break contact at a single point on the opposite side. In spur gears teeth suddenly meet at a line contact across their entire width causing stress and noise. Spur gears make a characteristic whine at high speeds and can not take as much torque as helical gears. Whereas spur gears are used for low speed applications and those situations where noise control is not a problem, the use of helical gears is indicated when the application involves high speeds, large power transmission, or where noise abatement is important. The speed is considered to be high when the pitch line velocity exceeds 25 m/s. Quite commonly helical gears are used with the helix angle of one having the negative of the helix angle of the other; such a pair might also be referred to as having a right-handed helix and a left-handed helix of equal angles. The two equal but opposite angles add to zero: the angle between shafts is zero – that is, the shafts are parallel. Where the sum or the difference (as described in the equations above) is not zero the shafts are crossed. For shafts crossed at right angles the helix angles are of the same hand because they must add to 90 degrees. Helical gear nomenclature Helix angle, ψ Angle between a tangent to the helix and the gear axis. Is zero in the limiting case of a spur gear. Normal circular pitch, pn Circular pitch in the plane normal to the teeth. Transverse circular pitch, p Circular pitch in the plane of rotation of the gear. Sometimes just called "circular pitch". pn = pcos(ψ)
p = Circular pitch = d g. p / Z g = d p. p / Z p p n = Normal circular pitch = p .cosβ P n =Normal diametrical pitch = P /cosβ p x = Axial pitch = pc /tanβ m n =Normal module = m / cosβ α n = Normal pressure angle = tan -1 ( tanα.cos β ) β =Helix angle d g = Pitch diameter gear = z g. m d p = Pitch diameter pinion = z p. m a =Center distance = ( z p + z g )* m n /2 cos β a a = Addendum = m a f =Dedendum = 1.25*m
MARINE ENGINES RECIPROCATING DIESEL ENGINES: The engine used in 99% of modern ships is diesel reciprocating engines. The rotating crankshaft can power the propeller directly for slow speed engines, via a gearbox for medium and high speed engines, or via an alternator and electric motor in diesel-electric vessels. The reciprocating marine diesel engine first came into use in 1903 when the diesel electric river tanker Vandal was put in service by Branobel. Diesel engines soon offered greater efficiency than the steam turbine, but for many years had an inferior power-to-space ratio. Diesel engines today are broadly classified according to: (i) Their operating cycle (a) two-stroke (b) four-stroke (ii) Their construction (a) Crosshead (b) trunk (c) opposed piston (iii) Their speed • Slow speed: any engine with a maximum operating speed up to 300 revs/minute, although most large 2-stroke slow speed diesel engines operate below 120 revs/minute. Some very long stroke engines have a maximum speed of around 80 revs/minute. The largest, most powerful engines in the world are slow speed, two stroke, and crosshead diesels. • Medium speed: any engine with a maximum operating speed in the range 300-900 revs/minute. Many modern 4-stroke medium speed diesel engines have a maximum operating speed of around 500 rpm. • High speed: any engine with a maximum operating speed above 900 revs/minute. Most modern larger merchant ships use either slow speed, two stroke, crosshead engines, or medium speed, four stroke, trunk engines. Some smaller vessels may use high speed diesel engines. The size of the different types of engines is an important factor in selecting what will be installed in a new ship. Slow speed two-stroke engines are much taller, but the area needed, length and width, is smaller than that needed for four-stroke medium speed diesel engines. As space higher up in passenger ships and ferries is at a premium, these ships tend to use multiple medium speed engines resulting in a longer, lower engine room than that needed for two-stroke diesel engines. Multiple engine installations also give redundancy in the event of mechanical failure of one or more engines and greater efficiency over a wider range of operating conditions.
