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- Corresponding Authors:
- Bardsley A

Energy and Power Management Research Group

School of Computing, Engineering and Physical Sciences

University of Central Lancashire,Preston, PR1 2HE, UK

**Tel:**+44 1772 201201

**E-mail:**[email protected]

Whitty JPM

Energy and Power Management Research Group

School of Computing, Engineering and Physical Sciences

University of Central Lancashire, Preston, PR1 2HE, UK

**E-mail:**[email protected]

**Received date:** December 09, 2014; **Accepted date:** January 26, 2015; **Published date:** February 04, 2015

**Citation:** Bardsley A, Whitty JPM, Howe J, Francis J (2015) A Review of *in-situ* Loading Conditions for MathematicalModeling of Asymmetric Wind Turbine Blades. J Fundam Renewable Energy Appl 5:153. doi:10.4172/2090-4541.1000153

**Copyright:** © 2015 Bardsley A, et al. This is an open-access article distributed under the terms of the Creative Commons Attribution License, which permits unrestricted use, distribution, and reproduction in any medium, provided the original author and source are credited.

**Visit for more related articles at** Journal of Fundamentals of Renewable Energy and Applications

This paper reviews generalized solutions to the classical beam moment equation for solving the deflexion and strain fields of composite wind turbine blades. A generalized moment functional is presented to effectively model the moment at any point on a blade/beam utilizing in-situ load cases. Models assume that the components are constructed from inplane quasi-isotropic composite materials of an overall elastic modulus of 42 GPa. Exact solutions for the displacement and strains for an adjusted aerofoil to that presented in the literature and compared with another defined by the Joukowski transform. Models without stiffening ribs resulted in deflexions of the blades which exceeded the generally acceptable design code criteria. Each of the models developed were rigorously validated via numerical (Runge-Kutta) solutions of an identical differential equation used to derive the analytical models presented. The results obtained from the robust design codes, written in the open source Computer Aided Software (CAS) Maxima, are shown to be congruent with simulations using the ANSYS commercial finite element (FE) codes as well as experimental data. One major implication of the theoretical treatment is that these solutions can now be used in design codes to maximize the strength of analogues components, used in aerospace and most notably renewable energy sectors, while significantly reducing their weight and hence cost. The most realistic in-situ loading conditions for a dynamic blade and stationary blade are presented which are shown to be unique to the blade optimal tip speed ratio, blade dimensions and wind speed.

Wind turbine blades; Mathematical modeling

In the wind turbine industry it is paramount that components
are designed to obtain the greatest obtainable efficiency. As time
has progressed selecting the correct materials and aerofoil shape has
become predominant in the determination of the optimum blade
tip-speed ratio, thus increasing a wind turbines power coefficient,
(C_{p}) and overall turbine efficiency (ƞ_{total}). It is in the industry's best
interest to produce high performance turbine components at the
lowest costs, without compromising the turbines structural stability.
Solving problems of this ilk drive wind turbine design, research and
development forward and one such problem is the subject of this paper.

There is currently a surprising lack of computational models in literature, which accurately predict the deflexion and strain fields of wind turbine blades. However, it has proven from relatively simple physical tests on the blades that relatively straightforward and well developed underpinning theory can be used to predict the deflexion of wind turbine blades remarkably well [1]. This said, testing protocols are quite restrictive, especially for the larger expensive components. For such components therefore, effective computational predictive modeling is required to ascertain specific structural static and dynamic design limits; such models can be employed to virtually test any component in a safe and cost free environment.

The work described throughout this paper considers small scale wind turbines (5m blades) which are representative of those used on wind-turbines producing outputs of 1.5 - 50 kW; additionally these size of blades are ideal for validation experimentally [1,2]. Moreover validation of larger components becomes technically and financially prohibitive [3]. Whence, modeling the performance and the actual working limits of the blades (i.e. blades angular velocity and tip speed ratio) will be taken from real case studies evident in the literature [4].

