Section
3 Crankshaft Design
3.1 Application
3.2 Scope
3.2.1 The formulae given in this Section are applicable to solid or semi-built crankshafts
of forged or cast steel, having a main support bearing adjacent to each crankpin.
3.2.2 This section uses the statically determinate method; alternative
methods, including a fully documented stress analysis, will be considered.
3.2.3 Calculations are to be carried out for the maximum continuous power
rating for all designed operating conditions.
3.2.4 Designs of crankshafts not included in this scope will be subject to special
consideration.
3.2.5 Where a crankshaft design involves the use of surface treated fillets, or when
fatigue parameter influences are tested, or when working stresses are measured, the
relevant documents with calculations/analysis are to be submitted to LR.
3.2.6 The design of crankshafts 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; and
- outlets of crankpin oil bores.
3.2.7 When the journal diameter is equal to or larger than the crankpin
diameter, the outlets of main journal oil bores are to be formed in a similar way to
the crankpin oil bores, otherwise separate documentation of fatigue safety will be
specially considered.
3.2.8 Calculation of crankshaft strength consists initially in determining the
nominal alternating bending (see
Pt 5, Ch 2, 3.6 Calculation of bending stresses) and nominal alternating
torsional stresses (see
Pt 5, Ch 2, 3.7 Calculation of torsional stresses) which when multiplied
by the appropriate stress concentration factors (SCF) (see
Pt 5, Ch 2, 3.8 Stress concentration factors), result in an equivalent
alternating stress (uniaxial stress) (see Pt 5, Ch 2, 3.10 Equivalent
alternating stress). This equivalent alternating stress is then compared with the
fatigue strength of the selected crankshaft material (see
Pt 5, Ch 2, 3.11 Fatigue strength). This comparison will show whether or
not the crankshaft concerned is dimensioned adequately (see
Pt 5, Ch 2, 3.12 Acceptability criteria).
3.2.9 Further information and guidance on crankshaft design is provided in the
LR Guidance Notes for the Calculation of Stress Concentration Factors, Fatigue
Enhancement Methods and Evaluation of Fatigue Tests for Crankshafts.
3.3 Information to be submitted
3.3.1 For the calculation of crankshafts, the documents and particulars listed
below are required. This information is provided by completing LR Form 2073 and
submitting the applicable plans required in Table 2.1.1 Plans and particulars to be
submitted:
- crankshaft drawing (which must contain all data in respect of the
geometrical configurations of the crankshaft);
- type designation and kind of engine (in-line engine or V-type
engine with adjacent connecting rods, forked connecting rod or articulated-type
connecting rod);
- operating and combustion method (2-stroke or 4-stroke cycle/direct
injection, precombustion chamber, etc.);
- number of cylinders;
- output power at maximum continuous rating (MCR), in kW;
- output speed at maximum continuous power, in rpm;
- maximum firing pressure, Pmax, in MPa;
- mean indicated pressure, in MPa;
- charge air pressure (before inlet valves or scavenge ports,
whichever applies), in MPa;
- digitised gas pressure/crank angle cycle for MCR (presented at
equidistant intervals at least every 5° CA);
- mean piston speed;
- compression ratio;
- vee angle αv, in degrees;
- firing order numbered from driving end, see
Figure 2.3.1 Designation of cylinders;
- direction of rotation;
- cylinder diameter, in mm;
- piston stroke, in mm;
- centre of gravity of connecting rod from large end centre, in
mm;
- radius of gyration of connecting rod, in mm;
- length of connecting rod between bearing centres,
LH, in mm;
- mass of single crankweb (indicate if webs either side of pin are of
different mass values), in kg;
- centre of gravity of crankweb mass from shaft axis, in mm;
- mass of counterweights fitted (for complete crankshaft) indicate
positions fitted, in kg;
- centre of gravity of counterweights (for complete crankshaft)
measured from shaft axis, in mm;
- all individual reciprocating masses acting on one crank, in kg;
- crankshaft material specification(s) (according to ISO, EN, DIN,
AISI, etc.);
- mechanical properties of material (minimum values obtained from
longitudinal test specimens):
- tensile strength, in N/mm2
- yield strength, in N/mm2
- reduction in area at break, percentage
- elongation, percentage
- method of manufacture (free form forged, continuous grain flow
forged, drop-forged, etc.; with description of the forging process);
- for semi-built crankshafts – minimum and maximum diametral
interference, in mm; and
- particulars of alternating torsional stress calculations, see
Pt 5, Ch 2, 3.7 Calculation of torsional stresses.
