Modeling and Simulation for Material Selection and Mechanical Design Part 17 potx

Normally, the screw shank is the location of fatigue failure, but the clamped part or nut thread component can also end in fatigue failure, e.g., thin sheet metals as clamped part and a

with rolled thread after heat treatment depends significantly on the preload level (reason: strain hardening and residual stresses from thread rolling are not compensated by a new grain structure from heat treatment, so nonlinear profiles from loading stresses and residual stresses are superposed) The test principle for determining axial fatigue load saspermt2is defined

in ISO 3800 [30] or more detailed in DIN 969 [10] for threaded fastening ele-ments Normally, the screw shank is the location of fatigue failure, but the clamped part or nut thread component can also end in fatigue failure, e.g., thin sheet metals as clamped part and a screw head with locking teeth If the location of fatigue failure is at screw head fillet, significant bending of screw

is probable (see alsoFig.54)

Minimizing fatigue problems can be realized by reducing the screw stressing (e.g., larger screw size, lower additional force for screw in fastening system), proper screw section design (see Fig 2, e.g., sufficient radius at head-to-shank-fillet, perpendicularity of screw axis and head support, smooth transition of each discontinuity at screw shank, such as different diameters of screw, running out of thread to unthreaded shank), no overlap-ping of stress concentrations (e.g., chamfer at clamped part under head or at

Figure 53 Typical cross-section of screw M18 1.5-12.9 failed in fatigue, result of testing procedure according ISO 3800, stress amplitude sa¼ 80 MPa, mean stress

555 MPa, symmetric axial force without bending moment

first nut thread flank, avoiding corrosive pittings at thread flanks or at head-shank-transition) The most established actions to increase the fatigue limit of the screw itself are discussed inFig 41

Figure 53 contains a cross-section of a screw M18 1.5 which has failed in fatigue A plane fracture zone can be seen at outer regions of the cross-section (area of crack initiation and crack propagation) and an unplane fracture area in the center of the cross-section (residual area of rapid failure under preload)

Figure 54in contrast to Fig 53, gives an impression of a fatigue failure with significant bending moment under external loading This result was obtained with a transversal vibrational test (see alsoFig 76).Now, the area

of crack initiation and crack propagation with ‘‘cycle lines’’ is clearly differ-ent from the residual fracture area The size of this second part of cross-sec-tion gives informacross-sec-tion whether the acting preload at the event of failure was high or not (fatigue failures often are induced by wrong tightening or loss of preload caused by relaxation or self-loosening)

F Aspects of Quality Management

The overall objective for quality aspects of a threaded fastening system is

to guarantee sufficient preload This preload normally is not specified directly Due to this reason, a large number of details must fit together, which have to be realized by different responsibilities Figure 55 demon-strates seven main groups of authorities which have to give their contri-bution to quality of the fastening system The drawn boxes make clear that every authority has its objective, its risk and takes actions due to different criteria Besides this, it is important for clarifying failures that each authority in most cases belongs to different business units or com-panies, so various communication interfaces exist, which have to work without deficit or error So, from the point of organization, a ‘‘fastening manager’’ is recommended

A few more quality aspects are:

The calculation of threaded fastening systems in any case is an approximation because numerous details have to be estimated like real screw loading in the system, local fatigue strength of screw, material inhomogeneities, real external loading spectrum or others

By these uncertainties, the compatibility of a designed fastening system with experience from former solutions is valuable, and critical calculations have to be verified by experimental testing But proper designing due to guidelines of this chapter avoids a lot of failure situations and reduces testing expense

every screw failure during operation, the assembly process has to be ana-lyzed too

Always, three general reasons for failures during operation have to be considered:

1 wrong initial preload (tightening process, poor design),

2 wrong residual preload (relaxation by creeping of materials, gaskets), and

3 overloading mechanically, thermal or reactive (too high operating load with plastification or sliding, too high temperature with creeping or decreasing of strength, too strong environment with significant corrosion)

G Cost Accounting of Fastening Systems

Always cost accounting of a fastening system has to be done due to life cycle

of the component system (product) because only this life cycle cost can be compared to the customer value Then, all boundary conditions ofFigs 1 and 32 have to be included and evaluated monetarily—the fastening element

is only one contribution to this life cycle cost.Figure 56proposes a funda-mental approach to total cost accounting, which takes into account the main types of cost related to a fastening system