**Previous work**

Recently [1] researchers have used the classical beam elasticity differential equation in order to effectively model the displacement of standard aerodynamic geometries. Though this work was informative in predicting the required displacement and hence strains fields remarkably well given the simplicity modeling method, these researchers do submit that further work is required particularly with respect to the use of composite materials under salient loading conditions. Moreover, conducting such investigations better predictions can be made of blade behavior when attached to the turbine. In practice it is generally accepted that the geometry of a blade is determined by the aerodynamic properties which the designer requires. The ruling characteristics are aerodynamic coefficients: chord, twist and thickness distribution. When designed blade materials are selected with the aim of minimizing tip deflexion, which reduces the probability of a blade/ tower collision [5]. It is also important that these components last the desired lifetime, around 20 years, while under accepted loads [6].

The aforementioned Whitty et al paper [1] presented closed form
solutions to the classical beam elasticity differential equation. The
developed analytical method was utilized to effectively model the
displacement of standard aerodynamic geometries used for wind turbine
blades. These models assume that the components are constructed from in-plane quasi-isotropic composite materials, forming a blade of
shell thickness with 5 mm, with or without a stiffening web. In these
cases, a composite comprising of 40 layers of laminated carbon and
glass substrate with an epoxy matrix was used, the effective modulus of
the material being 42 GPa. The authors [1] derive an explicit solution
for the strain and displacement fields for an elliptical shell formulation
which is presented for the first time in the literature. This is then
expanded to encompass the FX66-S-196 and NACA 63-621 symmetric
aerofoil approximations. However, it has been shown that the aerofoils
stated above are not symmetric (**Figure 1**). Thus this work investigates
the effects of the asymmetry on the displacement and strain fields using
comparable methodologies.

The experimental static testing of a composite turbine blade has also been carried out by a Jordanian group [2]. While focusing mainly on the manufacturing of the blade, it promotes a link between their local academic work and local industries. The works in depth look into the design and construction of the blade as well as a thorough application of static testing protocols provided valuable data in which to verify modeling work [1] and ultimately design codes [2].

**Aerofoil selection**

Since it was demonstrated that small changes in the geometry had
negligible effect on the structural response, the original Whitty et al [1]
aerofoil (**Figure 1**) used in their simulations was significantly simplified.
The aerofoil is effectively half an ellipse forming the leading edge and a
triangle forming the trailing edge. Following a survey of literature it was
found that there is normally no symmetry between the top half and the
lower half of the aerofoil, thus the Whitty et al. (adjusted) aerofoil was
designed. These are then compared with the FX66-S-196 and NACA
63-621 aerofoils which are plotted using discrete points obtained from
“Airfoil Tools” [7]. Finally the Continuous NACA aerofoil is aerofoils
constructed from the following equations, derived from the Joukowski
transform [8]:

(1)

(2)

Both equations depend on the length of chord, x, the diameter of
the aerofoil at the furthest point, d_{top} and d_{bottom}, respectively and the
total chord length, lc. The constants A, B, C, D and E are all specific
to the aerofoils aerodynamic properties [9]. Equations (1) and (2) are
effectively the same equation manipulated to produce two connected
curves at (0,0) and (0,600). These sections were plotted using the CAS
Maxima (**Figure 1**).

**Scope**

This paper provides a review of asymmetry of particular aerofoil sections for the modeling of displacement and strain fields of a wind turbine blade in service. A generalized blade moment functional which can incorporate any loading condition on the blade is also reviewed. These are employed to define the so called generalized aerodynamic load equation and different benchmarking procedures are described which should be routinely used for verification and validation of the computational models used across the industry.

**Methods**

In this section we review mathematical and simulation methods which can be employed to describe standard industrial testing protocols. These take the form of a generalized moment functional aerodynamic loading conditions, tip-speed optimization for the model blades considered and salient benchmarking procedures.

**Generalized moment equation derivation**

The most important equation which is needed to model these loading cases is the second order differential equation for the displacement of a beam (equation (3)) [10].

(3)

The differential equation comprises of two other function which vary depending on, z, the distance along the blade. M (X, z), is the generalized moment functional, this equation will depend upon the loading case being modeled. These will be discussed and explained in the sections to follow. I (z), is the second moment of area, this function also depends on, z.

