Redesign of a Wind Turbine Hub
MASSACHUSETTS INSTITUTE
OF TECHNOLOGY
by
JUL 3 0 2014
Bridget I. Hunter-Jones
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Submitted to the Department of Mechanical Engineering
in partial fulfillment of the requirements for the degree of
Bachelor of Science in Engineering as Recommended by the Department of
Mechanical Engineering
at the
MASSACHUSETTS INSTITUTE OF TECHNOLOGY
June 2014
( Bridget I. Hunter-Jones, MMXIV. All rights reserved.
The author hereby grants to MIT permission to reproduce and distribute
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Department of Mechanical Engineering
Jan 17, 2014
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Signature
Certified by ...............
.................
Sanjay Sarma
Professor of Mechanical Engineering
Signature redactedThesis
Supervisor
A ccepted by ...............................................................
Anette Hosoi
Professor of Mechanical Engineering
Undergraduate Officer
Redesign of a Wind Turbine Hub
by
Bridget I. Hunter-Jones
Submitted to the Department of Mechanical Engineering
on Jan 17, 2014, in partial fulfillment of the
requirements for the degree of
Bachelor of Science in Engineering as Recommended by the Department of Mechanical
Engineering
Abstract
The current designs of wind turbine hubs contain many faults. The slew ring bearing that
connects the blade to the hub takes on a large bending moment that in many cases causes
the joints to fail and the blade to break off. The design of the hub is very large, heavy, and
expensive to machine. The objective of this research is to reappraise the existing design of
wind turbine hub and to suggest an alternative design, which does not have the design faults
that are currently found.
Drawings and sketches are presented along with all necessary calculations for the current
and new wind turbine hub. The proposed redesign of the hub will have a bearing pack that
will distribute the load of the blade and a new hub which is lighter, less complex, and simple
to machine. This new design will be much safer than the current design and will produce
more reliable wind turbines.
Thesis Supervisor: Sanjay Sarma
Title: Professor of Mechanical Engineering
2
Acknowledgments
I would like to thank my supervisor, Professor Sanjay Sarma, for assisting with my project
that was carried over from my semester abroad at Oxford University. I would also like to
thank my supervisor at Oxford, Dr. Stuart Turnbull, a fellow at St. Peter's College, for his
assistance and the time that he dedicated to the project. His encouragement and enthusiasm
were inspiring. I'd also like to acknowledge the guidance and advice from my brother Nick.
In addition my parents, Lynette Jones and Ian Hunter, deserve a great deal of recognition
for always being supportive and making my life run smoothly. I don't know what I would
do without their constant encouragement.
3
Contents
Abstract
2
Acknowledgments
3
Contents
4
List of Figures
5
1
6
Introduction
Background ...........
. . . . . . . . . . . . . . . . . . . . . . .
6
1.2
Failure of Bolts Due to Fatigue
. . . . . . . . . . . . . . . . . . . . . . .
7
1.3
Current Design Flaws . . . . .
. . . . . . . . . . . . . . . . . . . . . . .
12
1.3.1
Primary Issue . . . . .
. . . . . . . . . . . . . . . . . . . . . . .
12
1.3.2
Bolt Load Calculations
. . . . . . . . . . . . . . . . . . . . . . .
13
1.3.3
Current Hub Design
. . . . . . . . . . . . . . . . . . . . . . .
13
.
.
.
.
.
.
.
.
.
1.1
2 Design Philosophy
16
3 Design
17
4
26
Conclusion
Summary
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
26
4.2
Future Work. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
26
.
.
4.1
Bibliography
27
4
List of Figures
1-1
Concrete pumping machine with slew ring bearing attaching the boom.
. .
7
1-2 Sketch of how the booms turn-table is fixed to the slew ring. . . . . . . . . .
8
1-3
. . . . . . . . . . . . . . . . . . . . .
9
1-4 Bolts that have failed under large bending moment. . . . . . . . . . . . . . .
