OMMI, 2008, Volume 5, Issue 1, May
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HRSG optimisation for cycling duty based on Euro Norm EN 12952-3
Ir. Pascal Fontaine1 and Ir. Jean-François Galopin2
1
Marketing Manager CMI Energy Av Greiner 1, 4100 Seraing BELGIUM.
2
Design Manager CMI Energy Av Greiner 1, 4100 Seraing BELGIUM.
pascal.fontaine@cmigroupe.com jean-francois.galopin@cmigroupe.com
Mr Pascal Fontaine.
In 1990, Mr Pascal Fontaine graduated as professional Engineer in Power Generation from the University of
Liège in Belgium. Within CMI Energy, he gained his experience on HRSG design and management, where he
occupied the following positions: Process Engineer for 4 years, Technical Project Manager for 5 years,
Proposal Manager for 4 years, and currently Marketing Manager for CMI Energy. Mr Fontaine is the author of
several papers and international publications on both Horizontal and Vertical HRSG technologies and HRSG
cycling related issues.
Abstract
Nowadays, most of the combined cycle power plants are working on day/night and
weekday/weekend cycling. The onset of deregulation and consequent merchant power have
brought combined cycle plants to supply electrical power to the grid as and when needed with
minimum notice. Combined cycles are also often forced to run on partial loads. Even units
which were originally designed for base load are all eventually forced to cycle as new more
efficient power plants are built. As generally recognised nowadays, the cycling criterion is to
be integral part of the HRSG design. In this paper, the HRSG fatigue analysis using the
European Norm EN 12952-3 will be discussed. This paper will present how the European
Norm can be used to assess HRSG cumulative damage and how the HRSG cyclic life time
can be optimised.
Key Words. European Norm EN 12952-3, HRSG, Fatigue
1. Introduction
Nowadays, most of the combined cycle power plants are working on day/night and
weekday/weekend cycling. The onset of deregulation and consequent merchant power have
brought combined cycle plants to supply electrical power to the grid as and when needed with
minimum notice. Combined cycles are also often forced to run on partial loads. Even units
which were originally designed for base load are all eventually forced to cycle as new more
efficient power plants are built. As generally recognised nowadays, the cycling criterion is to
be integral part of the HRSG design. In this paper, the HRSG fatigue analysis using the
European Norm EN 12952-3 will be discussed. This paper will present how the European
Norm can be used to assess HRSG cumulative damage and how the HRSG cyclic life time
can be optimised. The theory is first reviewed as a base to explain how Euro Norm can be
used for cyclic optimisation. In complement to construction code (i.e ASME), this paper will
explain how the boiler cyclic life can be optimised with progressive pressure start-up ramp
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rates reducing the overall start-up time. This paper focuses on fatigue and we will not discuss
here the interaction mechanisms between fatigue and creep on hot superheater.
2. Stress over component thickness
There are mainly two kinds of stress inside component wall: the mechanical stress originated
from internal pressure, and the thermal stress originated from thermal expansion. During startup, those stresses are in opposite directions, while they are in the same directions during
shutdown. Eventually, these stresses combine together into Von Mises stresses (Fig. 1).
Thermal stress sth is:
- fonction of transient f(t) only
- proportionnal to delta T?over e
- compression during start-up
- traction stress during stoppage
Insulated
Q ext = 0
sth
=
sp
+
sth
Ma
in
co
mp
on
e
nt
dia
me
ter
D
sTotal
Internal
Pressure,
Temperature
= fonction (time)
sp
sp ~=
Thickness e (up to 6 inch for HP drum)
e
e
Pression * Diameter
2*e
sTotal Stress
Mechanical stress sp is:
- proportionnal to pressure P
- proportionnal to diameter D
- nearly constant over thickness
- always a traction stress
Fig. 1 Mechanical and thermal stress over component wall thickness during transients
3. Stress cycle during a cold start-up
A cold start-up, meaning from sub-cooled water, is special compared to a hot start because the
inner shell is heated up while the internal pressure is atmospheric and constant. That induces
compression thermal stress which is not yet counter-balanced by any tensile stress from
pressurization. As soon as pressure starts to increase, mechanical tension stress will release
thermal stress as they are in opposite directions. Therefore, the largest amplitude of this stress
cycle will occur during such a cold start-up, which is the most severe condition. In addition,
the internal protective magnetite layer could crack during the early stage of start-up. This
ferrous oxide Fe3O4 is created naturally under pressure at drum normal operating condition.