To our knowledge no derivation of a generalized moment
functional for a wind-turbine blade is evident in the literature. This
derivation requires the consideration of **Figure 2** which shows a blade
fixed at one end with general load intensity applied. Taking moments
about the root renders:

(4)

From the diagram (2), there are four variables, M, the moment, W, the force, q, the load intensity and ζ, the centroid of the section of load intensity to the left of z. Additional there are spacial coordinates z and ζ. These being respectively the distance from the root and an arbitrary distance to the centroid to the left of z. consider an infinitely small section of the load distribution, dz; the force at this point is given by:

dW = q (z). dz

The total force on the component is therefore found from the integral of the distribution over the whole length:

(5)

The moment equation can be derived, considering the load distribution acting over the infinitely small distance dz, whence the moment is:

If higher order infinitesimals are neglected:

dM = q (z). z dz

Hence, the moment at the root is found by integrating over the length of the blade:

To make a general solution the moment equation must consider a changing centroid as moments are taken at points further from the root. It is important therefore that the moment equation be considered in the ζ direction. As the moment equation varies with length this creates a unique problem, when any load distribution is modelled the centroid of said load distribution will continually change the further along the blade moments are take. Hence, the well known expression for the centroid of area is employed [11].

It is now possible to substitute all the known parameters into equation (4) which gives the required generalized form:

(6)

Where z, is the distance along the blade length from the root, the function X(ζ) is the load distribution function at ζ, which is the distance along z, from the root, at which the centroid of the load distribution occurs. With the displacement field now solvable, the global strain field is required. This takes a much simpler algebraic relationship:

(7)

Equation (7) is employed in the models of this work and it is important to note that these models will only consider cross-sectional variations in the strain with material properties remaining constant.

**Force Distribution Derivations**

In reality there are a number of forces on a turbine blade that can cause deflexion. The purpose of designing an aerofoil for a wind turbine application is so that at relatively low wind speeds the lift force, produced by the aerodynamic properties of the blade, is large enough to overcome the drag force and the force on the blade due to gravity. This said, in this work the models will not consider the effect of gravity on the system. It is important to simulate the worst possible scenario when testing the blades after manufacturing; this ensures that the blade meets the design standards and is to for purpose. When considering tip deflexion the worst case scenario is to assert a point load on the blade at the end (i.e. furthest distance from the rotor axis). The magnitude of this force would be the maximum possible dictated by the particular standard under consideration [1,12,13], although this is usually unrealistic, though informative as it imposes a further factor of safety into the blade design process. When the blade is modeled, determining a suitable load is very important. The lift force, FL, can be determined by utilizing equation (8):

(8)

Here C_{L}, is the Lift coefficient, ρ, is the density of air, 1.225 kg/m3, L,
is the Blade length, L_{C}, is the chord length in meters which, due to the
tapering, depends on the blade length thus:

(9)

Where b_{r}, is the breadth at the root, and bt, is the breadth at the
tip of the blade. The parameter z here is the changing distance along
the blade. Finally the important function V^{2}_{R} , must be explained, this
is the relative velocity along the surface of the blade, depending upon
the tangential wind velocity, V_{t}. m/s, the angular velocity of the blade,
ɷ, the radius of the turbine, r, and the axial velocity, V_{θ}, equation (10):

(10)

Equation (10), can be determined from the Bernoulli momentum
equation for a free flow stream. From equations (8) and (10) a worst
case lift force can be determined and used in the models, the worst case
occurring at the highest possible tangential velocity that the turbine
could experience in its lifetime [5]. As previously indicated (**Figure 2** and equation (6) an important parameter which must be determined
for all the models is load intensity, q(z), that is, the force per unit
length; which is relent on equation (8).

(11)

where, ω(z) is the load intensity for the particular loading case (e.g.
PL, UDL, etc.), as indicated in **Table 1**.

Loading Case | Equation | Peak Load intensity, (N/mm) |
---|---|---|

DDL (PL) | (12) | 7600 |

UDL | (13) | 1900 |

RTDL | (14) | 3600 |

QDL | (15) | 5400 |

BDDL | (16) | 2000 |

BSDL | (17) | 2750 |

**Table 1:** Load distributions for specific cases.

These load distributions were developed after some key findings
during modeling. Originally the UDL distribution was considered to
be analogous to the stationary distribution on the blade, along with the
RTDL which was an approximation to the dynamic load distribution.
However, further investigation reveled that the most realistic loading
case was that of the so called Bernoulli Dynamic Distributed Load
(BDDL) and the special case^{1} of this the BDDL termed the Bernoulli
Stationary Distributed Load BSDL. It should be noted that the Point
Load (PL) is modeled here using the theory of distributions. That is,
multiplication of the Dirac distribution^{2} thus concentrating the force at
a point, rendering what is referred to in **Table 1,** as the Dirac Distributed
Load (DDL). **Figure 3**, depicts the comparisons between each of the
dierent loading cases, here application of equation (5) shows that the
total load in each case corresponds to the, required Det Norske Veritas
standard DNV-DS-J102, test load of 7600N.