10
1-5 View of the failed bolts. . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
10
1-6
Wind turbine in Lincolnshire, UK with one blade broken loose. . . . . . . . .
11
1-7
Goodman diagram used to calculate fatigue life of a material based on mean
Damage on the slew ring of the boom.
and alternating stresses. . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
12
1-8
Structural analogy of hub/blade attachment. . . . . . . . . . . . . . . . . . .
13
1-9
Manufacturing processes of the current hub. . . . . . . . . . . . . . . . . . .
14
1-10 Exploded view of a wind turbine showing the attachment of the blades to the
hub via a slew ring. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
3-1
15
Open spindle with the two bearings and gear, attached to the blade (Created
in SolidW orks). . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
18
Hub redesign drawing.
. . . . . . . . . . . . . . . . . . . . . . . . . . . . . .
19
3-3 Blade Data (looking at GloBladel for this report). . . . . . . . . . . . . . . .
19
3-4 Garter seal model and specifications.
. . . . . . . . . . . . . . . . . . . . . .
21
3-5 Timken Single Row Tapered Roller Bearings. . . . . . . . . . . . . . . . . . .
23
3-6 Exploded view of the new hub design .. . . . . . . . . . . . . . . . . . . . . .
25
3-2
5
Chapter 1
Introduction
1.1
Background
Variable pitch propellers have been in use since the beginning of the
2 0 th
century. In aircraft
they were introduced so that the angle of attack of the propeller blade could be altered to
change forward velocity independently of motor revolutions. They also made it possible for
the engine to be seen as a constant power device rather than the constant torque device,
which is the characteristic of most internal combustion (IC) engines. The first practical,
hydraulically operated, variable pitch propeller was designed by the English engineer Dr
Henry Hele-Shaw in 1924. After some development it was fitted to most of the fighter
planes used by the RAF in the Battle of Britain in 1940. In 1927, a Canadian engineer,
Wallace Turnbull developed an electrically operated mechanism which was taken up by the
American company Curtiss-Wright. It was popular with many pilots of multi-engine planes
during World War II as the propellers could be feathered when the engine was stopped and
not produce hydraulic pressure.
It is therefore surprising that when the first large wind turbines were first designed in the
1900's the designers did not follow the tried and tested method of attaching the blades to
the hub. The designers of variable pitch propellers had used a bearing pack which separated
the way the bearings carried the loads (i.e. axial load, bending moment and shear), but the
designers of the wind turbines decided to have all of these loads carried by a single bearing,
that bearing being a "slew ring bearing".
6
Figure 1-1: Concrete pumping machine with slew ring bearing attaching the boom. [1]
Slew rings have been in use for many decades and in many forms. The gun turrets of
warships were fitted with a form of slew ring, as were the gun turrets of tanks. All tower
cranes use slew rings to allow the tower's arm to rotate through 360 degrees. However, they
all have one thing in common: the load on the ring is predominately axial, with very little
bending or shear present. Figure 1-1 shows a concrete pumping machine that has a boom
fitted with a slew ring. The turn-table of the crane was attached to the slew ring with
several bolts as shown in Figure 1-2. The bearing eventually failed, as shown in Figure 1-3,
because of the large bending moment it had to support and resulted in the death of the
crane operator.
1.2
Failure of Bolts Due to Fatigue
The bolts that held the slew ring on the crane in Figure 1-3, underwent a large bending
moment and once one of the bolts failed the rest consecutively yielded as well. Figure 1-
7
PDOW
MW -4111,w
1u-1m1
I
I
R~L
I
_
_
-THE
FIKURE I
ARICULATEb BOOM
Figure 1-2: Sketch of how the booms turn-table is fixed to the slew ring. [2]
8
Figure 1-3: Damage on the slew ring of the boom. [1]
4 illustrates how these bolts have failed by fatigue. Cratering and pitting can be seen in
bolts after fatigue failure as shown in Figure 1-5. This could be the reason for the number
of failures that have occurred on wind turbines in the last few years. In one instance in
Lincolnshire, U.K., shown in Figure 1-6, had a 20 meter (65 foot) blade broke off a wind
turbine and flung to the ground. It was reported that the blade came loose after bolts
attaching it to the hub failed [3].