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As modern ships' propellers are at their most efficient at the operating speed of most slow speed diesel engines, ships with these engines do not generally need gearboxes. Usually such propulsion systems consist of either one or two propeller shafts each with its own direct drive engine. Ships propelled by medium or high speed diesel engines may have one or two (sometimes more) propellers, commonly with one or more engines driving each propeller shaft through a gearbox. THEORETICAL DESIGN CALCULATIONS CALCULATIONS FOR STEEL [40 Ni2 Cr1 Mo28 STEEL]: Input parameters: Power = P=9000 KW Speed of Pinion: N= 3500 r.p.m Gear Ratio, i=7 Helix Angle, β=25° Material Selection: The material for pinion & Gear is 40 Ni2 Cr1 Mo28 steel. Its design compressive stress & bending stresses are [σc=11000kgf/cm2 ], [σb=4000kgf/cm2] Properties for 40 Ni2 Cr1 Mo28 Steel: Density of 40 Ni2 Cr1 Mo28 Steel (ρ) = 7860 kg / m3 Young’s Modulus = 215*10 5 N /m m2 Poisson’s ratio (ν) = 0.30 i= 7 =(97420 x 9000 x 1.8)/3500 = 325661.14 kgf-cm Now, minimum centre distance based on the surface compressive strength is given by a≥(i+1) 3√[0.7/σc]2 x E[Mt]/(iψ) a≥88.41 cm Minimum module based on beam strength: Lewis form factor YV =0.154 -0.912/ ZV =0.4205 mn ≥1.15cos25x3√[ 325661.14/(0.4205*4000*10*18 )] No of teeth on gear, Z2 =i Z1 =126 Diameter of the pinion= (mn* Z1 )/ cos β =1.8*18/cos25° = 35.74 cm Diameter of the gear= i*d1 =7*35.74=250.24 cm Centre distance, a = (d1+ d2)/2 =142.99 cm Face width, b= ψa= 0.3*142.99= 43 cm (or) b= ψm mn =10*1.8=18 Taking maximum value of both b is 43 cm Checking Calculations: σc =0.7*[( i+1)/a ] √(i+1)*E*[MT] / (ib) ≤ [σc ] σb = 0.7(i+1) [MT] / (ab mn YV ) ≤ [σb] Based on the compressive stress : σc = [0.7*8*√(8*2.15*10 6 *325661.14) / (7*43)]=180.90 N/mm2
Based on the bending stress : σb = [(0.7*8*325661.14) / (88.4*43*0.4205)]=228.35 N/ mm2 From the calculations, σc & σb values are less than the [σc ] & [σb] values of given material, i.e 40 Ni2 Cr1 Mo28. Therefore our design is safe. Addendum, mn =18 mm, Dedendum=1.25* mn=22.5 mm, Tip circle diameter of the pinion= d1+(2*addendum) =357.4+3.6 = 393 mm Tip circle diameter of the gear= d2+(2*addendum) =2502+3.6 = 2538 mm Root circle diameter of pinion= d1-(2*addendum) =357.43.6 = 312.4 mm Root circle diameter of gear= d2-(2*addendum) =25023.6 = 2457.4 mm When the gear transmits power P, the tangential force produced due to the power is given by, FT=(P*KS)/V V=(π DP NP )/(60*1000) = (357.4*3500* π)/(60*1000)=65.42 m/s FT=(9000*1000*2)/65.42=275145.21 N The velocity factor CV is developed by Barth. It depends on the pitch line velocity and the workmanship of the manufacturer and it is given by CV = (5.5+√V)/5.5 for V>20m/s. FD=275,145*2.47 = 679771.90 N According to Lewis equation, the beam strength of helical gear tooth is given by FS= [σb]*b* π* mn* YV = 4000*43* π*1.8*0.4205=4089938.94N Since, FS > FD, Our design is safe. Considering the above conditions, Buckingham derived an equation to find out the maximum load acting on the gear tooth which is given by FD= FT + FI Where FD = Maximum Dynamic Load FT =Static load produced by power FI=Incremental load due to dynamic action Incremental load depends on the pitch line velocity, face width of a gear tooth, gear materials, accuracy of cut and the tangential load and it is given by FI= [0.164Vm (cbcos2β+ FT) cosβ]/ [0.164Vm+1.485√ cbcos2β+ FT] Where Vm is the pitch line velocity in m/s b is the face width of the gear tooth in mm c is the dynamic factor or deformation factor in N/mm. Deformation factor c can also be selected from table 41, 42 PSG 8.50 Here c = 11860*e, c= 11860*0.026=308.36 N/mm. FT=137572.60 N 180
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Vm=65.42 m/s=65.42*103 mm/s b=38.16 cm=381.6mm β=25° FD = FT + FI = FT + [0.164Vm (cbcos2β+ FT) cosβ] / [0.164Vm+1.485√ cbcos2β+ FT] =140776.62 N FW = (D1*b*Q*KW)/cos3β = (357.4*430*1.75*2.553)/cos225 =900086.75N FD=140776.62N Since FW > FD our design is safe.
