Testing of rotor blades of wind turbines Arno van Wingerde ...

Testing of rotor blades of wind turbines
Arno van Wingerde, Fraunhofer Center Windenergie und Meerestechnik
Fraunhofer-Center für Windenergie und Meerestechnik, Am Seedeich 45, D-27572 Bremerhaven.
tel.: +49-471-902629-23; E-Mail: vanwingerde@cwmt.fraunhofer.de; www.cwmt.fraunhofer.de
Summary
With wind turbines set to reach 200 m diameter and rotor blades approaching lengths
of 90 m [1], major investments are needed to develop new concepts. The certification
of the rotor blades is a major cost factor and the industry and research institutes are
looking into possibilities to optimize this process as well. The process entails material
tests as well as static and sometimes cyclic testing of the blade. Especially the latter
tests, compulsory by some wind turbine rotor blade certification agencies, are a major
cost post, since the tests can take over half a year to completion, thus delaying
the time-to-market. Also, the level of realism of these tests can be questionable.
Therefore the need arises to check alternatives such as component testing.
Another problem for the further development of offshore applications is the increased
need for reliability of wind turbines. For instance on the North Sea, wind turbines are
all but inaccessible for half a year at a time, so that even minor technical problems
result in major financial losses.
Scaling aspects of wind turbines
There exists a clear trend with the industry towards ever larger wind turbines. The
size of some wind turbines versus the year of introduction can be seen in
Figure 1.
Boeing
747
Figure 1 Trends in wind turbine size
Rational behind for upscaling and the square-cube law

Apart from marketing reasons where there is a certain advantage associated with
having the largest/fastest/best … there are also some technical and economical reasons
for upscaling, especially for offshore wind turbines, where certain aspects, such
as installation, maintenance, cables (offshore) and the foundation costs, do not fully
scale to the diameter of the turbine. For deep water offshore, these aspects account
for 40%-60% of the total cost of the wind turbine.
However, there exists a major hurdle called the square-cube law which all but prevents
economical upscaling of the wind turbine. When the blade is simply scaled up,
the size of the swept area, the disk of air captured by the wind turbine, and thereby
the potential energy production, increases with the square of the rotor diameter.
However, just scaling up the 3 dimensions of the blade will result in a weight that increases
with the third power of the structure.
Influence of the dead weight
The cubically increasing dead weight of the blade with the rotor diameter is a major
problem for the industry but, for large blades, the weight of the blade itself becomes a
major load component. Thus extra material is needed to carry the weight which in
turn further increases the weight of the blade.
An example may serve to get a feeling for effect of the dead weight of the blades with
increasing lengths. Consider the maximum length of a rod with a diameter D and a
length l, which is clamped at one side and only loaded by its dead weight, as shown
in Figure 2Error! Reference source not found.. Furthermore the rod is made of
glass-epoxy, the predominant material used for blades of wind turbines, with the specific
density ρ of glass-epoxy of about 2.3, or 23000 N/m 3 and a strength of 800
L
l
D
MPa= 800·10 6 N/m 2 .
Figure 2 Dead weight problem
Notation used:
D: the diameter of the rod
L: the length of the rod
F: the dead weight of the rod (which can be considered to act halfway the rod)
M: the bending moment at the clamped end of the rod, due to the dead weight
W: the elastic moment of the rod
σ max : the maximum bending stress in the rod
The dead weight of the rod, also the load on the rod is: F=L·¼·π·D 2·ρ
(1)
The resulting bending moment M at the clamped end is: M=½·F·L
(2)
The maximum stress is the rod, σ max = M/W
(3)