Many engineers believe installation preload is critical to preventing failures [4]. High
preloads in stiff joints transfer less of the externally applied loads to the bolt, reducing cyclic
loading [5]. However, one study done on the effect of clamping force on the fatigue behavior
of bolted plates showed that if the bolted joint is subjected to high magnitudes of cyclic
load there will not be any significant improvement in fatigue strength from applying a high
torque to firmly tighten the joint [6].
Since torquing the bolts up to yield stress does not ensure that they will not fatigue, a
large factor of safety needs to be used in any high-loaded bolt application. We can use the
Goodman relation to calculate the safe cyclic loading the bolts can withstand by using the
9
Figure 1-4: Bolts that have failed under large bending moment.
Figure 1-5: View of the failed bolts. [1]
10
[1]
Figure 1-6: Wind turbine in Lincolnshire, UK with one blade broken loose. [3]
11
b
.r4
4-D
-uts
Mean stress (am)
Figure 1-7: Goodman diagram used to calculate fatigue life of a material based on mean and
alternating stresses. [7]
mean stress and applied stress. For bolted joints, the mean stress is very close to the yield
stress +
induced stress. The Goodman relation is given as
gat = Of
\O'uts/
(1.1
where aol is the alternating stress, am is the mean stress, and of and uut, are the fatigue
limit and ultimate tensile strength (UTS) of the material. Goodman's relation is illustrated
in Figure 1-7, where the area below the curve defines the region which the material should
not fail given the stresses and the area above represents likely failure of the material. In
order to prevent failure of the bolts, the stresses on the bolts need to be known.
1.3
1.3.1
Current Design Flaws
Primary Issue
The slew rings, which attach the blade to the hub on a wind turbine, carry a very large
bending moment and axial load, both of which result from static and dynamic loads, and
the bolts which attach the slew ring to the hub and blade have to carry these loads. An
12
Metal will fail due to "tearing"
Case B
Load carried by inner and outer walls
and stresses are direct stresses.
-
Case A
Load carried by outer wall only
and stresses are bending stresses.
Figure 1-8: Structural analogy of hub/blade attachment. [8]
analogy of the existing design is shown in Case A in Figure 1-8, and the problems associated
with this design is obvious. An improved arrangement is shown in Case B where the load is
carried by both the inner and outer walls and this will be the basis of the new design.
1.3.2
Bolt Load Calculations
It is easy to calculate the load on the individual bolts due to the axial load, but calculating
the uneven load on the individual bolts due to the bending moment is not an easy task. The
problem is that the stiffness of both halves of the joint have to be considered and that is not
possible without resorting to finite element (FE) methods. In addition, the local stiffnesses
are not known so a full 3D model would have to be used. Even when using a 3D model some
assumptions have to be made about the boundary conditions and this can lead to an under
estimation of the load being carried by the bolts.
To overcome this problem it is proposed to replace the slew ring with a bearing pack which
separates the way the loads are carried. The advantage of this is that an easy calculation
can be made for the way the bearings inside the pack carry the loads and by reducing the
thickness of the blade, more of the blade can be in the form of an airfoil thus allowing the
blade to develop a greater torque.
1.3.3
Current Hub Design
The current design of the hub should also be reconsidered. Most wind turbine hubs are made
by casting large, heavy, steel pieces which are very expensive to machine as shown by the
13
Hub
Back Facing
Face Milling
Drilling
Profiling
Boring
Face and Shoulder
Milling
I'1.