The Euro Norm allows an empirical stress range of -600 N/mm² and +200 N/mm². As this
magnetite layer is formed under pressure, it is already under compression strain at cold
conditions, and therefore, subject to overpass the lower limit during a cold start-up (Fig. 2).
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OMMI, 2008, Volume 5, Issue 1, May
(N/mm?
www.ommi.co.uk
Inner shell surface
circonferentiel stress
Forbidden zone according to Euro Norm 12952
+ 200
MaxsT
sTnom
MaxsT
P, sp
- 600
Overall material
fatigue range:
MinsT
T
sTOTAL =sth+ sp
Time
sth
MinsT
START-UP
Forbidden zone according to Euro Norm 12952
(internal magnetite layer protection)
STOPPAGE
Fig. 2 Schematic approach of the cycling stress range over drum thickness for a cold start-up
During the subsequent pressure increase, the drum shell is heated from the inside by boiling
water and the heat is transmitted from inside to outside generating a temperature profile. A
significant temperature difference between inner and outer shell surfaces builds up. The outer
shell surface temperature lags behind the inner one. This temperature profile over the drum
thickness is constant and slides up (Fig. 3). This shifting temperature profile has roughly a
quadratic shape versus wall thickness, and so is the induced thermal stress profile. Once
pressure is stabilized at the normal operating condition, this temperature profile will vanish,
so will thermal stress, and only this mechanical stress will remain in the equipment wall.
During shutdown, the same opposite phenomenon occurs with reversed temperature profile
between external and internal surfaces on drum shells. This complete cycling stress, including
start-up and shutdown, shall be considered for component fatigue analysis. Regardless of the
stress transients, the ASME code considers the permanent operating conditions on creep for
thickness calculation.
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OMMI, 2008, Volume 5, Issue 1, May
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F in a l
te m p e ra tu re
u n ifo rm o v e r
th ic k n e s s
T
In te rn a l
w a ll
s th
t1
t2
T e m p e ra tu re
p ro file is
s lid in g
E x te rn a l
w a ll
t3
t4
t5
T e m p e ra tu re
p ro file is
b u ild in g
T im e
P ro file s o f
te m p e ra tu re a n d
th e rm a l s tre s s
In te rn a l
w a ll
E x te rn a l
w a ll
D ru m w a ll
th ic k n e s s
T e m p e ra tu re
p ro file b u ild in g
T e m p e ra tu re p ro file
slid in g
T e m p e ra tu re
p ro file v a n ish in g
T im e
Fig. 3 Schematic evolution of stress and temperature on internal/external walls during start-up
4. Impact of component thickness
Temperature difference between inner and outer walls is roughly proportional to the square of
the drum thickness, and so is also the thermal stress. This is why thin walls are so beneficial
in terms of cycling fatigue. For instance, it is generally considered that the thickness of a
superheater outlet header made of P91 material can be reduced by a factor of 2 compared to
the same header made in P22 material. Consequently, the induced thermal stress would be
reduced by a factor of 4 for such P91 header compared to P22 header (Fig. 4).
Thincomponent
sth
Temperature/ thermal stressprofiles
Thickcomponent
4
DT°'
1
DT°
e
If e' =2e, thensth' =4sth
Time
e' =2e
Fig. 4 Illustration of thermal stress which is proportional to square of component thickness
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OMMI, 2008, Volume 5, Issue 1, May
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5. Material fatigue induced by cycling stress
This start-up/shutdown cycle represents one simplified stress range. The allowable number of
cycle is calculated with the Euro Norm EN 12952-3 (Fig. 5). In practice, cold, warm and hot
cycle stress numbers are considered as per plant cyclic service informed by the specification.
Other stress cycles can also be accounted, such as partial cycle when the first unit is started on
a 2-2-1 arrangement, or even Low Cycle Fatigue (LCF) due to attemperation in operation.
