PDF -Diesel Engine Combustion - MIT - CALCULATION OF CRANKSHAFTS diesel engine.pdf
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rules dnvgl docs pdf DNV cn 2002 01 CN41 3 pdf CLASSIFICATION NOTES No 41 3 DET NORSKE VERITAS Veritasveien 1, N 1322 Hłvik, Norway Tel +47 67 57 99 00 Fax +47 67 57 99 11 CALCULATION OF CRANKSHAFTS FOR DIESEL

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Publications P (Additional Rule Requirements) issued by Polski Rejestr Statków complete or extend the Rules and are mandatory where applicable

Publication No

- 8/P – Calculations of Crankshafts for Diesel Engines – 2012,

is an extension of the requirements contained in Part VI – Machinery Installations and Refrigerating Plants and Part VII – Machinery,

of the Rules for Classification and Construction of Sea-going Ships

This Publication was approved by PRS Executive Board on 29 December 2011 and enters into force on 1 January 2012

- 8/P – Calculation of Crankshafts for Diesel Engines – 2007

PRS/AW,

12/2011

SPIS TREŚCI str

- 1 General Provisions
- 1 Application
- 2 Principles of Calculations
- 3 Drawings and Particulars to be Submitted

5 5 5 6

- 2 Calculation of Stresses
- 1 Calculation of Alternating Stresses Due to Bending Moments and Radial Forces
- 2 Calculation of Alternating Torsional Stresses
- 3 Evaluation of Stress Concentration Factors

1 General

- 2 Crankpin Fillet
- 3 Journal Fillet (not applicable to semi-built crankshafts)
- 4 Outlet of Crankpin Oil Bore
- 15 15 18 18 19
- 4 ADDITIONAL BENDING STRESSES
- 5 Calculation of Equivalent Alternating Stress

1 General

- 2 Equivalent Alternating Stress

20 20 20

- 6 CALCULATION OF FATIGUE STRENGTH
- 7 CALCULATION OF SHRINK-FITS IN SEMI-BUILT CRANKSHAFTS

1 General

- 2 Maximum Permissible Hole in the Journal Pin
- 3 Necessary Minimum Oversize of Shrink-fit
- 4 Maximum Permissible Oversize of Shrink-fit
- 22 22 23 23 24
- 8 Acceptability Criteria

GENERAL PROVISIONS Application

- 1 The requirements specified in the present Publication are applicable to determining the dimensions of crankshafts of diesel engines intended for propulsion and auxiliary purposes,

where the engines are capable of continuous operation at their rated power when running at rated speed

Crankshafts which do not comply with the requirements specified in the present Publication are subject to PRS’ consideration in each particular case

- as far as practicable,

calculations/analysis or measurements,

- shall be submitted to PRS

– fatigue parameter influences are tested,

– working (actual) stresses are measured,

PRS may,

after analysis of the above particulars,

approve the crankshaft concerned

- 2 The requirements specified in the present Publication apply to solid-forged and semi-built crankshafts of forged or cast steel,

with one crank throw between the main bearings

The design of crankshaft is based on an evaluation of safety against fatigue in the highly stressed areas

The calculation is also based on the assumption that the areas exposed to highest stresses are: – fillet transitions between the crankpin and web,

as well as between the journal and web,

– outlets of crankpin oil bores

When journal diameter is equal or larger than the crankpin one,

the outlets of main journal oil bores shall be formed in a similar way to the crankpin oil bores,

otherwise separate documentation of fatigue safety may be required

The outlets of oil bores into crankpins and journals shall be so formed that the fatigue strength safety factor in way of the bores will be not less than that accepted in the above-mentioned fillets

the engine manufacturer shall submit the oil bore design supporting documentation

the calculations shall be made for: – the crankpin fillet,

- – the journal fillet

the calculations shall be made for the crankpin fillet only,

- provided the requirements,
- specified in paragraph 7

concerning the shrink-fit of semi-built crankshafts are complied with

Calculation of concentrated stresses and their superposition in accordance with the given formulae allow to calculate the equivalent alternating stresses

The equivalent alternating stress shall be compared with the fatigue bending strength of the selected crankshaft material

This comparison will show whether or not the crankshaft concerned is dimensioned adequately

The calculations shall be made for the loads occurring during the engine operation at the engine rated power and at rated speed

Drawings and Particulars to be Submitted

the engine manufacturer shall submit the following particulars and drawings: – rated power,