The elastic moment W of a rod is: W= 1 / 32· π·D 3
(4)
The maximum allowable moment M follows from (3): M= σ max·W
Combining (1) and (2): M=½·L·¼·π·D 2·ρ·L. The maximum length L follows therefore
from: ½· L·¼·π·D 2·ρ·L = 800·10 6·1/ 32·π·D 3 or: L =√ (200·10 6·D/23000).
For instance a rod with a diameter of 1 m could have a maximum length of 93 m before
breaking under its own dead weight. This is the length of future rotor blades!
Obviously blades of wind turbines have a considerably more favourable shape: a larger
radius, a hollow cross section, parts of which are executed as sandwich panels,
the blades are not prismatic but smaller towards the tip etc. Still, from this simple example
it can be gathered that the dead weight is becoming a major load case in itself
for larger diameter wind turbines. In order to establish such structures in an economically
feasible way, the material will have to be utilized optimally.
If the same rod is made of carbon-epoxy, with the same strength, but a lower density
of about 1.5, the maximum length becomes: L =√ (200·10 6·D/15000) =115 m, a bit
better than glass-epoxy.
Instead, using a glass-epoxy hollow rod would result in: W= 1 / 32· π·(D 4 -d 4 )/D and
F=L·¼·π·(D 2 -d 2 )·ρ. Consider a rod with the inner diameter d=0.8* the outer diameter
(thus, a wall thickness of 0.1 D). Then, W= 1 / 32· π·0.59·D 3 and M=½
·L·¼·π·0.36·D 2·ρ·L. The maximum length follows from ½·L·¼·π·0.36·D 2·ρ·L
= 800·10 6·1/ 32· π·0.59·D 3 , or L =√ (200·10 6·D·(0.59/0.36)/23000)=119 m, again not all
that much better.
Consider a rod with the inner diameter d=0.98* the outer diameter (thus, a wall thickness
of 0.01 D). Then, W= 1 / 32· π·0,039·D 3 and M=½ ·L·¼·π·0,020·D 2·ρ·L.
The maximum length follows from ½· L·¼·π·0.020·D 2·ρ·L = 800·10 6·1/ 32· π·0.039·D 3 ,
or L =√ (200·10 6·D·(0.039/0.020)/23000) =130 m, 40% longer than the massive rod.
Scaling in practise
From the square-cube law, is can be expected that the weight increases with the
third power of the rotor diameter and possible worse due to the increasingly important
influence of the dead weight of the blades. However, the wind industry has continuously
worked on using the material more efficiently, so that the actual increase of
the blade weight with the rotor diameter is considerably lower.
It would seem a tribute to the technological vigour of wind energy industry that the
actual increase of the blade weight is only about rotor diameter to the power 2.4, rather
than the factor 3 that the square-cube law would predict. However, according to
newer trends observed in [2], this is only the case when very old, not optimised
blades are taken into account. In case of relatively newer blades, the increase in
weight is actually in accordance with the square-cube law. Up to a point, this is actually
to be expected as the major inefficiencies in blade design have been eliminated
for some time now.

Carbon fibre reinforced materials (CFRP), which exhibit a higher (fatigue) strength at
a lower specific weight as well as a higher stiffness. These materials would make an
attractive alternative to glass fibre, especially if used in high loaded areas and in the
tip [3], but the problems in securing the needed large quantities of the carbon fibre
has prevented widespread use in the past. Furthermore, unidirectional CFRP is notoriously
sensitive to misalignment, leading to premature buckling, see 0.
Structural verification of Rotor Blades
In order to be able to insure wind turbines, a certification of the turbine by a certification
body such as Germanische Lloyd (GL) [4] or Det Norske Veritas (DNV) [5] is
typically needed. The experimental verification of the structural integrity of the blades
using full-scale testing of blades has been documented by the IEC [6].
Material testing
Lacking general data, certification bodies generally allow a conservative estimation of
the fatigue performance of the material. For instance, GL allows the derivation of the
fatigue properties from the static property using a slope of 1:10 for epoxy. In practise
the slope of the S-N line can be 1:12. Also the constant life diagram, used for S-N
lines with other stress ratios has been shown to deviate significantly from the simple
triangular diagram that was adopted in the GL guideline [7]. Another material aspect
is the decrease in static strength due to fatigue loading as discussed in [8]. Therefore,
in order to fully utilize the material, most manufacturers have their selection of
materials tested more extensively. A material test can consist of a few static tests to
determine static strength and stiffness and use conservative rules to estimate fatigue
performance (which is quite OK in cases where fatigue is not the driving issue) up to
full S-N curves and constant life diagrams where the material needs to be fully utilized.
Static tests
Blades are generally tested for static strength by bending the blade in flap-wise (perpendicular
to the plane of the rotor) and edge-wise (in the plane of the rotor) directions.
If the blade is loaded in pure flap-wise and pure edge-wise direction only in
both directions, this results in 4 load cases. In practise tests in other directions are
also carried out. Although the design is checked separately, still some blades fail
these tests, because of poor production quality. The static test is therefore a useful
way for checking the structural integrity of the blade.
Buckling
Buckling often occurs either because of misaligned fibres or because of failing
bonded joints. A material particularly sensitive to buckling due to misalignment is UD
direction carbon-reinforced plastics. The carbon fibre has a typical strength of about
2000 MPa, whereas the epoxy is limited to about 30 MPa. It is clear that in case of a
compressive loaded part unidirectional carbon-epoxy, the discrepancy between
strength and stiffness of the fibre and the matrix material is such that even a slight
misalignment of the carbon fibres cannot be overcome by the matrix material and
hence buckling occurs. In practise, failure at less than half the nominal strength has
occurred during actual blade tests due to this phenomenon.