Figure 2.3.1 Designation of cylinders
Figure 2.3.2 Articulated-type connecting rod
3.3.2 The following information is also required for appraisal of the crankshaft (not
contained in Form 2073):
- for engines with articulated-type connecting rod (see
Figure 2.3.2 Articulated-type connecting rod);
- distance to link point LA, in mm
- link angle αN, in degrees
- connecting rod length LN, in mm
- firing interval (if applicable) i.e. if not evenly distributed
- mass of connecting rod (including bearings), in kg;
- mass of piston (including piston rod and crosshead where
applicable), in kg;
- every surface treatment affecting fillets or oil holes shall be
specified so as to enable calculation according to Chapter 3 of the LR
Guidance Notes for the Calculation of Stress Concentration Factors,
Fatigue Enhancement Methods and Evaluation of Fatigue Tests for
Crankshafts;
- this is to include crankshaft fatigue enhancement factors
K1 and K2 where applicable.
- maximum alternating torsional stress τa
(N/mm2)
- mechanical properties of material (minimum values obtained from
longitudinal test specimens), in addition to the information listed above:
- Impact energy KV, in Joules.
3.4 Symbols
3.4.1 For the purposes of this Chapter the following symbols apply, see
also;
B |
= |
transverse breadth of web, in mm |
D |
= |
crankpin diameter, in mm |
DA |
= |
the outside diameter of web or twice the minimum distance
between centreline of journals and outer contour of web, whichever is
less, in mm |
DBH |
= |
diameter of axial bore in crankpin, in mm |
DBG |
= |
diameter of axial bore in journal, in mm |
DG |
= |
journal diameter, in mm |
Do |
= |
diameter of radial oil bore in crankpin, in mm |
Ds |
= |
shrink diameter of main journal in web, in mm |
E |
= |
pin eccentricity |
Em |
= |
Young’s modulus of crankshaft material, in
N/mm2 |
F |
= |
area related to cross-section of web, in
mm2 |
Ke
|
= |
bending stress factor (considers the influence of adjacent
crank and bearing restraint) |
K |
= |
fatigue enhancement factor (K =
K1.K2) |
K1 |
= |
fatigue enhancement factor due to manufacturing
process |
K2 |
= |
fatigue enhancement factor due to surface treatment |
Ls |
= |
length of shrink fit, in mm |
MBON |
= |
alternating bending moment calculated at the outlet of
crankpin oil bore |
MBRFN |
= |
alternating bending moment related to the centre of the web,
in Nm |
MTN |
= |
maximum alternating torque, in Nm |
MTmax |
= |
maximum value of the torque, in Nm |
MTmin |
= |
minimum value of the torque, in Nm |
QRFN |
= |
alternating radial force related to the web, in N |
RH, RG |
= |
fillet radius at junction of web and pin or journal, in
mm |
S |
= |
pin overlap, in mm ![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e897.png) |
TH, TG |
= |
recess of pin or journal fillet radius into web measured
from web face, in mm |
W |
= |
axial thickness of web, in mm |
Weqw |
= |
section modulus related to cross-section of web, in
mm3 |
We |
= |
section modulus related to cross-section of axially bored
crankpin, in mm3 |
Wp |
= |
polar section modulus related to cross-section of axially
bored crankpin or bored journal, in mm3 |
y |
= |
distance between the adjacent generating lines of journal
and pin, in mm
Note For
y≥ 0,05DS. Where y is less than
0,1DS, special consideration is to be given to
the effect of the stress due to the shrink fit on the fatigue
strength at the crankpin fillet.