Options for cost optimizing are thread rolling(Fig 7), standard mate-rials (Fig 10), coarse tolerances for geometry (Table 2) But one must always remember that the guaranteeing of reliable function has to be of higher priority than the cost for a technical system Otherwise, the product has no customer value and therefore no market

IV EXAMPLES OF DESIGN

A Fastening with Optimized Initial Preload

Traditionally, screws are tightened with torque control (see Fig 19) The tightening torque Ttot is specified for the conditions with lowest friction This is shown on the left side ofFig 57(as a supplement to Fig 51)for a steel screw 8.8, tensile strength of 800 MPa The first case A considers a low friction situation with coefficients of mt¼ mh¼ 0.08 This screw at

20 N m tightening torque is stressed up to 0.9 Rp0.2 (rhomb marking) and produces a preload of almost 20 kN Because of high screw stressing, the torque specification cannot be increased over 20 N m if yielding or breaking of screw has to be avoided in any case

For the same screw in large series assembly lines, the frictional situa-tion can change to mt¼ mh¼ 0.16 (curve B) Then, for the same tightening

defined minimum plastification before fracture, therefore, the manufacturing process must be optimized In principle, using yield point controlled tighten-ing, the screw cannot be overloaded Using angular controlled tightentighten-ing, the screw cannot be overloaded, if the snug torque is low enough (e.g

10 N mþ 908 inFig.57).Figure 26explains why also a screw tightened with angular control can be loaded additionally

As a side-effect, the diagrams in Fig 57 make it clear that torque-con-trolled tightening (and a steel screw for aluminum components) is an ‘‘old-fashioned’’ and not very optimized solution from the viewpoint of engineer-ing threaded fastenengineer-ing systems The diagrams confirm that a torque value mentions ‘‘nothing’’ about the preload acting in the joint

B Fastening with Small Thread Engagement

For every component design, the necessary minimum thread engagement needs space and therefore generates component weight A minimization of thread engagement is required Figure 33 contains the basic mechanics and suggests the use of relatively high-strength nut thread material or the use of relatively low-strength screw material

Figure 58 Measured low minimum thread engagement for AluformTMscrews M6 (From Ref 15.)

Figure 58gives a practical verification of the calculation fromFig 33 for a pull-out-test with RIBE-Aluform# screws [60] M6 (Rm> 400 MPa) engaged to an aluminum nut thread plate (Rm¼ 300 MPa) with certain lengths of thread engagement te The bar diagram shows the maximum pull-out-forces Fzmax in the event of failure The upper level of 9 kN belongs to a tensile screw breaking in the screw shank The lower level belongs to nut thread stripping

The transition begins exactly at the point 0.7 d which is also pre-dicted by Fig 33 (do not forget to consider chamfer of 1 P in Fig 33)

So, indeed Aluform#screws have a reliable behavior against stripping also for low-strength nut thread components and low thread engagements te, which would never be fulfilled with a steel screw 8.8 or higher

C Fastening of High-strength Components

High-strength components provide the chance for small threaded fastening systems because contact pressure at screw head as well as thread flanks can

be in the range between minimum of Rp0.2and Rmof the materials in con-tact Besides this, hard surfaces are almost unaffected by roughness

chan-Figure 59 Screw head design with small head diameter for tightening on hard surface

ging with adhesive or abrasive wear meachanisms, so that the frictional situation is constant over a wide range of preload

Figure 59demonstrates a small-diameter screw head design for tigh-tening on hardened component surface in the range of Rm¼ 1400 MPa (for high-strength screw materials seeTable 9) Such a screw of dimension M11 1.5 and a screw tensile strength of 1150 MPa produces an initial pre-load of about 60 kN This leads to a mean contact pressure of 950 MPa (compare also diagram in Fig 39 and explanations related) In addition, for such design, the lubrication of the screw is of significant importance These screws offer the possibility for small space flanges and in consequence for a compressed design of component In contrast to light metal com‘ ponents, these high-strength materials possess significant mass density, but can be used for an extreme compact design

The same aspects are valid for thread engagement but at least

te¼ 0.8  d should be realized to avoid stripping of screw thread flanks (compare asymptotic behavior of Fig 33 for high nut thread strength

Rmnut) If fastening high-strength components, precise support geometries and small contact roughness is required, then peak contact pressure is avoided, which can be the origin of crack propagation and fatigue failure

of the component

Threaded fastenings with components made of high-strength materials provide the possibility for meeting small space requirements (low thread engagement, small head diameter, small screwing boss diameter at clamped part) If this is combined with overelastic tightening and reliable lubrication, then also lightweight fastening is possible