- – rated speed,

– type designation and kind of engine (in-line engine,

articulated-type connecting rods),

– operating and combustion method (2-stroke or 4-stroke cycle/direct injection,

- precombustion chamber,
- ) – number of cylinders,
- – crankshaft drawing,

which must contain all data in respect of the geometrical configurations of the crankshaft,

– direction of rotation (see Fig

– firing order with the respective ignition intervals and,

- where necessary,

αv driving shaft flange counter clockwise

driving shaft flange counter clockwise clockwise

- clockwise
- 3-1 Cylinders configuration

– particulars for determining alternating torsional stresses,

- see paragraph 2

– for the engine rated power and rated speed – radial components of the gas and inertia forces,

- within one working cycle,

acting on the crankpin at equidistant intervals of the crank angle

(The intervals shall not be greater than 5o

In the case of V-type engines,

the V-angle shall be integrally divisible by the intervals 6

– – – – – – – – –

and the simultaneous radial components shall be added algebraically and expressed as a combined force

the total radial force for which Pmax – Pmin in the engine working cycle is the greatest,

- shall be taken
- the permissible,
- assumed by the manufacturer,

total alternating torsional stresses from the crankshaft and engine occurring in the cylindrical parts of the crankpin and journal,

- cylinder diameter,
- stroke,
- maximum cylinder pressure,
- charge air pressure,

(before inlet valve or scavenge port,

- whichever applies),
- nominal compression ratio,

[ – ] connecting rod length,

all individual reciprocating masses acting on one crank,

for engines with articulated-type connecting rod (see Fig

- 3-2): – distance to link point,
- – link angle,
- – link rod length,

for the cylinder with articulated-type connecting-rod: – maximum cylinder pressure,

- – charge air pressure,

(before inlet valve or scavenge port,

- whichever applies),
- 3-2 Articulated-type connecting rod

mechanical properties of material (minimum values obtained from longitudinal test specimens),

required by PRS: – material designation (according to PN or EN,

- or DIN,
- or ISO or AISI,
- – tensile strength,
- – yield strength,

– reduction in area at break,

- – elongation A5,
- – impact energy KV,

– material casting process (open-hearth furnace,

- electric furnace,

– type of forging (free form forged,

- continuous grain flow forged,
- drop-forged,

with forging process description),

- – heat treatment,

– surface treatment of fillets,

journals and pins (induction hardened,

- flame hardened,
- nitrided,
- roll hardened,
- shot peened,

with full details of hardening process),

- – surface hardness,

– extension of surface hardening,

– particulars for determining alternating torsional stresses,

- see paragraph 2

The calculation is based on a statically determined system,

composed of a single crank throw,

supported in the centre of adjacent main journals

- shown in Fig

Bending moments,

are calculated in the relevant section based on triangular bending moment diagrams due to the radial component FR and tangential component FT of the connecting-rod force,

- respectively,
- see Fig

the relevant bending moments are obtained by superposition of the two triangular bending moment diagrams according to phase

Connecting rod forces (FR or FT)

Crank throw for in-line engine and for V-type engine with 2 adjacent connecting rods

L1 – distance between main journal centre line and crankweb centre (see also Fig 2

- 1 for crankshaft without overlap) L2 – distance between main journal centre line and connecting-rod centre L3 – distance between two adjacent main journal centre lines 9

Bending Moments and Radial Forces Acting in Web

The bending moment MBRF and the radial force QRF are taken as acting in the centre of the solid web (distance L1) and are derived from the radial component of the connecting-rod force

The alternating bending and compressive stresses due to bending moments and radial forces shall be related to the cross-section of the crank web

- 1 Reference area of crank web cross-section

- 2): MBRO – the bending moment of the radial component of the connecting-rod force,

The alternating stresses due to these bending moments shall be related to the cross-sectional area of the axially bored crankpin

- 2 Crankpin section through the oil bore

Generally,

the radial and tangential forces due to gas and inertia loads acting upon the crankpin at each connecting-rod position will be calculated over one working cycle

Upon PRS’ agreement,

a simplified procedure for calculating these components may be used

Using the forces calculated over one working cycle and taking account of the distance from the main bearing midpoint,

the time curve of the bending moments MBRF,

MBTO and radial forces QRF will then be calculated in accordance with paragraphs 2