Whence, in that case of the DDL, equation (12), the resulting
distribution is that of an impulse showing a zero value until the tip of the blade. These distributions (**Figure 3**) were all obtained using
the parameters given in **Table 2**. We note here for completeness that
unrealistically high speeds of 68 m/s and 110:4 m/s must be applied
in order to maintain the required factor of safety as dictated by the
standards [12,13].

Parameter | Nomenclature | Value | Units |
---|---|---|---|

Wind Speed | v_{n} |
68 | m/s |

Lift Coefficient | C_{L} |
1 | - |

Rotational speed | ω | 26.16 | rad/s |

Drag Speed | v_{θ} |
0.5 | m/s |

**Table 2: **Modeling parameters.

**Tip speed optimization**

When the initial blade selection took place, the blade dimensions were taken directly from reference [1] and then modified to give the asymmetry. The blade is then designed for the optimization of the tip speed ratio which is the most important factor. The tip speed ratio is defined by:

(18)

where ω, is the rotational speed in radians per second and U_{∞}, is the
free stream velocity. Typically accepted values for the tip speed ratio
lie between 4 - 8 [5,3]. This is the classical mean of blade design (e.g as
used in industry). From reference [4], it is shown that the optimal tip
speed ratio for a wind turbine depends on the number of blade n, and
can be calculated:

In the case utilized in this work, the turbine has 3 blades; as such
application of this expression renders an optimum tip speed ratio of
5.45. This tip speed ratio should be maintained at rated wind speed,
which for the vast majority of small scale turbines is 10 m/s [3].
Rearrangement of equation (18) and the substitution of this expression,
the rotational speed of the turbine at 10 m/s is 10.9 rad/s. As this work considers the worst case scenarios which occurs at wind speeds of
greater than 24 m/s, when the optimal tip speed ratio is maintained the
rotational speed of the turbine is increased by a factor of 2.6.^{1}

**Blade testing:**

In static testing, the blade should be loaded to its most severe loading. In accordance to the Det Norske Veritas standard DNVDS- J102, deflexion of the blade should always be less than a specified upper limit in order to avoid blade contact with the tower or other components. The design criterion for deflexion of the blade is given by the following inequality [12]:

(19)

Here, F_{k} is the characteristic load, Δd is the largest tip deflection
when passing the tower, Δi is the smallest distance from blade tip to the
tower or other obstacle in the unloaded condition, f is the load factor,
ϒ_{m} is the material factor and ϒ_{n} is the consequence of failure factor.
Typically, the load factor takes a value between 0.9 and 1.5 depending
upon the desired situation. In the case where the blade could experience
the most load (transportation and installation) the suggested value is 1.5
[13]; when under favorable loads this can be as low as 0.9. According to
the aforementioned design code the so called material safety factor is a
product of a number of other factors in the range

(0.95 -1.3). Finally the consequence of failure factor is normally taken to be 1.1 [13]. Utilizing equation (19), explicit knowledge of Classical Lamination Theory (CLT) from composite materials mechanics that composites and the maximum strain failure criterion, a maximum tip deflexion of 10% of the blade length can be determined. For the blade modeled in this work therefore the design criterion was such that the blade could not deflect more that 420 mm under the test load of 7600 N.

During static testing the applied test load is greater than the
design load to account for in influences from temperature, humidity,
production variations and other environmental aspects during the life
of the blade. The test load, F_{T}, is determined as:

(20)

where, ϒsu is the blade-to-blade variation factor considered to be 1
(worse case testing), ϒ_{f} is the load factor and F_{k} is the characteristic load
(i.e. the in-situ load). The characteristic load for this blade is ~5050N,
rendering the aforementioned test load of ~7600N [12].