|
αB |
= |
bending stress concentration factor for crankpin
fillet |
αT |
= |
torsional stress concentration factor for crankpin
fillet |
βB |
= |
bending stress concentration factor for main journal
fillet |
βQ |
= |
compression stress concentration factor for main journal
fillet
Note βQ 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.
|
βT |
= |
torsional stress concentration factor for main journal fillet |
γB |
= |
bending stress concentration factor for outlet of crankpin oil bore |
γT |
= |
torsional stress concentration factor for outlet of crankpin oil
bore |
σadd |
= |
additional bending stress due to misalignment and bedplate
deformation as well as due to axial and bending vibrations, in
N/mm2 |
σB |
= |
specified minimum UTS of crankshaft material, in N/mm2 |
σBFN |
= |
nominal alternating bending stress related to the web, in
N/mm2 |
σBG |
= |
alternating bending stress in journal fillet, in N/mm2 |
σBH |
= |
alternating bending stress in crankpin fillet, in N/mm2 |
σBO |
= |
alternating bending stress in the outlet of the oil bore, in
N/mm2 |
σBON |
= |
nominal alternating bending stress in the outlet of the oil bore related to
the crankpin diameter, in N/mm2 |
σQFN |
= |
nominal alternating compressive stress due to radial force related to the
web, in N/mm2 |
σDW |
= |
allowable fatigue strength of crankshaft, in N/mm2 |
σSP |
= |
minimum yield strength of material for journal pin, in N/mm2 |
σSW |
= |
minimum yield strength of material for crankweb, in N/mm2 |
σTO |
= |
alternating torsional stress in the outlet of the crankpin oil bore, in
N/mm2 |
σy |
= |
equivalent alternating stress for crankpin fillet, journal
fillet or outlet of crankpin oil bore as applicable, in
N/mm2 |
τH |
= |
alternating torsional stress in crankpin fillet, in N/mm2 |
τG |
= |
alternating torsional stress in journal fillet, in N/mm2 |
τN |
= |
calculated nominal alternating torsional stress referred to crankpin or
journal (as applicable), in N/mm2 |
τa |
= |
manufacturer stated crankshaft half range torsional stress limit, in
N/mm2 |
Figure 2.3.3 Crank dimensions for overlapped crankshaft
Figure 2.3.4 Crank dimensions for crankshaft without overlap
Figure 2.3.5 Crankpin section through the oil bore
Figure 2.3.6 Crankthrow of semi-built crankshaft
Figure 2.3.7 Stress concentration factors in crankshaft fillets
Figure 2.3.8 Stress concentration factors and stress distribution at the edges of oil
drillings
3.5 Calculation of alternating stresses due to bending moments and radial forces – assumptions
3.5.2 The bending moments, MBR and MBT, 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
Figure 2.3.9 Bending moment and shear force
for in-line engine crankthrows). For crankthrows with two connecting-rods acting upon
one crankpin the relevant bending moments are obtained by superposition of the two
triangular bending moment diagrams according to phase (see
Figure 2.3.10 Bending moment and shear force
for V engine crankthrows).
L1 |
= |
Distance between main journal centreline and crankweb centre |
L2 |
= |
Distance between main journal centreline and connecting - rod
centre |
L3 |
= |
Distance between two adjacent main journal centrelines |
Figure 2.3.9 Bending moment and shear force
for in-line engine crankthrows
L1 |
= |
Distance between main journal centreline and crankweb centre |
L2 |
= |
Distance between main journal centreline and connecting - rod
centre |
L3 |
= |
Distance between two adjacent main journal centrelines |
Figure 2.3.10 Bending moment and shear force
for V engine crankthrows
3.5.4 The two relevant bending moments for bending acting on the outlet of
crankpin oil bores are taken in the crankpin cross-section through the oil bore.
See;
MBRO is the bending moment of the radial component of the
connecting-rod force and MBTO is the bending moment of the
tangential component of the connecting-rod force. The alternating stresses due to
these bending moments are to be related to the cross-sectional area of the axially
bored crankpin. Mean bending stresses are neglected.