1 and 2

In the case of V-type engines,

the bending moments – progressively calculated from the gas and inertia forces – of the two cylinders acting on one crank throw shall be superposed according to phase

Different designs (forked connecting-rod,

articulatedtype connecting-rod or adjacent connecting-rods) shall be taken into account

the calculation shall cover all crank variants

The decisive alternating values will then be calculated in accordance with the formula: [X N ] = ± 1 [X max − X min ] 2 where: XN – is considered as alternating force,

- moment or stress,

Nominal alternating bending and compressive stresses shall be calculated in accordance with the formulae:

σ QFN = ±

QRFN ⋅103 ⋅ K e ,

1-1) (2

where: σBFN – nominal alternating bending stress related to the web,

- 1 M BRFmax − M BRFmin ,

- [mm3] M BRFN = ±

– empirical factor considering to some extent the influence of adjacent crank and bearing restraint with: Ke = 0

- 8 for 2-stroke engines,
- 0 for 4-stroke engines,

QRFN = ±

- 1 QRFmax − QRFmin ,

- [mm2] F = B · W,

σ BON = ±

M BON ⋅103 ,

where: σBON – nominal alternating bending stress related to the crank pin diameter,

MBON – alternating bending moment calculated at the outlet of the crankpin oil bore,

[Nm] 1 M BON = ± M BOmax − M BOmin ,

where: ψ – angular position of oil bore (see Fig

- [mm3] We = 2
- 4 ⎤ π ⎡ D'4 − DBH

Calculation of Alternating Bending Stresses in Fillets

as well as for the journal fillet

For the crankpin fillet,

alternating bending stresses shall be determined in accordance with the formula:

αB – stress concentration factor for bending in the crankpin fillet (see formula 3

alternating bending stresses shall be determined from the below formula (not applicable to semi-built crankshafts):

σBG = ± (βB ·σBN + βQ ·σQN),

βQ – stress concentration factor for compression due to radial force in journal B

- fillet (see formula 3

Calculation of Alternating Bending Stresses in Outlet of Crankpin Oil Bore

4-1) 13

where: γB – stress concentration factor for bending in crankpin oil bore (see formula 3

Calculation of Alternating Torsional Stresses General

The calculation of nominal alternating torsional stresses shall be carried out by the engine manufacturer in accordance with paragraph 2

The manufacturer shall specify the maximum nominal alternating torsional stress

The maximum and minimum torques shall be determined for every mass point of the complete dynamic system and for the entire speed range by means of a harmonic synthesis of the forced vibrations from the 1st order up to and including the 15th order for 2-stroke cycle engines and from the 0

- 5th order up to and including the 12th order for 4-stroke cycle engines

allowance shall be made for the damping that exists in the system and for unfavourable conditions (misfiring in one of the cylinders)

The speed step calculation shall be so selected that any resonance found in the operational speed range of the engine shall be detected

they shall be so arranged that satisfactory operation is possible despite their existence

- 8 for normal firing conditions

The nominal alternating torsional stress in every mass point,

which is essential to the assessment,

results from the below formula: M τ N = ± TN ⋅ 103 ,

- 2-1) Wp M TN = ±
- 1 M Tmax − M Tmin ,

16 ⎜⎝

16 ⎜⎝

where: MT – nominal alternating torgue in the crankpin or journal,

Wp – polar section modulus related to cross-section of axially bored crankpin or bored journal,

MTmin – maximum and minimum values of the torque,

the nominal alternating torsional stress considered in further calculations is the highest value,

calculated according to the above method,

occurring at the most torsionally loaded mass point of the crankshaft system

the torsional stresses within these ranges shall not be considered for assessment calculations

The approval of crankshaft will be based on the installation having the largest nominal alternating torsional stress (but not exceeding the maximum figure specified by engine manufacturer)

For each installation,

it shall be ensured by suitable calculation that the approved nominal alternating torsional stress is not exceeded

This calculation shall be submitted to PRS for assessment

Calculation of Alternating Torsional Stresses in Filletsand Outlet of Crankpin Oil Bore

The calculation of alternating torsional stresses for the crankpin fillet shall be performed in accordance with the formula:

- 3-1) where: αT – stress concentration factor for torsion in crankpin fillet (see formula 3

The calculation of alternating torsional stresses for the journal fillet shall be performed in accordance with the formula:

- 3-2) where: βT – stress concentration factor for torsion in journal fillet (see formula 3