Failure of bonded joints
A rotor blade is typically made up of two outer parts, bonded together at the ends as
well as a number of stiffeners inside.
The bonded joints, particularly those of the stiffener, often fail prematurely because of
serious production failures, for instance the bonding paste at some cross section is
simply not applied at the right location, shown in Figure 4. It is not too hard to image
how this can happen, since when the mould is closed around the two halves, the
relatively flexible stiffener has to fit perfectly over a length of 50 m.
Skin
Stiffener
Bonding
paste
Stiffeners
Cross section
of a blade
Detail showing bonded joint
to stiffener
Figure 4 correctly bonded joint (photo, A&R) and poorly bonded joint (drawing)
Stress distribution in the cross section
If a blade is bent in flap-wise and edge-wise directions, in both positive and negative
directions, the extremes of the cross section are all tested for tensile and compressive
stresses. However this does not mean that the complete cross section is adequately
tested, see Error! Reference source not found..

Flap-wise loading
IFAM 4/9/07 2:42 PM
Formatted: Indent: Left: 0.63 cm
Highly loaded
Mildly loaded
Virtually unloaded
Loaded in Shear
Figure 5 Stresses in a cross section of the blade
A test in flap-wise direction causes the widest parts of the blade to be highly loaded,
darker parts in Error! Reference source not found.a, the shear webs are loaded in
shear. For an edge-wise loading the picture is rather opposite: the leading and trailing
edge are loaded whereas the widest part of the blade is virtually unloaded (apart
from shear in the sides), see Error! Reference source not found.b. The problem is
that the grey parts are not loaded to their maximum in either case and are therefore
not adequately tested. Also, some buckling panels in the cross section are now conveniently
supported by nearby parts loaded in tension, whereas this would not necessarily
be the case for a ± 45° loading. This problem can be mitigated by loading
the blade in more directions, besides pure edge-wise and flap-wise.
IFAM 22/4/08 6:23 PM
Formatted: Font:10 pt, English (UK)
IFAM 22/4/08 6:23 PM
Formatted: Font:10 pt, English (UK)
Bending moment distribution along the blade
Not only the stress distribution across the cross section is an issue in proper rotor
blade verification, also the bending moment distribution along the length of the blade
has to be observed. However, applying an appropriate loading is not entirely trivial. In
its service life, the blade is loaded by a distributed loading, whereas in testing typically
concentrated loads are applied to the blade by actuators. As a result, the moment
distribution in the blade is inaccurate, see for instance Figure 6a where the
bending moment distribution due to the concentrated load at the tip deviates considerably
from the actual bending moment distribution that the wind turbine is subjected
to in practical application. There are essentially two ways to counter this effect:
Static test at several positions simultaneously
The first option is to load the blade at several positions simultaneously, which results
in the bending moment distribution of Figure 6c, which for the example shown seems
to approach the actual moment distribution fairly accurately.
However, the area where the load is introduced itself is not tested properly, because
there are massive, typically wooden, blocks which clamp the blade at the position of