Modifyed model
DESIGN OF HOLE: Design is based on the torque transmitted by the shaft Torque transmitted T= (60*P)/2πN= (60*9000*103)/(2π*3500)=24555.3N-mm We know that torque transmitted by the shaft T= (πd3*fs)/16=d=125 mm Key dimensions: Width of key (w) =d/4=32 mm Thickness of key (t) =d/6=18 mm Introduction To CAD/CAM/CAE The Modern world of design, development, manufacturing so on, in which we have stepped can’t be imagined without interference of computer. The usage of computer is such that, they have become an integral part of these fields. In the world market now the competition in not only cost factor but also quality, consistency, availability, packing, stocking, delivery etc. So are the requirements forcing industries to adopt modern technique rather than local forcing the industries to adapt better techniques like CAD / CAM / CAE, etc. The Possible basic way to industries is to have high quality products at low costs is by using the computer Aided Engineering (CAE), Computer Aided Design (CAD) And Computer Aided Manufacturing (CAM) set up. Further many tools is been introduced to simplify & serve the requirement CATIA, PRO-E, UG are some among many. CAD: Computer Aided Designing (Technology to create, Modify, Analyze or Optimize the design using computer. CAE: Computer Aided Engineering (Technology to analyze, Simulate or Study behaviour of the cad model generated using computer. CAM: Computer Aided Manufacturing (Technology to Plan, manage or control the operation in manufacturing using computer. Model:
Introduction to ANSYS This Tutorial will use a readymade file to speed up the learning process for the student. This file is provided in Parasolid format. The intention of this tutorial is to get the student to run a straight forward simulation. By the end of this tutorial a check list for the required procedure can be formulated by the student. ANSYS as a software is made to be user-friendly and simplified as much as possible with lots of interface options to keep the user as much as possible from the hectic side of programming and debugging process. Why is it that such a simple model is used? During this tutorial a simple geometry is used, the objective of that is that the student masters the steps to get to run a simple simulation, once that’s done the student can model any kind of geometry he sees necessary for his studied case. DEFINITION OF PROBLEM Marine engines are among heavy-duty machineries, which need to be taken care of in the best way during prototype development stages. These engines are operated at very high speeds which induce large stresses and deflections in the gears as well as in other rotating components. For the safe functioning of the engine, these stresses and deflections have to be minimized. In this project, we have performed static-structural analysis on a high speed helical gear used in marine engines, using different materials. The results obtained in the static analysis are compared with those obtained in theoretical and a conclusion has been drawn on the material to be used for the gear. ELEMENT CONSIDERED FOR STATIC ANALYSIS: According to the given specifications the element type chosen is SOLID 45. SOLID45 is used for the 3-D modeling of solid structures. The element is defined by eight nodes having three degrees of freedom at each node: translations in the nodal x, y, and z directions. The element has plasticity, creep, swelling, stress stiffening, large deflection, and large strain capabilities. The element is defined by eight nodes and the orthotropic material properties. Orthotropic material directions correspond to the element coordinate directions. The element coordinate system orientation is as described in Coordinate Systems. The following Figure shows the schematic diagram of the 8-noded static solid element.
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SOLID-45, 8-NODE STATIC ELEMENT
ASSUMPTIONS AND RESTRICTIONS: Zero volume elements are not allowed. Elements may be numbered or may have the planes IJKL and MNOP interchanged. Also, the element may not be twisted such that the element has two separate volumes. This occurs most frequently when the elements are not numbered properly.
The above image shows stress
All elements must have eight nodes. A prism-shaped element may be formed by defining duplicate K and L and duplicate O and P node numbers. A tetrahedron shape is also available. The extra shapes are automatically deleted for tetrahedron elements. MATERIALS
YOUNGSMODULES (N/mm2)
DENSITY (Kg/mt3)
POISSIONSRATIO
Steel
215*105
7850
0.30
Aluminum alloy
340*105
3900
0.220
The above image shows strain RESULTS FOR CERAMICS:
BOUNDARY CONDITIONS: A) Geometry boundary conditions: The shaft is fixed at the centre along with its key. B) Loads applied: FT=137572.60 N FR =55248.70 N Finally the boundary conditions are verified before going for a solution. ANSYS RESULTS RESULTS FOR STEEL:
The above image shows total deformation
The above image shows total deformation
Modified steel
The above image shows stress
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International Journal of Research and Innovation (IJRI)
Modified ceramic
CONCLUSION The present project work deals “with CONTACT STRESS ANALYSIS OF HELICAL GEAR BY USING FINITE ELEMENT ANALYSIS” to provide variation percentage between analysis and calculations for better solution to reduce contact stresses.