The calculation of alternating torsional stresses for the crankpin oil bore shall be performed in accordance with the formula:

σ TO = ±(γ T ⋅ τ N ) ,

where: γ T – stress concentration factor for torsion in outlet of crankpin oil bore (see formula 3

The stress concentration factors shall be evaluated in accordance with the formulae given in paragraphs 3

3 and 3

The stress concentration factor formulae concerning the oil bore are only applicable to a radially drilled oil hole

Where the geometry of the crankshaft is outside the boundaries of the analytical stress concentration factors (SCF),

the calculation method detailed in Appendix III may be used

βB) is defined as the ratio of the maximum equivalent stress (Von Mises) – occurring in the fillets under bending load – to the nominal bending stress related to the web cross-section (see Appendix I)

βT) is defined as the ratio of the maximum equivalent shear stress – occurring in the fillets under torsional load – to the nominal torsional stress related to the axially bored crankpin or journal crosssection (see Appendix I)

The stress concentration factor for compression (βQ) in the journal fillet is defined as the ratio of the maximum equivalent stress (Von Mises) – occurring in the fillet due to the radial force – to the nominal compressive stress related to the web cross-section

The stress concentration factors for bending (γB) and torsion (γT) are defined as the ratio of the maximum principal stress – occurring at the outlet of the crankpin oil-hole under bending and torsional loads – to the corresponding nominal stress related to the axially bored crankpin cross-section (see Appendix II)

When reliable measurements which can allow direct assessment of stress concentration factors are not available,

these values may be determined from formulae given in paragraphs 3

3 and 3

- 4 applicable to the journal and crankpin fillets of solidforged crankshafts and to the crankpin fillets of semi-built crankshafts

All crank dimensions necessary for the calculation of stress concentration factors are shown in Fig

- 1 Crank dimensions necessary for the calculation of stress concentration factors

D DBH Do RH TH DG DBG RG 16

– – – – – – – –

- crankpin diameter,

diameter of axial bore in crankpin,

diameter of oil bore in crankpin,

- fillet radius of crankpin,
- recess of crankpin fillet,
- journal diameter,

diameter of axial bore in journal,

- fillet radius of journal,
- – recess of journal fillet,
- – pin eccentricity,
- – pin overlap,

B(*) – web width,

(*) In the case of 2-stroke semi-built crankshafts: – when TH > RH,

the web thickness shall be considered as equal to: Wred = W – (TH – RH),

- [mm] (see Fig

– web width B shall be taken in accordance with Fig

- the related dimensions,
- specified in Table 3
- shall be applied

- 1 Crankpin fillet

- r = RH/D

r = RG/D s'w w b do dG dH tH tG

= S/D = W/D crankshafts with overlap = Wred/D crankshafts without overlap = B/D =Do/D =DBH/D =DBG/D =TH/D =TG/D

The stress concentration factors are valid for the ranges of related dimensions for which the investigations have been carried out

- 2 ≤ w ≤ 0
- 1 ≤ b ≤ 2
- 03 ≤ r ≤ 0
- 13 0 ≤ dG ≤ 0
- 8 0 ≤ dH ≤ 0
- 8 0 ≤ do ≤ 0

Low range of s'may be extended down to large negative values if : – the calculated f (recess) < 1,

then factor f (recess) shall not be taken into account (f (recess) = 1),

- – s'< – 0
- then f (s,
- w) and f (r,

s) shall be evaluated replacing the actual value of s'by – 0

Crankpin Fillet

shall be determined from the formula: B

- 6914 ⋅ f (s,w) ⋅ f (w) ⋅ f (b) ⋅ f (r) ⋅ f (dG) ⋅ f (dH) ⋅ f (recess) B
- f (s,w) = – 4

1883 + 29

- 2004 ⋅ w – 77
- 5925 ⋅ w2 + 91
- 9454 ⋅ w3 – 40
- 0416 ⋅ w4 + +(1–s) ⋅ (9
- 5440 – 58
- 3480 ⋅ w + 159
- 3415 ⋅ w2 – 192
- 5846 ⋅ w3 + 85
- 2916 ⋅ ⋅ w4) + (1 – s)2 ⋅ (– 3

8399 + 25

- 0444 ⋅ w – 70
- 5571 ⋅ w2 + 87
- 0328 ⋅ w3 – – 39
- 1832 ⋅ w4) f (w) = 2
- 1790 ⋅ w0,7171 f (b) = 0
- 6840 – 0,0077 ⋅ b + 0
- 1473 ⋅ b2 f (r) = 0
- 2081 ⋅ r–0,5231 f (dG) = 0