Then, the Miner’s rule is used to account each of those fractions of cumulative fatigue
damage. Application of the norm shows that a cold start is up to 20 more damaging than a
warm start, and that the stress range resulting from a hot start is typically below the fatigue
limit and not contributing to the total fatigue damage (except for the damaging quenching
issue, see §7). The fatigue damage is very sensitive to stress range because of its logarithmic
nature (see the double logarithmic scale of Fig. 5). A small variation in stress amplitude
impacts largely the corresponding number of cycles, even more so for small amplitude 2fa of
warm start-ups. Fatigue calculation does not fix exactly the line between a crack and a nocrack initiation, but it is rather a statistical probability of crack occurrence under Na, fa
conditions, eventually representing a percentage of risk of failure. The sensitivity and
probabilistic nature of fatigue results in an uncertainty in fatigue lifetime analysis. Robust
modeling by finite element can reduce the other uncertainties which come from simplified
stress ranges and boundaries.
Fig. 5 Wöhler curve from EN 12952-3 showing material fatigue for symmetric stress range
(amplitude fa) versus allowable number of cycles Na for various material tensile strength Rm.
5. Allowable temperature ramp rates for start-up and shutdown
ASME I considers continuous operation at design conditions, but it does not mandate
assessment for fatigue analysis. European boiler manufacturer perform this analysis using the
European Norm 12952-3 to calculate the acceptable pressure/ time gradients. The European
Norm 12952-3 derives largely from the German code TRD 301, which was commonly used
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previously for cyclic assessment. Whereas the HRSG is designed according to ASME, it is
checked, in addition, for fatigue analysis using Euro Norm. Input data are basically the
components diameter, material, thickness and operating pressure and temperature. Two
methods can be applied, either the number of cycles is calculated for given start-up and
shutdown rates, or those allowable rates are calculated for a given number of cycles. An
interesting feature is the ramp rate variation versus operating pressure (Fig. 6): as pressure
increases, the allowable pressure ramp rate also increases. In practice, EN fatigue analysis is
applied to the thick HP drum wall as well as to outlet headers of reheater and HP superheater.
Fig. 6 Example of results from TRD301 (European Norm 12952-3) for determination of acceptable
temperature gradients on an HP drum 90mm thick in SA302B up to 110 bar A.
As explained above, a cold start is the most stressful event in terms of fatigue effect,
inasmuch as it includes the largest cycle range. Unless required otherwise by a specification,
the number of cold start-ups considered for design by CMI is 2000 over a 20-year design
lifetime. By applying the norm for an extended number of cold start-ups, we calculate
negative temperature gradients (Fig. 7). This is of course impossible, and it shall be
interpreted as a statistical fact that material would have no more strength reserve to perform
such extended number of start-ups, meaning that such cumulative cycling damage has
exhausted the component fatigue lifetime.
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Fig. 7 Euro Norm results showing temperature gradient versus number of cold start-ups
7. Euro Norm 12952 improvements compared to former TRD 301 code
TRD 301 was generally looked upon as the reference for pressure vessels cycling analysis.
However, when compared to finite element (FE) analysis, this code revealed itself to be rather
conservative. For instance, TRD considered empirical stress concentration factor around
nozzles which were proven to be high by FE, as cautions are taken in construction to avoid
sharp edges on those openings. Also, the feedwater nozzle on drums is always equipped with
an internal thermal sleeve to accommodate the induced thermal shock without undue stress
occurrence when cold feedwater is first admitted into a warm drum. The main improvement
of the Euro Norm compared to the former TRD 301 for cyclic analysis is on this
determination of stress concentration factors, as finite element analysis can be used (Fig. 8).
Fig. 8 Finite Element nozzle analysis for stress concentration calculation
In order to fully apprehend all changes between EN 12952-3 and TRD 301, an overall survey
of the two codes was conducted by CMI on a specific boiler under engineering. TRD 301 and
EN 12952-3 calculation methods are similar, but some differences have been identified,
mainly related to the empirical coefficients used within the formula on stresses evaluation:
σ = αm .