3.6 Calculation of bending stresses
3.6.1 The radial and tangential forces due to gas and inertia loads acting upon
the crankpin at each connecting-rod position are to be calculated over one working
cycle. Using the forces calculated over one working cycle and taking into account of
the distance from the main bearing midpoint, the time curve of the bending moments,
MBRF, MBRO and MBTO, and
radial forces, QRF, as defined in Pt 5, Ch 2, 3.5 Calculation of alternating stresses due to bending moments and radial forces – assumptions 3.5.3
and Pt 5, Ch 2, 3.5 Calculation of alternating stresses due to bending moments and radial forces – assumptions 3.5.4
are then calculated.
3.6.2 Nominal bending stresses are referred to the web bending modulus.
3.6.3 In case of V-type engines, the bending moments – progressively calculated from the
gas and inertia forces – of the two cylinders acting on one crankthrow are
superposed according to phase. Different designs (forked connecting-rod,
articulated-type connecting-rod or adjacent connecting-rods) shall be taken into
account.
3.6.4 Where there are cranks of different geometrical configurations in one crankshaft, the
calculation is to cover all crank variants.
3.6.5 The decisive alternating values will then be calculated according to:
XN |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e1896.png) |
where
XN is considered as the alternating force, moment or
stress
Xmax is the maximum value within one working cycle
Xmin is the minimum value within one working cycle
3.6.7 Nominal alternating bending stress in the outlet of the crankpin oil bore is
calculated as follows:
σBON |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e2572.png) |
where
MBON is taken as the ˝ range value
MBON = ± ˝ (MBOmax –
MBOmin) Nm
and
MBO |
= |
(MBTO cosψ + MBRO sinψ),
ψ = angular position in degrees, see
Figure 2.3.5 Crankpin section through the oil bore |
MBRO |
= |
bending moment of the radial component of the connecting-rod force |
MBTO |
= |
bending moment of the tangential component of the connecting-rod
force |
We |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e2766.png) |
3.7 Calculation of torsional stresses
3.7.1 The nominal alternating torsional stress, τN, is to be taken
into consideration. The value is to be derived from forced-damped vibration
calculations of the complete dynamic system. Alternative methods will be given
consideration. The engine designer is to advise the maximum level of alternating
vibratory stress that is permitted (τa).
3.7.3 Nominal alternating torsional stress is calculated as follows:
τN |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e3365.png) |
where
MTN |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e3486.png) |
Wp |
= |
for the crankpin, or for the journal |
τN is to be ascertained from assessment of the torsional
vibration calculations where the maximum and minimum torques are 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 first 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. Whilst doing so,
allowance must be made for the damping that exists in the system and for
unfavourable conditions (misfiring in one of the cylinders when no combustion occurs
but only on the compression cycle). The speed step calculation shall be selected in
such a way that any resonance found in the operational speed range of the engine
shall be detected.
3.7.4 For the purpose of the crankshaft assessment, the nominal alternating
torsional stress considered in calculations is to be the highest calculated value,
according to the method described in Pt 5, Ch 2, 3.7 Calculation of torsional stresses 3.7.3, occurring at the most torsionally loaded mass point
of the crankshaft system.
3.7.5 The approval of the crankshaft will be based on the installation having
the largest nominal alternating torsional stress (but not exceeding the maximum
figure specified by the engine manufacturer). For each installation it is to be
ensured by calculation that the maximum approved nominal alternating torsional
stress is not exceeded. See
Pt 5, Ch 8, 2 Torsional vibration.
3.8 Stress concentration factors
3.8.1 Stress concentration factors (SCF) are to be calculated using the analytical formulae
outlined in this Section.
3.8.3 Where the geometry of the crankshaft is outside the boundaries
(see
Table 2.3.2 Crankshaft variable boundaries for analytical SCF calculation) of the analytical SCF the
calculation method detailed in Chapter 1 and Chapter 4 of the LR Guidance Notes
for Calculation of Stress Concentration Factors, Fatigue Enhancement Methods and
Evaluation of Fatigue Tests for Crankshafts may be undertaken.