the load introduction. Because the loads take a bit to distribute properly across the
cross section of the blade (say 1x the width of the blade), the areas next to the load
introduction are not tested properly as well. Assuming a zone of the width of the
blade next to the load introduction blocks, yields fairly large areas which are not
tested properly, shown as the shaded areas in Figure 6c. In this way, the load cases
considered here (flap-wise and edge-wise in both directions as discussed previously)
can be carried out as four static tests.
Static test at several positions subsequently
Instead of putting on loads at two positions simultaneously, it is also possible to carry
out two separate tests. The first one with only a load near the tip, as shown in Figure
6a, and subsequently applying a load at another position, as shown in Figure 6b.
This test procedure allows almost the whole blade to be tested: only the dashed area
near the load introduction of Figure 6a is not tested properly. The main disadvantage
is that each test (in case of only purely flap-wise and edge-wise, this would be altogether
4 tests) has to be carried out for each position separately, for a total of 8 static
tests. The advantage is that using this procedure, almost the whole blade can be
properly tested with only marginally more effort, so that this test procedure is preferred
over the simultaneous test procedure.
Ok
LOAD 3
INVALID: SUPPORT
LOAD 2
LOAD 1+3
Bending moment
due to test load
Actual bending moment
distribution
INVALID: LOAD
INTRODUCTION
UNDERLOAD
UNLOADE
D
a) Bending moment distribution along blade axis due to a concentrated load near the tip
LOAD 3
LOAD 2
Actual bending
LOAD 1+3
moment distribution
Bending moment
due to
test load
b) Bending moment distribution along blade axis due to a concentrated load

LOAD 3
LOAD 2
LOAD 1+3
c) Bending moment distribution along blade axis due to two concentrated loads at the blade
Figure 6 bending moment distribution along the blade length
Fatigue testing of rotor blades
A fatigue test of a rotor blade is notoriously more complex than a static test. All the
previously outlined problems, regarding the need for a clean representation of the
actual stress states across the cross section and along the length of the blade are
still there, but a number of additional problems occur.
Required testing time
Actual bending
moment distribution
Bending moment
due to
test load
Wind turbines experience about 10 8 -10 9 cycles, much more than virtually any other
known structure, see Figure 7. Also, the loadings vary more than other structures,
making fatigue testing of rotor blades particularly difficult.
Figure 7 Overview of fatigue loaded structures [9]
Assuming a frequency of 1 Hz, 10 8 -10 9 cycles would take about 3 to 32 years, which
is highly impractical. Therefore the tests are carried out at a raised load level, so that
10 6 or 2·10 6 cycles would suffice to reach the equivalent fatigue damage.
However, many blades are tested at their natural frequency, which is typically below
1 Hz, especially for larger blades, in which case 0.3 Hz. and lower are possible. Testing
outside of the natural frequency is possible, but requires vastly more force and

hence energy for testing, raising the cost of the blade test considerably. If a blade is
tested at higher frequencies, the test may be harder to carry out due to higher harmonic
excitations of the blade. An interesting option to increase the natural frequency
of the blade is the removal of the tip part, which also helps to limit the maximum deflections
of the actuators, which would otherwise exceed strokes of 10 m.
Testing a blade for a million cycles at 0.3 Hz takes only 38.5 days, slightly over one
month, and is therefore entirely feasible. Since the stresses are raised and the
maximum stresses are already close to the static strength, the spectrum would be
compressed anyway, making the step to a constant amplitude loading a small one.
Required fatigue loading
Many fatigue tests have been carried out as a single loading at a single cross section,
so that the bending moment is only correct for a limited length of the blade.
Typically the blade is first loaded in flap direction and then in edge direction, thus
doubling the total test time. Carrying out the test at another cross section, as was
suggested for static tests, doubles the total testing time yet again to the better part of
a year. This would be a major cost factor for the manufacturer, who has invested
large sums in the design and moulds of the new blade and wants to certify and produce
and sell the blades as soon as possible. Furthermore some parts might actually
be overloaded due to the subsequent fatigue tests.
Instead, for the fatigue test it is suggested to load the blades at several cross sections
simultaneously, so as to get the bending moment distribution right along the
length of the blade, rather than just at a limited length. Moreover, the blade should be
tested bi-axially, so as to get the whole cross section properly loaded. A bi-axial test
is carried out with cylinders in two directions at a cross section, so as to make the
blade move in an elliptical way, rather than a uni-axial movement. However, controlling
the loading in two or more cross sections, while at the same time loading the
blade in two directions, poses a major challenge for the controller of the test set-up.
In spite of its shortcomings and difficult execution, fatigue tests have often exposed
failures in blades which had already passed the static tests. Hence the fatigue tests
must be considered a valuable check on the integrity of the blade. On the other hand,
the effort and time spent is growing for larger blades and the realism of the fatigue
test is often questionable, because of an unrealistic stress field in the blade. The
aforementioned disadvantages are the reason for the search for establishing alternative
test methods.
New Testing Methodology for Rotor Blades
Currently, a debate is raging worldwide concerning the need for compulsory fatigue
testing in the certification process. Given the previously discussed large technical
and economical problems associated with a fatigue test of a blade, the industry and
research institutes are currently investigating possible alternatives. Although these
alternatives may not totally eliminate the need for fatigue testing of rotor blades, they
could reduce the absolute need for such tests, especially in case of modifications of
existing blades or scaled-up version of an existing design.
A possible new testing methodology has been suggested in an earlier publication [9].
In order to replace full size fatigue blade tests, emphasis is given to material tests
and tests of simple components, supported by more extensive numerical analyses.