Hub
Material: GGG70
i~-
H
Height: -4m
Weight: -11 ton
Weight to machine: -10%
Figure 1-9: Manufacturing processes of the current hub.
extensive manufacturing process in Figure 1-9. Fabricating the hub requires less machining
and would be a cheaper process. The current hub, shown in more detail in Figure 1-10, is
very large and not easily serviced. The blades have a large diameter because they need large
slew rings to take on the load. If an alternative bearing arrangement was used, the diameter
of the blades could be reduced and still produce the same power.
14
Figure 1-10: Exploded view of a wind turbine showing the attachment of the blades to the
hub via a slew ring. [9]
15
Chapter 2
Design Philosophy
The purpose of redesigning the hub is to eliminate the need for the use of FE modeling and
to be able to make an accurate estimate of the loads being carried by the bearings and the
bolts securing the bearing assembly. It is proposed to do this by separating the way the
loads are carried; that is the three major loads, gravitational force, centripetal force and the
torque due to the wind, are each to be carried by separate components of the structure. If
this is not possible, then components which have been specifically manufactured to carry
dual loads, such as taper roller bearings, should be used. A further objective of the redesign
is to improve the way the bearing assembly is sealed against the ingress of dirt and water,
be it rain or sea water. This will be done by using garter seals combined with a preliminary
seal, such as a labyrinth seal.
The new design should be such that it is possible to have as much of the assembly work
carried out on the "bench," away from the main structure. The bearing pack will allow this
by being pre-assembled and then simply bolted to the hub after. The hub should not require
complicated machining and be a simple structure. Maintenance of the mechanism should be
possible without the need to remove any of the major components.
16
Chapter 3
Design
A 3D CAD model (created in SolidWorks) of the open spindle and blade is shown in Figure 31. The proposed design of the wind turbine hub is shown schematically in Figure 3-2. It
should be noted that this is not an engineering drawing, hence elements of the drawing are
not drawn to scale. The design data is based on LM Wind Power's GloBladel which is a 1.5
MW wind turbine blade [10]. The specifications of the blade are shown in Figure 3-3.
Please refer to Figure 3-2 for the following reference letters (A-I).
A: Spindle Shaft
The spindle shaft is the rotating axis along which the pitch of the wind turbine blades will
be altered [11]. The shaft extends into the interior of the blade so that no bending moment
will be applied to the bolts. The shaft also extends into the hub so that the load of the
5.9 tonne blade will be supported by the interior and exterior walls as shown in Case B in
Figure 1-8.
In order to determine the diameter that the shaft needs to be in order to support the
bending moment being applied, the area moment of inertia for a solid shaft I, is used
1= -1rR,
4
(3.1)
along with the relationship between moment M, yield stress of steel ay (using 250 MPa),
17
Figure 3-1: Open spindle with the two bearings and gear, attached to the blade (Created in
SolidWorks).
18
TK
partial view in direction of arrows
Figure 3-2: Hub redesign drawing.
BLADE DATA
GloD9adel
LNM 40N3 P2
LM
37.3 P2
37
Blade length (meters)
42.1
Est. blade mass (kg)
5900
6
5590
Est. moment (kNm)
763
75
624
Est. AEP compared to LM 40.3 P2
5%
Est. AEP compared to LM 37.3 P2
+15%
Wind class
IIl/IV
n
Figure 3-3: Blade Data (looking at GloBladel for this report). [10]
19
and the distance to the neutral axis y,
M
I
- = - .
(3.2)
Given that the distance to the neutral axis is R, we find that the minimum diameter needed
to withstand the loads is 412.1 mm.
The material being used for the shaft will be rust and heat resistant steel called AIS1321
Annealed Stainless Steel which has an ultimate tensile strength (UTS) of 640 MPa. It
therefore has a factor of safety greater than 2. We will use a 500 mm diameter solid shaft
that has a slightly smaller diameter (2-3 mm decrease) on the bearing end in order for the
outer bearing to fit along the shaft and then be forced on.