3.8.4 Where reliable experimental measurements and/or calculations are
available, which can allow direct assessment of SCF, these can be used. The relevant
documents and their analysis are to be submitted for consideration in order to
demonstrate their equivalence. This is always to be performed when dimensions are
outside the boundaries shown in Table 2.3.2 Crankshaft variable boundaries for analytical SCF calculation.
3.8.5 Chapters 1 and 4 of the LR Guidance Notes for Calculation of Stress
Concentration Factors, Fatigue Enhancement Methods and Evaluation of Fatigue
Tests for Crankshafts describe how finite element (FE) analyses can be used
for the calculation of the SCF. Care needs to be taken to avoid mixing equivalent
(von Mises) stresses and principal stresses.
Table 2.3.1 Crankshaft variables for SCF calculation
Variable
|
Function
|
r
|
=
RH/D for crankpin fillet
=
RG/D for journal fillet
|
s
|
=
S/D
|
w
|
=
W/D crankshafts with overlap
=
Wred/D crankshafts without
overlap
|
b
|
=
B/D
|
do
|
=
DO/D
|
dG
|
=
DBG/D
|
dH
|
=
DBH/D
|
tH
|
=
TH/D
|
tG
|
=
TG/D
|
Table 2.3.2 Crankshaft variable boundaries for analytical SCF calculation
Lower
bound
|
Variable
|
Upper
bound
|
|
s
|
≤
0,5
|
0,2 ≤
|
w
|
≤
0,8
|
1,1
≤
|
b
|
≤
2,2
|
0,03
≤
|
r
|
≤
0,13
|
0 ≤
|
dG
|
≤
0,8
|
0 ≤
|
dH
|
≤
0,8
|
0 ≤
|
do
|
≤
0,2
|
Notes
The lower
bound of s can be extended down to large negative values
provided that:
- If calculated f(recess) < 1, then
the factor f(recess) is not to be considered
(f(recess) = 1)
- If s < –0,5 then
f(s,w) and
f(r,s) are to be evaluated
replacing actual value of s by –0,5.
|
3.8.6 Crankpin SCF are calculated as follows:
- Bending
αB |
= |
2,6914 · f(s,w) · f(w) ·
f(b) · f(r) ·
f(dG) ·
f(dH) · f(recess) |
- where
f(s,w) |
= |
–4,1883 + 29,2004w – 77,5925w˛ +
91,9454w3 – 40,0416w4 + (1 –
s)(9,5440 – 58,3480w + 159,3415w˛ –
192,5846w3 + 85,2916w4)
+ (1 – s)˛(–3,8399 + 25,0444w – 70,5571w˛ +
87,0328w3 –
39,1832w4) |
f(w) |
= |
2,1790w0,7171 |
f(b) |
= |
0,684 – 0,0077b +
0,1473b2 |
f(r) |
= |
0,2081r(–0,5231) |
f(dG) |
= |
0,9993 + 0,27dG –
1,0211dG2 +
0,5306dG3 |
f(dH) |
= |
0,9978 + 0,3145dH –
1,5241dH˛ +
2,4147dH3 |
f(recess) |
= |
1 + (tH +
tG)(1,8 + 3,2s) |
- Torsion
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– 10,654b + 5,3482b2
– 0,857b3 |
f(w) |
= |
w(–0,145) |
3.8.7 Journal fillet SCF are calculated as follows(not applicable to semi-built
crankshafts):
- Bending
βB |
= |
2,7146
·fB(s,w)·fB(w)·fB(b)·fB(r)·fB(dG)·fB(dH)·f(recess) |
- where
fB(s,w) |
= |
–1,7625 + 2,9821w – 1,5276w˛ + (1 –
s)(5,1169 – 5,8089w + 3,1391w˛) + (1 –
s)˛(–2,1567 + 2,3297w – 1,2952w˛) |
fB(w) |
= |
2,2422w0,7548 |
fB(b) |
= |
0,5616 + 0,1197b +
0,1176b2 |
fB(r) |
= |
0,1908r(–0,5568) |
fB(dG) |
= |
1,0012 – 0,6441dG +
1,2265dG2 |
fB(dH) |
= |
1,0022 – 0,1903dH +
0,0073dH2 |
f(recess) |
= |
1 + (tH +
tG)(1,8 + 3,2s)
|
- Compression due to the radial force
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 |
fQ(w) |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e5539.png) |
fQ(b) |
= |
b –0,5 |
fQ(r) |
= |
0,5331r(–0,2038)
|
fQ(dH) |
= |
0,9937 – 1,1949dH +
1,7373dH2 |
f(recess) |
= |
1 + (tH +
tG)(1,8 + 3,2s) |