The above image shows stress
DISCUSSION OF RESULTS RESULTS FOR STEEL AS MATERIAL: The following table shows the comparison between theoretical results & experimental results PARAMETER
DESIGN STRESSES
INDUCED STRESSES
BENDING STRESS
400 N/mm2
228.35 N/mm2
COMPRESSIVE STRESS
1100 N/mm2
178.59 N/mm2
COMPARISION OF RESULTS PARAMETERS
STEEL
CERAMICS
MODIFIED STEEL
MODIFIED CERAMIC
DEFLECTIONS
0.0001278
0.00011502
3.3167e-5
2.985e-5
STRESS
2.516
2.2644
2.4799
2.2319
STRAIN
2.0421e-5
2.1619e-5
1.3832e-5
1.244e-5
GRAPHS
Data collection and letrecher survey is done to understand approach for theoretical and numerical calculations. Theoretical calculations are done to find out gear parameters and load conditions, Analysis work is done to validate design and materials & geometry. Contact stress analysis is done to find the stresses at the contact region by varying material and tooth profile. As per the above results THEORETICAL RESULTS& ANSYS RESULTS are having negligible variation in BENDING STRESS, COMPRESSIVE STRESS and von-misses stress is having 60% less stress than design stresses. By observing all the results Modified ceramic is the best solution to reduce contact stresses and modifications in tooth profile will reduce stress intensity. REFERENCES 1. Thirupathi R. Chandrupatla & Ashok D.Belegundu ., introduction to finite eleement in engineering, pearson ,2003 2. Joseph Shigley, Charles Mischike ., Mechanical Engineering Design, tmh,2003 3. Maithra ., handbook of gear design,2000 4. V.B.Bhandari., design of machine elements,tmh,2003 5. R.S.Khurmi ., machine design, schand,2005 6. Darle w dudley.,hand book of practical gear design,1954 7. Alec strokes., high performance gear design,1970 8. Khurmi, R.S, Theory of Machines, S.CHAND 9. Schunck, Richard, "Minimizing gearbox noise inside and outside the box.", Motion System Design. 10. Doughtie and Vallance give the following information on helical gear speeds: "Pitch-line speeds of 4,000 to 7,000 fpm [20 to 36 m/s] are common with automobile and turbine gears, and speeds of 12,000 fpm [61 m/s] 183
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have been successfully used." – p.281. 11. McGraw Hill Encyclopedia of Science and Technology, "Gear", p. 742.
19. Kravchenko A.I., Bovda A.M. Gear with magnetic couple. Pat. of Ukraine N. 56700 – Bul. N. 2, 2011 – F16H 49/00.
12. Canfield, Stephen (1997), "Gear Types", Dynamics of Machinery, Tennessee Tech University, Department of Mechanical Engineering, ME 362 lecture notes.
20. ISO/DIS 21771:2007 : "Gears – Cylindrical Involute Gears and Gear Pairs – Concepts and Geometry", International Organization for Standardization, (2007)
13. Hilbert, David; Cohn-Vossen, Stephan (1952), Geometry and the Imagination (2nd ed.), New York: Chelsea, pp. 287,ISBN 978-0-8284-1087-8.
21. Hogan, C Michael; Latshaw, Gary L The Relationship Between Highway Planning and Urban Noise , Proceedings of the ASCE, Urban Transportation Division Specialty Conference by the American Society of Civil Engineers, Urban Transportation Division, 21 to 23 May 1973, Chicago, Illinois
14. McGraw Hill Encyclopedia of Science and Technology, "Gear, p. 743. 15.
Vallance Doughtie, p. 287.
16. Vallance Doughtie, pp. 280, 296.
AUTHOR
17. Doughtie and Vallance, p. 290; McGraw Hill Encyclopedia of Science and Technology, "Gear", p. 743.
Dr. Kareem Dakhil Jasym, Research Scholar, Department Of Mechanical Engineering, AL-Qadissiya University college of engineering, Iraq.
18. McGraw Hill Encyclopedia of Science and Technology, "Gear", p. 744.
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