The shaft will be manufactured by turning and facing to produce the lip extrusion that
will bolt to the blade. The slightly smaller diameter at one end will be produced with a rough
turning. The bearing seats will be finely turned or ground and the seal surface burnished.
The end of the shaft will be threaded to attach the locking nut (G) shown in Figure 3-2.
Both ends of the shaft will have chamfers to ease assembly.
B, F: Seal Housing and Seal
The seal housing will be made out of SG cast iron. It will secure the multi-layer seal that
protects the bearings. The current slew ring design does not have a feature protecting the
bearings from dirt and water. This new design will have a garter seal and labyrinth seal.
Figure 3-4 shows the Timken Model 64 Stainless Steel Garter Seal that will be used. One
design criteria is that the width of the garter seal must be less than that of the bearings so
that the seal can be bolted onto the bearing housing.
C: Bearing Sleeve
The bearing sleeve is made out of cast iron and will be custom made to fit the bearings,
the bearing sleeve being assembled away from the hub. This means that the bearings will
be less likely to be contaminated with dirt or water during assembly. The advantage of this
pre-assembly of the bearing pack is that no complicated assembly needs to be done inside
20
-odel 64
Millimeters
Inches
Product
Code
212384228
Model
Inner
Outer
Diameter Diameter
64-L
19.750
Width
21.750
0.875
Inner
Outer
Diameter Diameter
502
552
Width
22
Figure 3-4: Garter seal model and specifications. [12]
the hub. The finished assembly will then just need to slide in to the hub and be bolted on
to the outside.
All the bolts used in this design are M16 socket head cap screws in conjunction with an
even number of 16 mm dowels. The dowels are used to transmit torque because the bolts
should only carry the axial load. The UTS of the bolts is 1.3 GPa with a maximum load
of 204 kN each. The bolts that attach the blade to the shaft will be taking on most of the
load. The axial load will be due to the weight of the blade plus the centripetal force. This
can be calculated by using the equation for gravitational force
F = ma,
(3.3)
F, = mw 2 r ,
(3.4)
and the centripetal force
where the rotational speed of the turbine w, is 1.6 rad/s (15 rpm) and the distance to the
center of mass r, is 13.18 m (this was calculated from the blade specifications using the
moment and the mass).
This gives the centripetal force as 250 kN. With 16 bolts and 4
dowels securing the blade each bolt will take on 12.5 kN, which is far lower than their failure
21
load. These bolts will be about 100 mm in length since they need to fit between the bearing
and the lip for assembly but still extend through the lip and attach the blade.
D: Outer Bearing
In order to ensure the bearing pack in the new design would be able to withstand the high
loads, forces on the bearings were calculated. The three main forces the bearings have to
withstand are gravitational force, centripetal force and the torque due to the wind. The
axial load due to gravitational force and centripetal force was calculated above. The force
due to the torque of the wind F, can be found using the power relationship
P = wFyr.
(3.5)
Balancing these forces gives the force on each bearing, which are 1 m apart. The shear force
on the inner bearing is 1.85 MN and the axial force is 250 kN as the inner diameter must be
500 mm (from shaft diameter).
Single row tapered roller bearings will be used because they are able to take large axial
forces as well as sustain large radial forces. Spherical roller bearings were considered but
they have a degree of rotation that is not needed in this application. The specifications
for the Timken Tapered Roller Bearings which satisfy the load requirements are shown in
Figure 3-5.
E: Inner Bearing
The inner bearing will be taking on less load than the outer bearing and so will be safe to
have the same load rating as the outer bearing. The diameter of the inner bearing will need
to be slightly smaller because the shaft narrows slightly at the end in order for the outer
bearing to slide on.