- Torsion
βT |
= |
αT if the diameters and fillet radii of
crankpin and journal are the same, or |
βT |
= |
0,8f(r,s).f(b).f(w)
if crankpin and journal diameters and/or radii are of different
sizes |
where
f(r,s),
f(b) and f(w) are to be determined in
accordance with Pt 5, Ch 2, 3.8 Stress concentration factors 3.8.6.(b), however,
the radius
of the journal fillet is to be related to the journal diameter:
3.8.8 Crankpin oil bore SCF for radially drilled oil holes are calculated as follows:
- Bending
γB |
= |
3 – 5,88do +
34,6do2 |
- Torsion
3.9 Additional bending stress
3.9.1 In addition to the alternating bending stresses in fillets (see
Pt 5, Ch 2, 3.6 Calculation of bending stresses 3.6.8) further bending stresses due to
misalignment and bedplate deformation as well as due to axial and bending vibrations
are to be considered by applying σadd as given by Table 2.3.3 Additional bending stresses
Table 2.3.3 Additional bending stresses
Type of
engine
|
σadd
|
Crosshead
engines
|
± 30
N/mm2 (see Note 1)
|
Trunk piston
engines
|
± 10
N/mm2
|
Note 1. The additional stress of ±30 N/mm 2 is composed
of two components:
- an additional stress of ±20 N/mm2 resulting
from axial vibration
- an additional stress of ±10 N/mm2 resulting
from misalignment/bedplate deformation
|
3.9.2 It is recommended that a value of ±20 N/mm2 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. Where axial vibration calculation results of the complete dynamic system
are available, the calculated figures can be used instead.
3.10 Equivalent alternating stress
3.10.1 In the fillets, 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
Figure 2.3.7 Stress concentration factors in crankshaft fillets). As a result the equivalent alternating stress is to be
calculated for the crankpin fillet as well as for the journal fillet by using the
von Mises criterion.
3.10.3 The above two different ways of equivalent stress evaluation both lead to stresses
which can be compared to the same fatigue strength value of crankshaft assessed
according to the von Mises criterion.
3.10.4 Equivalent alternating stress, σ v, is defined as:
- For the crankpin fillet:
σv |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e6103.png) |
- For the journal fillet:
σv |
= |
N/mm |
- For the outlet of crankpin oil bore:
σv |
= |
![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e6400.png) |
3.11 Fatigue strength
3.11.1 The fatigue strength is to be 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 can be evaluated by means of the
following formulae:
- Related to the crankpin diameter:
σDW |
= |
N/mm2 |
with
RX |
= |
RH in the fillet area |
RX |
= |
Do/2 in the oil bore area |
- Related to the journal diameter:
σDW |
= |
N/mm2 |
- where
K |
= |
K1K2 |
K1 |
= |
fatigue endurance factor appropriate to the
manufacturing process |
= |
1,05 for continuous grain flow forged or drop-forged
crankshafts |
= |
1,0 for free form forged crankshafts (without
continuous grain flow) |
= |
0,93 for cast steel crankshafts with cold rolling
treatment in fillet area manufactured by companies using a LR
approved cold rolling process |
K2 |
= |
fatigue enhancement factor for surface treatment.