β .E
d ms
. p + α t . t t . ΔT
2 . ems
1 −ν
mechanical
stresses
with :
dms
thermal stresses due to
temperature gradients
= equipment diameter
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OMMI, 2008, Volume 5, Issue 1, May
ems
αm
p
αt
ν
βt
Et
ΔT
www.ommi.co.uk
= equipment mean thickness
= stress concentration factor for membrane stresses
= pressure
= stress concentration factor for thermal stresses
= Poisson’s ratio
= differential linear thermal expansion coefficient
= modulus of elasticity
= temperature variation with respect to time
Comparison of results is as follows:
stress
concentration
factor α m
stress
concentration
factor α t
TRD 301
The 2 codes use
different
correlations.
EN 12952-3
It seems to be more precise due to the fact that it
incorporates the branch mean thickness within its
formula (as opposed to TRD 301).
α t is fixed to 2 or
α t is evaluated with the help of the mean branch
1.5 depending on
the head shape
(cylindrical or
spherical)
diameter (dmb), the mean body diameter (dms) as well as
the heat transfer coefficient h :
2
h
h + 2700
⎡
⎤
α t = ⎢2 −
.z +
. (exp( −7.z ) − 1)⎥ + 0,81.z 2
h + 1700
h + 1700
⎣
⎦
in which
: z=
⎧1000 W / m² K
h=⎨
⎩3000 W / m² K
stresses
calculations :
TRD 301
Ck and Ct*
(or equivalent
parameters) are
not used
d mb
d ms
for steam
for water
EN 12952-3
A notch factor (Ck) is used to take into account the
influence of surfaces and weld finish:
σ ∗ = σ . Ck
Where Ck=1 for our application
A corrective factor (Ct*) is used to take into account the
influence of temperature as well as the type of material
(either ferritic or austenitic):
σ ∗ (t ∗ ) =
comparison
criteria
The sum of the
cumulated
damages shall be
smaller than 50%
σ∗
Ct ∗
In which : t* = 0,75 . tmax + 0,25 . tmin
(the same definition is used for t* in TRD 301)
The sum of the cumulated damages shall be smaller
than 100%
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As mentioned before, the critical components are the HP drum and high temperature
superheater / reheater headers and/or manifolds. The hereafter results are based on detailed
connection between branches (nozzles) and main body on an HP drum and an HP superheater:
Drum
SA 302 B
2170
122
SA 105
526.52
80
Superheater
SA 335 P91
219.1
28.6
SA 213 T91
38
4
Type of start-up
cold
warm
hot
50
1250
5000
7.97
7.97
7.97
Type of start-up
cold
warm
hot
50
1250
5000
7.97
7.97
7.97
294.7
294.7
294.7
565
565
565
8
-1.8
0.08
7
-1.8
1.65
6.5
-1.8
4.78
17
-26.7
0.07
17
-26.7
1.87
17
-26.7
8.01
0.56
10.71
27.25
0.22
5.54
24.20
Main body Material
Diameter [mm]
Thickness [mm]
Branch
Material
Diameter [mm]
Thickness [mm]
Number of cycles
Max. operating pressure
[N/mm²]
Max. operating temperature
[°C]
ΔT at Pmin [°C/min]
ΔT at Pmax [°C/min]
Cumulative damage EN
[%]
Cumulative damage TRD301
[%]
When comparing the cumulative damages, results are as follows:
TRD 301
EN 12952-3
Total drum damage
38,52%
6,51% (set through nozzles),i.e.
Total SHT damage
29,96%
9,95% (set on, weld root machined)
Knowing that the maximum cumulated damage differ from one code to the other, TRD 301
results should be multiplied by a factor 2 to compare percentages of fatigue exhaustion.
This analysis confirms that the most important improvement is on the Stress Induced Factors.
Euro Norm appears to be less conservative than TRD 301, and more representative of physics.
8. Superheater quenching issue not accounted by the Euro Norm
Mechanical stress is high for HP drum, whilst thermal stress is high for superheater and
reheater, especially in case of water quenching. This quenching phenomenon occurs due to
condensate water inside still hot superheater during initial gas purge. This quickly chills out
parts of the component compared to others, and it induces low cycling fatigue cycle in the
overall warm start-up cycle. Although the phenomenon is very damageable, it is very difficult
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to account because of its empirical nature. Quenching issue is still not yet considered by Euro
Norm, and a good engineering practice shall prevail for a large and multiple drainage system.