9993 + 0

- 27 ⋅ dG – 1
- 0211⋅ d'G2 + 0
- 5306 ⋅ d'G3 f (dH) = 0

9978 + 0

- 3145 ⋅ dH – 1
- 5241 · d'H2 + 2
- 4147 · d'H3 f(recess) = 1 + (tH + tG) ⋅ (1
- 2 ⋅ s) The stress concentration factor for torsion,

shall be determined from the formula:

αT = 0

- 8 ⋅ f (r,
- s) ⋅ f (b) ⋅ f (w)
- where: f (r,s) = r[–0

322 + 0

- 1015 · (1 – s)] f (b) = 7
- 8955 – 10
- 654 ⋅ b + 5
- 3482 ⋅ b2 – 0
- 857 ⋅ b3 f (w) = w–0

Journal Fillet (not applicable to semi-built crankshafts)

shall be determined from the formula: βB = 2

- 7146 ⋅ fB (s,w) ⋅ fB (w) ⋅ fB (b) ⋅ fb (r) ⋅ fB (dG) fB (dH) ⋅ f (recess) (3
- 3-1) where: fB (s,w) = –1

7625 + 2

- 9821 · w – 1
- 5276 · w2 + (1 – s) ⋅ (5

1169 – 5

- 8089 ⋅ w + + 3
- 1391 ⋅ w2) + (1 – s)2 ⋅ (–2

1567 + 2

- 3297 ⋅ w – 1
- 2952 ⋅ w2) fB (w) = 2
- 2422 ⋅ w0
- 7548 fB (b) = 0

5616 + 0

- 1197 · b + 0
- 1176 · b2 fB (r) = 0
- 1908 ⋅ r(–0

5568) B

fB (dG) = 1

0012 – 0

- 6441⋅ dG + 1
- 2265 ⋅ d'2G B

0022 – 0

- 1903 ⋅ dH + 0
- 0073 ⋅ d'2H f (recess)= 1 + (tH + tG) ⋅ (1

2 ⋅ s) B

The stress concentration factor for compression due to the radial force,

shall be determined from the formula:

βQ = 3

- 0128 ⋅ fQ (s) ⋅ fQ (w) ⋅ fQ (b) ⋅ fQ (r) ⋅ fQ (dH) ⋅ f (recess)
- where: fQ (s) = 0

4368 + 2

- 1630 (1 – s) – 1
- 5212 (1– s)2 w fQ (w) = 0

0637 + 0

- 9369 ⋅ w fQ (b) = –0
- 5 + b fQ (r) = 0
- 5331 ⋅ r (–0,2038) fQ (dH) = 0

9937 – 1

- 1949 ⋅ dH + 1
- 7373 ⋅ d'2H f (recess) = 1 + (tH + tG) ⋅ (1
- 2 · s) The stress concentration factor for torsion,

shall be determined from the formula: βT = αT (3

- 3-3) if the diameters and fillet radii of crankpin and journal are the same,
- or βT = 0
- 8 ⋅ f (r,s) ⋅ f (b) ⋅ f (w) if crankpin and journal diameters and/or radii are of different sizes,
- where: f (r,

f (b) and f (w) shall be determined in accordance with paragraph 3

- 2 (see calculation of αT),
- however,

the radius of the journal fillet shall be related to the journal diameter: R (3

- 3-4) r= G DG 3

Outlet of Crankpin Oil Bore

shall be determined from the formula: γB = 3 – 5

- 88 · do+34
- 6 · do2 (3

The stress concentration factor for torsion,

shall be determined from the formula: γT = 4 – 6 · do+30 · do2 (3

further bending stresses due to misalignment and bedplate deformation,

as well as due to axial and bending vibrations shall be increased by applying σadd specified in Table 4

The additional stress of ±30 [MPa] involves two components: 1) an additional stress of ±20 [MPa] resulting from axial vibrations,

- 2) an additional stress of ±10 [MPa] resulting from misalignment / bedplate deformation