**Model verification (benchmarking)**

The benchmarking case was selected to show the confidence in
the analytical model developed by comparing the results between the
benchmarking cases in Whitty et.al. [1], the ANSYS-FE model data,
this comparison is shown in **Figure 4**. Here, the benchmarking case
used the original symmetrical aerofoil, under a point load of 8500 N
(**Figure 4**). The work described in this paper slightly underestimates the deflexion when compared to the original published work. This is
due to previous work allows the chord lengths and aerofoil maximum
height to vary independently, however the current work varies these
parameters simultaneously in an attempt to maintain the aerodynamic
properties of the aerofoil.

The Whitty et.al. aerofoil [1] varies the chord length and the
height^{3} of the blade linearly using two separate functions which are
independent of each other and therefore the deflexion is greater as
the blade is modeled to be less stiff; a comparison of the maximum
deflexion values is shown in **Table 3**.

Model | Deflexion |
---|---|

Analytical (this work) | 525 mm |

Whitty et.al. | 567 mm |

ANSYS | 462 mm |

**Table 3:** Comparison of maximum deflexion under point load.

The fact that the FE method over estimates the stiffness of any
continuum system coupled with the shell element formulation
employed in the model explains the lower predictions from the ANSYS
software. The new analytical method predicts 13.6% greater deflexion
than the FE model, whereas the method employed in the reference
[1] suggests a value 22.7% greater than the FE model. These modeling
methods suggest a displacement of 567± 89 mm^{4}.

Using the new analytical model the maximum strain on the blade
occurs not at the root of the blade, as originally expected, but it occurs
at ~2250mm along the blade giving strain between about ~0.36%-
0.41% depending upon the method employed. When compared to
the results obtained in [1] (**Figure 5**) then it is clear that although the
strain is expected to occur at that position, there is also an increase
in the maximum strain prediction. Prima-facia this is unexpected as
the maximum displacement is now less due to the increased overall
stiffness (**Table 3**). However, this may be attributed to blade not being
in a fully strained condition as reported in the reference [1]. That is,
up to half the span of the blade the newly developed analytical model
is most probably stiffer than that reported previously and ispo-facto is
more optimally strained. Accordingly CLT dictates that de-lamination
should not occur until much larger strains are realized.

When comparing the strain fields, it is important to note, for all results discussed in the proceeding section, that ANSYS-FE model maximum strain values at the root are too high due to the Saint Venant edge effect and therefore a more realistic maximum strain must be interpolated.

The following results were all obtained using the parameters
displayed in **Table 2 and Table 4**, both the developed analytical and
ANSYS-FE models used these parameters.

Parameter | Nomenclature | Value | Units |
---|---|---|---|

Blade length | L | 4200 | mm |

Chord length at the root | br | 600 | mm |

Chord length at the tip | bt | 200 | mm |

Total Force | W | 7600 | N |

Shell thickness | t | 5 | mm |

**Table 4:** Model parameters.

**Displacement field results**

**Figure 6** shows the application of equations (13) through (16); the
calculated deflexion of the adjusted aerofoil under the various loading
cases (**Table 1**). As expected, the significant reduction in the aerofoils
second moment of areas due to the asymmetry of the new design,
greater deflexions are predicted by all the modeling procedures.

The maximum deflexion is observed on application of DDL, where
the total force is distributed such that the moment is at its greatest.
The lowest deflexion is realized using the SDL, where the centroid of
the load distribution is closer to the rotor axis when compared to all
others. **Table 5** illustrates the maximum deflexion under each of these
load distributions.

Loading Case | Maximum deflexion (mm) | % Deflexion |
---|---|---|

BSDL | 231 | 5.5 |

UDL | 315 | 7.5 |

BDDL | 336 | 8 |

RTDL | 462 | 11 |

QDL | 588 | 14 |

DDL | 1008 | 24 |

**Table 5: **Maximum deflexions due to loading cases.

As the loading cases change as does the centroid of the loading distribution this affects the deflexion. The model predicts that the further along the centroid of the load distribution from the rotor axis, the greater the deflexion.

**Strain field results**

Analytical model results for the strain fields are shown in **Figure
7**. As expected, the largest strains are observed on the application of a
point load, ~0.72%. The strain due to the point load, unlike the other
loading cases, shows that the maximum strain occurs 2250 mm from
the root. This was expected when consulting literature on classical
lamination [14] and previously developed tapered beam theory [1].