22
TRB
Bearing Type
TS
Bearing Subtype
Cone (Inner)
M272749
Cup (Outer)
M272710
d - Bore
479.425
mm
D - Outer Diameter
679.450
mm
T - Width
128.588
mm
3730
kN
C90 - Dynamic Radial Load Rating - 90 M revs
968
kN
Ca90 - Dynamic Axial Load Rating - 90 M revs
551
kN
7400
kN
C1 - Dynamic Radial Load Rating - 1 M revs
CO - Static Load Rating
Figure 3-5: Timken Single Row Tapered Roller Bearings. [13]
23
G: Locking Nut
A locking nut will be put on the end of the shaft to secure it in place. It does not need to be
particularly wide since the majority of the load on a nut is carried in the first few threads.
H: Bevel Gears
On the end of the shaft is the bevel gear, which controls the pitch of the blade. The gear
must have large teeth since one or two teeth will carry all of the torque from the blade.
This gear will connect to a central bevel gear that will control the pitch of all three blades.
Therefore, its number of teeth must be a factor of 3.
I. Hub
The current design of the hub is very large, heavy, and expensive to machine. As shown in
Figure 1-9, the current hubs weigh about 11 tons and require at least 6 separate machining
processes to manufacture. The proposed new design of the hub is shown in Figure 3-6. Its
diameter is about 2.75 m and is an all-welded structure.
The front piece will be made from 20 mm steel plate. There is a front and back plate
that will secure the middle frame together. These discs will be flame cut from 20 mm plate.
The three holes in the plate have a dual function. The holes in each of the plates line up and
tubes equal in diameter will pass through them and be welded in place. This allows the two
side plate to be connected together and be stiffened mid-way between the inner and outer
rings. Their other function is to allow lifting cables to be safely attached to the completed
assembly. The middle section will be made out of three 120* segments of 20 mm plate that
will be butt-welded together. Cast steel tubes will be welded in the middle frame as shown
to secure the bearing shaft and to make the component easy to lift when attaching it to the
wind turbine frame (i.e. looping a rope through and raising). The flange at the end will be
made from 2 pieces of 25 mm plate butt-welded together and then connected to a web made
from 20 mm plate. This design is much more compact and simpler to manufacture. The
24
Figure 3-6: Exploded view of the new hub design.
empirical formula for the weldability of steel was derived by Dearden and O'Neill and is
CE=%C+
1
1
1
(%Cu+%Ni),
%Mn + (%Cr+%Mo+%V)+
15
6
5
(3.6)
where CE is the equivalent carbon content and has excellent weldability up to 0.35 [141.
25
Chapter 4
Conclusion
4.1
Summary
Gravitational force, centripetal force and the torque due to the wind all apply large forces
on a wind turbine blade and without a secure attachment to the hub these forces will break
the blade off. The current design of attaching the enormous blades to the hub with a slew
ring has been proven to fail by bolt fatigue since the bolts can not withstand these forces.
The objective of this project was to find an alternative design for the wind turbine hub that
would replace the current design and be able to withstand the high loads of the blade. This
report showed that the slew ring could be replaced by a bearing pack which can safely take
on the loads. The new design will not only be safer but also reduce the weight of the hub
and the manufacturing costs.
4.2
Future Work
Further work should be done on creating a 3D model of the new hub design and running
flow simulations to see if they produce the same force results as were calculated. As the root
of the blade can now be made much smaller, a new blade design should be considered. This
new hub design could also be implemented in other applications that involve altering the
pitch of blades relative to the hub. For example, it could be implemented in variable pitch
propeller on aircrafts and ships.
26
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University of Oxford.
[3] M. Moore, "UFO wind turbine 'broke due to mechanical failure not collision with flying
object'," The Telegraph. Viewed at www.telegraph.co.uk in November 2013.
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fatigue life and crack growth in torque tightened bolted joints," Aerospace Science and
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2013.
[11] J. Manwell, J. McGowan, and A. Rogers, Wind Energy Explained: Theory, Design and
Application. Wiley, 2010.
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[13] "Timken bearing data." Retrieved from www.timken.com in April 2013.
[14] N. Bailey and W. Institute, Weldability of FerriticSteels. Series in Welding and Other
Joining Technologies Series, Abington, 1994.
27