Material tests were used before to determine the basic material properties, but could
perhaps also be taken from actual blades to reflect the actual material properties of a
blade, as opposed to just some generic material properties from coupons that are
produced entirely separately from the blades, sometimes by other companies and
hence may not fully reflect the actual properties of the material in the blade, such as
the misaligned fibres discussed previously in the paper.
The material properties are used to establish numerical models of the blade. Some
critical areas of the blade could be modelled by performing simple component tests,
other areas will require more complex component tests, such as the T-bolt “IKEA”
connections typically used to connect the blade to the hub.
If a component is thought of as a part of a blade, it is generally not trivial to obtain the
correct boundary conditions of a component as they would occur in the blade, which
requires either a complex test set-up or complex test specimens. However, via the
numerical model, it is possible to establish an “equivalent” test which would be as
severe a case, but with more basic boundary conditions. In other cases, scaled-down
models might be tested to see whether they concur with the numerical results.
Other problems that seem to lend themselves well to component testing would be
either potential buckling problems or bonded joints. The use of smaller specimens
instead of a full blade also allows for more extensive testing with for example fatigue
tests at different stress ratios, or tests at other temperatures and at raised humidity
levels. Moreover, rather than just testing a single test specimen, a full series can be
tested, revealing important statistical properties of the production. Conceivably
specimens could be produced along with the blades (comparable to concrete cubes
used in the construction business to check the quality of concrete structures), to
check the consistency of the production in time. If implemented successfully, the new
testing technology, which reflects long established practises in airplane construction,
could result in a raised level of reliability, at reduced overall testing costs.
Conclusion
Performing and analysing static and fatigue testing of rotor blades of wind turbines
poses major challenges for test centres, blade manufacturers and certification bodies
alike. With the increasing size of wind turbines, testing the rotor blades is becoming
increasingly costly and time-consuming. The compulsory fatigue testing of rotor
blades might become a hindrance for the further development of the industry. Therefore,
an alternative test method, based on static blade tests and component tests, as
well as more extensive material tests might offer a way to satisfy reliability demands
at considerably lower costs. Especially in case of several blades of one family with
minor variations, or in case of scaled-up blades, the need for dynamically testing the
full blade may be all but eliminated using this test philosophy. This requires the definition
of accepted component tests, which should occur in a body like the IEC, a
committee is being formed to set up these tests.
Acknowledgements
The authors are grateful to valuable discussions with the German Competence
Group Wind Energy „Rotor Blade Testing and Rotor Blade Materials“.
The work is supported by the State of Bremen, Senate of Civil Engineering, Environment
and Transportation and Bremerhaven Economic Development Company
Ltd, the Federal Ministry for the Environment, Nature Conservation and Nuclear

Safety and the Federal Ministry of Education and Research with support of the
„European Regional Development Fund ERDF”.
The research project OptiMat Blades was funded in part by European Commission in
the framework of the specific research and technology development programme Energy,
Environment and Sustainable Development.
References
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Wind Energy?, DEWEK 2004, Wilhelmshaven, October 20-21, 2004
2. WIND ENERGY - THE FACTS, European Wind Energy Association, 2003
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symposium Wind Turbine Fatigue IEA Joint Action. Delft October 25-26 1999. Organized by:
TU Delft. ISSN0590-8809 CFRP
4. Guideline for the Certification of Offshore Wind Turbines, Chapter 6.2. Germanischer Lloyd,
Nov. 2003
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Norske Veritas, DNV-OS-J102, April 2007.
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Electrotechnical Commission 23, IEC/TS 61400-23, April 2001.
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of Technology, The Netherlands, 2006.
9. H-G Busmann, Ch. W. Kensche, A. Berg-Pollack, F. Bürkner, F. Sayer, K. Wiemann, Testing
of Rotor Blades, DEWI Magazin nr. 30, February 2007, pp.5-9.