These treatments are to be applied to the fillet radii |
A value for K2 will be assigned upon
application by the engine designers. Full details of the process,
together with the results of full scale fatigue tests will be required
to be submitted for consideration. See Chapter 2 of the LR
Guidance Notes for the Calculation of Stress Concentration
Factors, Fatigue Enhancement Methods and Evaluation of Fatigue Tests
for Crankshafts. Alternatively, the following values may be
taken (surface hardened zone to include fillet radii):
K2 |
= |
1,15 for induction hardened |
= |
1,25 for nitrided |
Where a value of K1 or K2 greater
than unity is to be applied, details of the manufacturing process are to be
submitted. An enhanced K1 factor will be considered, subject to
special approval of the manufacturing specification. See Materials and
Qualification Procedures for Ships, Book E, Procedure MQPS 5-2.
3.11.2 The formulae in Pt 5, Ch 2, 3.11 Fatigue strength 3.11.1 are subject to geometry limits. The junction of the
oil hole with the crankpin or main journal surface is to be formed with an adequate
radius and smooth surface finish down to a minimum depth equal to 1,5 times the oil
bore diameter and for calculation purposes RH,
RG or RX are to be taken as not less than 2
mm.
3.11.3 Fatigue strength calculations or, alternatively, fatigue test results
determined by experiment based either on full size crankthrow (or crankshaft), or on
specimens taken from a full size crankthrow, may be required to demonstrate
acceptability. The experimental procedure for fatigue evaluation of specimens and
fatigue strength of crankshaft assessment are to be submitted for approval by LR.
The procedure is to include as a minimum: method, type of specimens, number of
specimens (or crankthrows), number of tests, survival probability, and confidence
number. See also Chapter 2 of the LR Guidance Notes for the Calculation of
Stress Concentration Factors, Fatigue Enhancement Methods and Evaluation of
Fatigue Tests for Crankshafts.
3.11.4 When journal diameter is equal or larger than the crankpin diameter, the
outlets of main journal oil bores are to be formed in a similar way to the crankpin
oil bores, otherwise separate fatigue strength calculations or, alternatively,
fatigue test results may be required.
3.11.5 Only surface treatment processes approved by LR are permitted. Guidance
for calculation of surface treated fillets and oil bore outlets is presented in
Chapter 3 of the LR Guidance Notes for the Calculation of Stress Concentration
Factors, Fatigue Enhancement Methods and Evaluation of Fatigue Tests for
Crankshafts.
3.12 Acceptability criteria
3.12.1 The sufficient dimensioning of a crankshaft is confirmed by a comparison
of the equivalent alternating stress and the fatigue strength. The acceptability
factor, Q, is to be greater than or equal to 1,15 for the crankpin fillet,
the journal fillet and the outlet of crankpin oil bore:
Q |
= |
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3.13 Shrink fit of semi-built crankshafts
3.13.2 In general, the radius of transition, RG, between the main journal
diameter, DG, and the shrink diameter, DS, is to
be not less than 0,015DG or 0,5(DS –
DG) where the greater value is to be considered.
3.13.3 Deviations from these parameters will be specially considered.
3.13.4 The maximum permissible internal diameter in the journal pin is to be calculated in
accordance with the following formula, this condition serves to avoid plasticity in
the hole of the journal pin:
DBG |
= |
mm |
where
SR |
= |
safety factor against slipping; however, a value of not less
than 2 is to be taken unless documented by experiments. |
Mmax |
= |
absolute maximum value of the torque MTmax
in accordance with Pt 5, Ch 2, 3.7 Calculation of torsional stresses 3.7.3, in Nm |
μ |
= |
coefficient for static friction; however, a value of not
greater than 0,2 is to be taken unless documented by experiments. |
3.13.6 The minimum required diametral interference is to be taken as the
greater of:
where
QA |
= |
web ratio, QA = ![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e8189.png) |
QS |
= |
shaft ratio, QS = ![](svgobject/2Fwork2Ftemp2FLRSHIP_PT5_CH2_3.xml_d12071421e8256.png) |
3.13.7 The maximum diametral interference is not to be greater
than:
This condition serves to restrict the shrinkage induced mean stress in the
fillet.
3.13.8 Reference marks are to be provided on the outer junction of the crankwebs with the
journals.
|