It is recommended that a value of ± 20 [MPa] be used for the axial vibration component for assessment purposes where axial vibration calculation results of the complete dynamic system (engine/shafting/ gearing/propeller) are not available

the calculated figures may be used instead

as well as for the journal fillet by using the Von Mises criterion

bending and torsion lead to two different biaxial stress fields which can be represented by a Von Mises equivalent stress with the additional assumptions that bending and torsion stresses are time phased and the corresponding peak values occur at the same location (see Appendix I)

bending and torsion lead to two different stress fields which can be represented by an equivalent principal stress equal to the maximum of principal stress resulting from combination of these two stress fields with the assumption that bending and torsion are time phased (see Appendix II)

The above two different ways of equivalent stress evaluation both lead to stresses which may be compared to the same fatigue strength value of crankshaft assessed according to the Von Mises criterion

The equivalent alternating stress is calculated in accordance with the following formulae: 20

- – for the crankpin fillet:

- – for the journal fillet:

– for the outlet of crankpin oil bore: 2⎤ ⎡ 1 9 ⎛ σ TO ⎞ ⎥ ⎢ ⎟ ,

[MPa] σ v = ± ⋅ σ BO ⋅ 1 + 2 ⋅ 1 + ⋅ ⎜⎜ ⎢ 3 4 ⎝ σ BO ⎟⎠ ⎥ ⎦ ⎣

CALCULATION OF FATIGUE STRENGTH

The fatigue strength is understood as that value of equivalent alternating stress (Von Mises) which a crankshaft can permanently withstand at the most highly stressed points

the fatigue strength may be evaluated by means of the following formulae: – related to the crankpin diameter: ⎡

σ DW = ± K ⋅ (0

- 42 ⋅ σ B + 39
- 3) ⋅ ⎢0

264 + 1

- 073 ⋅ D'−0,

2 + ⎣⎢

- 785 − σ B 196 1 ⎤ + ⋅ ⎥ 4900 σB R X ⎥⎦
- with: RX = RH
- - in the fillet area RX = Do/2
- - in the oil bore area – related to the journal diameter: ⎡

- 42 ⋅ σ B + 39

3) ⎢0

264 + 1

- 073 ⋅ DG −0,

2 + ⎣⎢

- 785 − σ B 196 1 ⎤ + ⋅ ⎥ (6-2) σB RG ⎥⎦ 4900

σDW – allowable fatigue strength of crankshaft for bending,

- [ – ]

Values greater than 1 are only applicable to fatigue strength in fillet area

- 05 for continuous grain flow forged or drop-forged crankshafts,
- 0 for free form forged crankshafts (without continuous grain flow),

K – factor for cast steel crankshafts with cold rolling treatment in fillet area: = 0

- 93 for cast steel crankshafts manufactured by companies using a PRS approved cold rolling process

σB – minimum tensile strength of crankshaft material,

For other parameters – see paragraph 3

- it must be approved by PRS

the outlet of the oil bore and inside the oil bore (down to a minimum depth equal to 1

- 5 times the oil bore diameter) shall be smoothly finished
- – for calculation purposes,

RG or RX shall be taken not less than 2 mm

the fatigue strength of the crankshaft can be determined by experiment based either on full size crank throw (or crankshaft) or on specimens taken from a full size crank throw

In any case the experimental procedure for fatigue evaluation of specimens and fatigue strength of crankshaft assessment have to be submitted to PRS for approval (method,

- type of specimens,

number of specimens (or crank throws),

- number of tests,
- survival probability,

All crank dimensions necessary for the calculation of the shrink-fit are shown in Fig

Crank throw of semi-built crankshaft

DS – shrink diameter,

between the centre line of journals and the outer contour of web,

- whichever is the lesser,

– distance between the adjacent generating lines of journal and pin y ≥ 0

05 ⋅ Ds,

special consideration shall be given to the effect of the stress due to the shrink-fit on the fatigue strength at the crankpin fillet

1 and Fig

The radius of the transition from the journal to the shrink diameter shall comply with the following condition: RG ≥ 0

- 015 DG and RG ≥ 0
- 5 (DS – DG),

where the greater value shall be considered

The actual oversize Z of the shrink-fit shall be within the limits Zmin and Zmax calculated in accordance with paragraphs 7

3 and 7

Maximum Permissible Hole in the Journal Pin

The maximum permissible hole diameter in the journal pin is calculated in accordance with the following formula:

- 4000 ⋅ S R ⋅ M max ,

[mm] μ ⋅ π ⋅ DS2 ⋅ LS ⋅ σ SP

where: SR – safety factor against slipping,

- however,

a value not less than 2 shall be taken unless documented by experiments,

μ – coefficient for static friction,

- however,
- a value not greater than 0
- 2 shall be taken unless documented by experiments,