9. Optimisation of drum design
There are 4 materials often used for drum design, namely SA 302 Gr B, SA 299, SA 516 Gr
60, SA 516 Gr 70. As one can see from their MAS curves, there is a trade off between
(mainly) SA 302 Gr B and SA 299 around 100 bar as per CMI consideration, where a drum
made out of SA 299 would become far too thick compared to drum made out of SA 302 Gr B.
This material is generally used for HP drums in order to reduce wall thickness (Fig. 4). The
two other materials are mainly used for lower pressure drums, where fatigue is not an issue
(low pressure, low thickness). While performing the EuroNorm calculation with the same
inputs, but with the drum thickness as a free parameter, it is interesting to note that it features
an optimum in the allowable temperature gradient (Fig. 9), which can be explained as follows.
Considering the temperature profile generated over the drum thickness, thermal stress on the
inner shell increases roughly as the square of the wall thickness, while mechanical stress
decreases proportionally according to this thickness. Therefore, during cold start-up, the inner
thermal stress quickly becomes predominant, and consequently it reduces the allowable
temperature gradient. Coincidentally, the calculated wall thickness according to ASME code
is typically close to this optimum wall thickness calculated according EN for material fatigue.
Fig. 9 Optimum HP drum wall thickness for a typical case
If the main body diameter is of importance, the nozzle diameter is not to be set aside. Whilst
considering a two downcomer drum, and comparing it to a 3 or 4 downcomer equivalent drum
(meaning that circulation/speed criteria are conserved), calculation results below show that
the drum equipped with thicker downcomers has a better fatigue resistance:
Branch diameter [mm]
Branch thickness [mm]
Total damage as per EN
2 downcomers
(base case)
641,04
95
4,95%
3 downcomers
4 downcomers
526,52
80
6,51%
475,42
75
7,24%
However, this statement is not to be generalized. Indeed, when performing the same kind of
calculation with header and tube connection, the opposite could be noticed.
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10. Welds configuration at main body to branch on drum or superheater
Three design attachments are typically used in boiler industry, each using Stress Induced
Factors (SIF) as specified by the Euro Norm as follows:
Nozzle set through
Root not machined
Root machined
αm = 0,9 αm theory
αm = 1,6 αm theory
αm = 1 αm theory
Fig. 10. Various attachment configurations of nozzles on main body
It is therefore important to design the welding connections as per the most appropriate way.
CMI standard design is as follows:
- on drums : set-through nozzle, implemented for years (with/without internal extension)
- on superheater headers : set-on nozzles/branches without root weld reappropriation
whenever fatigue analysis allows it, and with ground over (i.e. stubs) when required by
fatigue analysis.
Note that attachments connections that are using reinforcing pads cannot be calculated whilst
using EN formulas directly. Then, a finite element analysis is required by Euro Norm.
For HP superheater and reheater made in alloyed steel and subject to water quenching, CMI
standard for tube to header welding attachments is always set-on in full strength penetration
welding (Fig. 11). This design is beyond ASME minimum requirements. But, for economizer
and evaporator, partial penetration welds (Fig. 12) according ASME is proven under cycling.
Fig. 11 Partial penetration welding detail
Fig. 12 Full penetration welding detail
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11. Optimisation of boiler start-up time
As noted earlier (Fig. 6), the allowable temperature gradient increases as pressure is building.
These calculated temperature/time gradients are converted into pressure/time gradients as
those are more accessible and controllable parameters during transients. This feature is used
to optimized the start-up by applying progressive pressure ramp rates (Fig. 14), which allows
optimisation of the overall boiler start-up time, without consuming any extra lifetime of the
boiler. Such progressive pressure gradients are always implement into the plant DCS by CMI,
either via a curve of variable controlled set points applied on the HP steam turbine by-pass
valve (Fig. 13), or via various HP pressure control set points during start-up. As long as the
condenser is not yet ready, the boiler relies only superheater start-up vents sized ad hoc. Only
the HP drum pressure needs to be controlled, and no limitations are imposed on IP and LP
circuits because those are less critical items for fatigue. However, the bottle neck of the
overall cold start-up time is typically on steam turbine side, rather than on boiler side.