D Fastening of Components Made of Brittle Materials

‘‘Brittle’’ means that a material has a low ductility before fracture, which leads to a sudden rupture without plastified deformation in the case of ten-sile testing or overloading a component As a guide, the fracture toughness from tensile test is for brittle materials smaller than 3–5%) For such a component (e.g., made of magnesium, titanium with hexagon crystal struc-ture or cast iron with high carbon content or ceramic materials), not only

is the mean stressing important, but also all local stress peaks have to be minimized

Therefore, thread engagement of a screw in brittle materials should be increased by at leastþ20% due to Fig 33 because of the inhomogeneous stress distribution up to the event of fracture, which is not compensated

by local plastification of the nut thread flanks

Brittle materials of low strength (cast iron, magnesium) tend to pro-duce increased adhesive-abrasive wear in the screw head contact zone

during tightening (Fig 60) This results in increased roughness, particles and undefined contact conditions, so the head frictional torque is increased and—if the screw is tightened by torque control—the preload

is reduced significantly To avoid this, the lubrication of screws for brittle materials should be enhanced

Use of thread rolling screws in brittle materials is critical Figure 61 presents an example from thread rolling with a high-performance thread rolling screw M8 (induction hardened forming point) in high-strength duc-tile gray iron GGG 50 The result is that particles of nut thread material are produced in an unacceptable amount They lead to poor nut thread quality

as well as screw thread damage and therefore to insufficient process capabil-ity for series production (see torquing diagram in Fig 61 with temporary breakdown of torque curve)

E Fastening of Light Metal Components

Light metals, such as aluminum and magnesium, are characterized mechani-cally by low strength and high thermal expansion coefficient Especially in

Figure 60 Adhesive–abrasive wear of head support area after angular controlled tightening of screw M14 2-11.9, specification 150 N m þ 908 þ 908 þ 908, preload app 90 kN, surface gray cast iron GG25

tightening or angular controlled tightening(Fig 26) and a screw with threaded screw shank(Fig 3)

4 If operating at elevated temperatures, a special adaptation of thermal fit for minimization of thermal stress increase is necessary (often screws made of aluminum are an effective alternative compared to steel screws,Fig 29)

5 For aluminum components, thread rolling screws are widely used (Fig 7, Fig 62)

6 For corrosion stability, use enhanced corrosion protection for steel screws or use aluminum screws(Fig 63)

Figure 64 contains the fastening of a magnesium component with thread rolling screw made of aluminum in two columns: the left side refers

to five repetitions of screwing with same screw into the same nut thread hole The right side refers to the situation where the same screw is screwed into a new pilot hole without nut thread for each repetition

The images of the screwing bosses confirm a high quality of the pro-duced nut thread in magnesium for both columns of Fig 64 The diagram

is very detailed due to the formation of positive torque and prevailing torque (negative) All values are proposed for 1 to 5 screwing operation

Figure 62 Trilobular stud M6 32 for thread rolling in aluminum component, dry lubrication, maximum thread forming torque 3 N m, tightening torque for stud

10þ 0.5 N m, thread engagement 11 mm, casted pilot hole with diameter 5.4–5.6 mm

Figure 65 Fastening of aluminum metal foam with screw–sleeve-combination, see alsoFig 44

Figure 66 Thermal decrease of preload by relaxation of threaded fastening systems with magnesium components at elevated temperatures; measured data (From Ref 16.)

Every case, from 1 to 7, shows the initial preload for steel screw (light bar) and aluminum screw (dark bar) before and after thermal exposure For case 1 (screw M10), this means 32 kN resp 19 kN before thermal exposure and 5 kN resp 9 kN after exposure The other cases confirm similar behavior The most extensive preload relaxation occurs for fastening systems with the widely used magnesium alloy AZ91 (high-strength alloy with

9 wt.% aluminum and 1 wt.% zinc; relatively stable corrosive behavior; low creeping resistance at temperatures above 1208C) Cases 3 and 5 confirm where the steel screw leads to almost no preload after thermal exposure— here aluminum screws are the only solution for reliable fastening systems

In general, for these situations, an aluminum screw always gives a higher residual preload than a steel screw

F Fastening of Components with Thread Rolling Screws

What is the difference between thread rolling screws and screws for existing nut thread in practice? The fundamental principle of thread rolling screws is

Figure 67 Functional behavior of thread rolling screw compared to a screw for existing nut thread used for the same application, see also Ref 63