This condition serves to avoid plasticity in the hole of journal pin

σ SW ⋅ DS Em

- 1 − Q A ⋅ QS 4000 S R ⋅ M max ⋅ ⋅ ≥ ,

[mm] μ ⋅ π Em ⋅ DS ⋅ LS (1 − Q A 2 ) ⋅ (1 − QS 2 ) QA = DS / DA

where: Zmin – minimum oversize,

Em – Young’s modulus for crank web material,

- [ – ],

Qs – shaft ratio,

- [ – ]

The maximum permissible oversize for a shrink-fit loaded with the maximum torque Zmax in the crankshaft shall be calculated from the formula:

⎛σ 0,8 ⎞ ⎟⎟ ,

[mm] Z max ≤ DS ⋅ ⎜⎜ SW + ⎝ Em 1000 ⎠

Adequate dimensioning of crankshaft is evaluated on the basis of acceptability factor for the crankpin fillet and journal fillet,

determined from the formula: Q=

σDW – fatigue strength of the crankshaft material for the crankpin fillet and journal fillet,

determined in accordance with paragraph 6,

determined in accordance with paragraph 5,

Adequate dimensioning of the crankshaft,

- except shrink-fit dimensions,

is ensured if the smallest of all acceptability factors satisfies the criterion: Q ≥ 1

Max σ1

- || σ3|| > σ1

σ1 > ||σ3||

Typical principal stress system Mohr‘s circle diagram with σ2 = 0

- 5*(σ1 – σ3) SCF = τequiv /τn for αT,

σequiv = σ12 + σ22

- - σ1· σ2 SCF = σequiv / σn for αB ,

APPENDIX II Stress Concentration Factors and Stress Distribution at the Edge of Oil Drillings

Tension + shear

Tension

Uniaxial stress distribution around the edge

Mohr‘s circle diagram

σα = σn ·γB /3 [1+2cos(2α)]

γB = σmax / σn for α= k·π

- 0 τn τn 0

APPENDIX III Alternative Method for Calculation of Stress Concentration Factors in the Web Fillet Radii of Crankshafts by Utilizing Finite Element Method CONTENTS 1 General

- 2 Model Requirements
- 1 Element Mesh Recommendations

2 Material

- 3 Element Mesh Quality Criteria
- 1 Principal stresses criterion
- 2 Averaged/unaveraged stresses criterion
- 29 29 30 30 30 30
- 3 Load Cases

1 Torsion

- 2 Pure Bending (4-Point Bending)
- 3 Bending with Shear Force (3-Point Bending)

1 Method 1

2 Method 2

- 31 31 32 33 34 34

The objective of the analysis is to develop Finite Element Method (FEM) calculated figures as an alternative to the analytically calculated Stress Concentration Factors (SCF) at the crankshaft fillets

The analytical method is based on empirical formulae developed from strain gauge measurements of various crank geometries and accordingly the application of these formulae is limited to those geometries

The SCFs calculated according to the rules of this document are defined as the ratio of stresses calculated by FEM to nominal stresses in both journal and pin fillets

When used in connection with the method presented in Publication No

- 8/P or the alternative methods,

von Mises stresses shall be calculated for bending and principal stresses for torsion

The procedure,

as well as evaluation guidelines are valid for both solid cranks and semi-built cranks (except journal fillets)

and unit loads of appropriate magnitude shall be applied for all load cases

It is advised to check the element accuracy of the FE solver in use,

by modelling a simple geometry and comparing the stresses obtained by FEM with the analytical solution for pure bending and torsion

Model Requirements

- 1 Element Mesh Recommendations In order to fulfil the mesh quality criteria it is advised to construct the FE model for the evaluation of Stress Concentration Factors according to the following recommendations: – The model consists of one complete crank,

from the main bearing centerline to the opposite side main bearing centerline – Element types used in the vicinity of the fillets: – 10 node tetrahedral elements – 8 node hexahedral elements – 20 node hexahedral elements – Mesh properties in fillet radii