7
6
FSR PT001A/01B
PRESS GRADIENT (BAR/MIN)
SETPOINT FOR POSITIVE HP DRUM
8
5
4
3
2
1
0
0
20
40
60
80
100
120
140
H P D R U M P R E S S U R E (B A R G )
F S R P T 0 0 1 A /0 1 B
Fig.13 Variable HP pressure control set point during start-up
Fig. 14 Start-up curve simulation based on 3 progressive start-up ramp rates, as performed by CMI
for a vertical HRSG 3 pressures behind gas turbine Alstom 13E2 from a cold condition.
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12. Conclusions
It is good to be reminded that each new component avails a fatigue life time, and fatigue
damage are always cumulative and cannot be reversed. Our main conclusions are as follows:
1. Compared to TRD 301, the Euro Norm 12952-3 improves methodology, but it is still
conservative. To stick to reality, stress concentration factors could be determined by
finite element analysis and used with the Euro Norm, as it is allowed by the code.
2. Branch to main components shall be engineered so as to minimize fatigue. This means
that set-through arrangements are preferred for drums, whereas set-on are preferred for
tube to headers arrangements.
3. Euro Norm shows that allowable pressure gradient varies with pressure. Then, boiler
start-up time can be optimized without consuming fatigue lifetime by using
progressive pressure gradients. This interesting feature, derived directly from a fatigue
analysis, is used by CMI on a standard basis.
4. If possible, exact alternate stress values shall be obtained either by finite elements, or
finite difference calculations, in order to have more representative results.
5. The sensitivity on allowable number of start, due to the logarithmic and probabilistic
natures of fatigue, results in an inherent uncertainty in fatigue lifetime analysis.
6. The superheater water quenching, which remains the big cycling issue for hot start-up,
especially for horizontal HRSGs, is not accounted by the Euro Norm. According to
good engineering practice, efficient HP superheater and reheater drainage remains the
key design point to overcome this quenching problem.
13. Acknowledgments
A special thank to Mr Brueckner from Siemens for the useful information on the subject, as
well as to Mrs Weber from CMI for her contribution for those fatigue calculations.
14. References
[1] Euro Norm NBN EN 12952, Part 3, February 2002 ‘Water-tube boilers and auxiliary
installations, design and calculation for pressure parts’
[2] David S. Moelling, P.E., Frank L.Berte, Ph.D.; Peter S. Jackson, P.E., Ph.D.; Tetra
Engineering Group, Inc., Power Gen 2002 Brussels ‘Cycling Experience of Large Heat
Recovery Steam Generators in the New England ISO Market’
[3] Dooley R.B., Todd A.K., Mc Naughton W., Paterson S.R., Pearson M., Shields K.J;
2002; EPRI, HRSGs Tube Failure
[4] Leafton S.A.; Besuner P.M.; Grimsrud G.P. Aptech Engineering Services Inc; Platts
Power Magazine; December 2002; ‘The real cost of cycling power plants’
[5] ABMA, Task Group on Cyclic Service, 2002 ‘Comparison of fatigue assessment
techniques for heat recovery steam generators’
[6] Michael Pearson, Power, February 1997, ‘Warning: cycling HRSGs can be dangerous
to your health’
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OMMI, 2008, Volume 5, Issue 1, May
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[7] ASME Boiler and Pressure Vessel Code Section I and Section VIII, Rules for
construction of Pressure Vessels, ASME, New York , 2001 Edition , 2002 Addenda
[8] Claxton K.J., Collier J.G., Ward J.A., November 1972 ‘H.T.F.S. correlations for two
phase pressure drop and void fraction in tubes (H.T.F.S. design report 28).
[9] Fontaine P., June 2003, ‘Cycling tolerance – Natural circulation vertical HRSGs’,
ASME paper IJPGC 2003 - 40114
[10] Galopin J-F, 1996, ‘Dynamic Constraints on HRSG drum design’
[11] Robert Swanekamp, Platts Power Magazine, October 2002, ‘Users group to publish
guidelines for operation, maintenance of HRSGs’
[12] TRD 301 Code, April 1979, ‘Zylinderschalen unter innerem Uberdruck’
[13] Michael Pearson, P Eng, J Michael Pearson & Associates Co Ltd. And Robert W
Anderson, Florida Power Corp, Power Magazine, July/August 2000, ‘HRSGs – Questions
about condensate quenching, prestart purging’.
© The copyright of this paper is retained by CMI Energy, Belgium.
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