The maximum strain for the remainder of the loading cases all
occur at the root (**Table 6**), however there is a trend, as the centroid of
the load distribution moves further from the root of the blade the strain
field becomes increasingly non-linear.

Loading Case | Maximum Strain, % |
---|---|

BSDL | ~0.25 |

UDL | ~0.31 |

BDDL | ~0.33 |

RTDL | ~0.41 |

QDL | ~0.47 |

DDL | ~0.72 |

**Table 6:** Maximum strain values for corresponding loading cases.

**Tables 5 and 6** compare the strain fields which are generally
increase with displacement fields, as expected. These results also hold
when considering the failure over time, from [5] it is known that wind
turbine blades tend to fail due to maximum strain which occur at the
root. The only loading case contradicting this phenomenon at present
is the point load but as demonstrated previously the point load is not a
realistic in-situ loading condition.

**Stiffening web**

A single 5mm thick stiffening web, in the centre of the aerofoil,
along the full length of the blade was added (**Figure 8**) in line with the
literature [1,2] and current industrial practices. The analytical and
ANSYS_FE model comparisons including the stiffening rib are shown
in **Figure 9**. Application of a point load renders the maximum deflexion
predicted is ~630 mm and 720mm respectively for the analytical and
FE model calculations respectively.

The points in **Figure 9** show the deflexion predictions from the
ANSYS-FE UDL case and the PL case. Following this solution, in line
with previous published work [1], the second moment of area of the
stiffening web was increased using the analytical model such that the
DDL deflexion field was within the design criteria of less than 10%
deflexion, therefore reducing proportionally the delfexion the other
loading cases considered (**Figure 10**).

It was found the analytical model increases the second moment of
areas from 12.5 million mm^{4} to 15 million mm^{4} when the stiffening rib
is included (**Figure 11**).

A mathematical method for solving the deflexion and strain fields of composite wind turbine blades has been reviewed. The model makes use of a generalized moment functional (equation 6) which determines the moment at any point on a blade or beam utilizing of a in-situ load cases, e.g. equation 11). The models shown use freely available, open source software (e.g. via Maxima website: http://maxima.sourceforge.net/). Work is now progressing within the Computational Mechanics Research group at the University of Central Lancashire to allow minor adjustments to the code together with the use of external aerofoil coordinate data (from such websites as: http://airfoiltools.com). This development will enable the deflexion and strain fields of any type of aerofoil constructed from quasi-isotropic materials. This being indiscriminate, whether said aerofoil is for a wind turbine, commercial aircraft or indeed other such industrial aerodynamic component. The salient conclusions gained from the work depicted in this paper are as follows:

The asymmetric nature of the blade sections designs reduces the
second moment of area significantly from between 20.5 - 24 million
mm4 for the symmetrical aerofoil to between 9.5 - 12 million mm^{4}, the
asymmetry results in a reduction of between around 51% - 55%.

Due to the reduction of the second moment of area, the stiffness
of the blades is reduced, rendering a higher predicted deflexion. The maximum deflexion of the asymmetrical blade and their corresponding
maximum strain values obtained from the analytical model are seen in **Table 7**.

Loading Case | Maximum deflexion, (mm) | Maximum Strain, (%) |
---|---|---|

BSDL | 231 | ~0.25 |

UDL | 315 | ~0.31 |

BDDL | 336 | ~0.33 |

RTDL | 462 | ~0.41 |

QDL | 588 | ~0.47 |

DDL | 1008 | ~0.72 |

**Table 7:** Maximum deflexion and strain results for the asymmetric blade, under
individual loading cases.

In general for this and the previous work [1,2] the solid formulation predicts greater than the shell counterparts, indicating that the shell formulations are stiffer, as expected.

The most realistic loading condition for a dynamic blade is that produced by the BDDL, equation (16). Whereas the most realistic loading condition for a stationary blade is that produced by the BSDL equation seen in equation (16). It is most important to note however, the lift force, equation (8), is unique to the blade. That is the lift force is unique to the optimal rip speed ratio, blade dimensions and wind speed.

^{1}obtained by setting ω to zero.

^{2}Sometimes erroneously referred to as the Dirac-delta function.

^{3}the height being the distance for the neutral axis to the outer most fibre of the
aerofoil.

^{4}calculated at the 95% confidence level.

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Peer Reviewed Journals

International Conferences 2018-19