The following applies to ±90 degrees in circumferential direction from the crank plane: – Maximum element size a=r/4 through the entire fillet,

as well as in the circumferential direction

the element size in the circumferential direction may be extended up to 5a

In the case of multiradii,

fillet r is the local fillet radius

(If 8 node hexahedral elements are used,

even smaller element size is required to meet the quality criteria

) – Recommended manner for element size in fillet depth direction – First layer thickness equal to element size of a – Second layer thickness equal to element to size of 2a – Third layer thickness equal to element to size of 3a – Minimum 6 elements across web thickness

- – Generally,

the rest of the crank should be suitable for numeric stability of the solver

– Counterweights have to be modelled only when influencing the global stiffness of the crank significantly

– Modelling of oil drillings is not necessary as long as the influence on global stiffness is negligible and the proximity to the fillet is more than 2r,

- see Fig

– Drillings and holes for weight reduction have to be modelled

– Sub-modelling may be used as far as the software requirements are fulfilled

Oil bore proximity to fillet

Material

- 8/P does not consider material properties such as Young’s Modulus (E) and Poisson’s ratio (ν )

In FE analysis,

those material parameters are required as strain is primarily calculated and stress is derived from strain using the Young’s Modulus and Poisson’s ratio

either as quoted in literature or as measured on representative material samples

For steel,

the following is advised: E = 2

- 05·105 MPa and ν =0

If the actual element mesh does not fulfil any of the following criteria at the examined area for SCF evaluation,

then a second calculation with a refined mesh shall be performed

- 1 Principal stresses criterion The quality of the mesh should be assured by checking the stress component normal to the surface of the fillet radius

Ideally,

- this stress should be zero

σ2 and σ3 ,

the following criterion is required:

- min (|σ1 |,

- 03·max (|σ1 |,

|σ2 |,

- 2 Averaged/unaveraged stresses criterion The criterion is based on observing the discontinuity of stress results over elements at the fillet for the calculation of SCF: Unaveraged nodal stress results calculated from each element connected to a node,

should differ less than by 5 % from the 100 % averaged nodal stress results at this node at the examined location

Load Cases

To substitute the analytically determined SCF in accordance with Publication No

the following load cases have to be calculated

The examined structure is loaded in pure torsion

surface warp at the end faces is suppressed

This node acts as the master node with 6 degrees of freedom and is connected rigidly to all nodes of the end face

- 1 Boundary and load conditions for the torsion load case

principal stresses are extracted and the equivalent torsional stress is calculated:

- σ2|,

The maximum value taken for the subsequent calculation of the SCF:

αT = τequiv,α / τN βT = τequiv,β / τN where: τN is nominal torsional stress referred to the crankpin and journal,

- respectively,
- as specified in paragraph 2
- 2 of Publication No
- with the torsional torque T:

The examined structure is loaded in pure bending

In the model,

surface warp at the end faces is suppressed

The bending moment is applied to the central node located at the crankshaft axis

This node acts as the master node with 6 degrees of freedom and is connected rigidly to all nodes of the end face

- 2 Boundary and load conditions for the pure bending load case

von Mises equivalent stresses σequiv are extracted

The maximum value is used to calculate the SCF according to:

αB = σ equiv,α / σ N βB = σ equiv,β / σ N Nominal stress σ N is calculated as specified in paragraph 2

- 1 of Publication No
- 8/P with the bending moment M:

The structure is loaded in 3-point bending

In the model,

surface warp at the both end faces is suppressed

boundary conditions are applied to the centre nodes

These nodes act as master nodes with 6 degrees of freedom

- 4 Load applications for in-line and V-type engines

The maximum equivalent von Mises stress σ3P in the journal fillet is evaluated

The results from 3-point and 4-point bending are combined as follows:

- - nominal bending stress in the web centre due to the force F3P [N] applied to the centre-line of the actual connecting rod,
- see Fig

σQ3P = Q3P / (B·W) where Q3P is the radial (shear) force in the web due to the force F3P [N] applied to the centre-line of the actual connecting rod,

- see also Fig
- 1 in Publication No

Method 2

In a statically determined system with one crank throw supported by two bearings,

the bending moment and radial (shear) force are proportional

the journal fillet SCF can be found directly by the 3-point bending FE calculation

The SCF is then calculated according to:

- see paragraph 3

the radial force and stress determination in accordance with Publication No

- 8/P becomes superfluous

- as specified in paragraph 2
- 3 of Publication No
- is then evaluated:

σBG = ± βBQ · σBFN Note that the use of this method does not apply to the crankpin fillet and that this SCF must not be used in connection with calculation methods other than those assuming a statically determined system specified in Publication No