Design of Small Centrifugal Compressors Performance Test Faciltity

8/10/2019 Design of Small Centrifugal Compressors Performance Test Faciltity

1/9

DRAFT

Proceedings of ASME Turbo Expo 2004

Power for Land

Sea, and Air

June

14-17,

2004

Vienna

Austria

GT2004-53821

DESIGN OF SMALL CENTRIFUGAL COMPRESSORS PERFORMANCE TEST

FACILITY

Al-Sulaiman

, Fahd A.

Graduate Assistant

AI-Qutub

Amro M

Associate Professor

Mechanical

Engineering Department

King Fahd University of Petroleum

and Minerals

Dhahran, 31261, Saudi Arabia

ABSTRACT

Actual performance testing is a key element in the design

stage, development and troubleshooting of centrifugal

compressors. The present work discusses the procedure for

designing the experimental setup and the selection of drive unit

for variable centrifugal compressors sizes. It starts with setting

criteria of selection. A survey over different types of drive units

and facility setup was conducted. It was found that the electric

drive unit with the aid of transmission for stepping-up the

speeds is the most suitable type. This is due mainly to the

excellent control property of electric motors allowing for wide

range of operational speed and power. A new methodology was

developed for selecting operational power and speeds of the

drive unit for different sizes of impellers. The code, used for the

analysis, was developed by the authors. It calculates the range

of input power, input torque, and rotational speeds, as well as,

the mass flow rate, total pressure and temperature ratios for

different sizes of impellers. This will aid in selecting the proper

instrumentation for the experiments. The code was validated

with experimental results in the literature. It is expected that the

present methodology will enhance selection procedure for

designing compressor test facility.

INTRODUCTION

Improper selection of the instrumentation for a single test

rig may

lead to inaccurate reading increase

cost or even failure

in the system, such as the failure of the mass flow measurement

at the venture nozzle, Colantuoni and Colella [I]. In other

words, estimating the input power, torque, mass flow rate, rpm,

and total pressure and temperature ratios for different sizes and

configurations is essential in planning for compressor testing.

Some commercial codes are available and designed to

estimate the operating conditions of the centrifugal

compressors. However,

most

of them are for industrial

applications and used to evaluate conditions during operation

and not for the designing purposes, such as, Centrifugal

Compressor Tracking Program developed by Ronald. P.

Lapina. On the other hand, there are limited codes that can

calculate the operating conditions of the compressor for

designing purposes. An example of that is the codes developed

by PCA Engineering and Concepts Incs. However, these codes

do not show on the same figure the effect of changing impeller

size or configuration. The present work involves the

development of a code that predicts centrifugal compressor

performance at different sizes and configurations. Presentation

of results is designed such as the selection of drive unit

characteristics can be done in a very efficient manner.

Designing a flexible centrifugal compressor test facility

where different compressor configurations and sizes can be

tested with minor modifications, means reduction in cost and

time. Also, the quality of the experimental results depends on

the design of the test facility. That is the instrumentation used

for the measurement and the layout of the test rig. This raises

the need to find a methodology to satisfy these requirements.

On the other hand, one of the

main requirements in

designing the test facility

is selecting the proper type of driver

unit. For instance, in the present

study, the driver has to be

flexible to test different types of compressors with good control

system

on the rotational speed

, especially near the surge

condition.

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Based on literature survey there are three major types of

drivers

electrical motor

gas turbine and blow

-down facility

with a turbocharger

Each one of them has its advantages and

disadvantages, such as cost

size, and power requirement as

discussed below

The present work evaluates these systems as

alternatives for driving different sizes of high- speed small-size

centrifugal compressors.

NOMENCLATURE

Symbols

bImpeller wdth

CAbsolute

elocity of the impeller

coTangential

velocity of the impeller

Cr

Constant

pressure specific heat

M Mach number

m Mass flow rate

N h

Number of

the blades in the impeller

PPressure

Pr Pressure ratio

PwPower

RGas constant

r Impeller radius

r p m

Rotational speed per minute

T Temperature

U Impeller blade

velocity

WRadial relative

velocity

of the impeller

Z The reference size

Greek letters

Subscripts

Impeller angle

Density

Specific heat ratio

Slip factor

Efficiency

Flow factor

Angular velocity

act

Actual case

cr Compressor rotor

c Compressor stage

h Valueat hub

ideal Ideal case

01 Impeller inlet (compressor) stagnation condition

02 Impeller exit stagnation condition

04 Compressor exit tagnation condition

I Impeller inlet

2 Impeller exit

4 Compressor exit

PERFORMANCE

ANALYSIS

One of the main requirements of designing the centrifugal

compressor test facility is estimating the required input power

and rotational speed to select the proper drive unit. This can be

achieved through

using

the proper analysis. In addition to that,

input torque , mass flow rate, stagnation pressure and

temperature ratios can help in estimating the range of operating

conditions and selection of proper instrumentation. Based on

the analysis below a computer program was constructed to

perform the calculations for different impeller sizes and

configurations.

Analysis

The operating conditions for a given centrifugal

compressor can be obtained theoretically through using

reasonable assumptions

as well as available information from

the literature. Equations for input power, input torque, mass

flow rate

stagnation pressure and temperature ratios, as well as

some other related equations were derived. The main derived

equations are shown below. Most of the analysis, with

additional modifications, is based on Hill and Peterson [2]

derivation for centrifugal compressor.

Geometry

The geometry of the analyzed impeller is shown in Figs. I

and 2.

Major

Assumptions:

Ideal gas

Bulk flow,

No preswirl

at the inlet

Adiabatic flow.

Fig. 1: Impeller inlet geome try and velocity vector

I-V

Fig. 2: Impeller exit geometry and velocity vector

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Governing Equations:

Actual mass flow rate

To account for the bulk flow assumption in the analysis an

estimated coefficient for the average mass flow rate was

implemented in the code

It was chosen based on a well-

established experimental investigation in literature for smooth

pipes Schlichting

[3]. Also,

to account for the flow separation

(off design condition

)

an estimated relative flow angle, 10

relative to inducer tip as flow off design limits, was

implemented.

The flow factor is used to account for

the off-design

condition

When the relative flow is parallel to inducer tip the

flow is considered to be in ideal case

design condition);

otherw ise it is off design condition.

Using

the definition of the flow factor

the actual mass

flow rate can be written as:

Mact =

1

m

deal

In this case, the inlet axial velocity is:

tan /3,

1 )

2 )

Using the definitions of the static isentropic relation, Mach

number, total temperature and inlet axial velocity, the mass

flow can be written as a function of operating conditions and

impeller geometry such that:

m-,=P01

( 0,Qri2

Ytan 31

2

0 O

tan /31

yRTo1- )

x 7 c r , 2

2) 0,

rI) 3

as =(1-2

N b

cosf2

X(1-

M

ace tan

t 2)

m

2P2 7r2 r2 b( )

30

The slip factor is also defined as:

U

Impeller exit static density

4 . a )

4 . b )

To calculate the slip factor, the impeller exit static density

needs to be found first. The general procedure to derive the

required equation to find the density was done by Hill and

Peterson [2]. Some improvements were done by taking into

account actual inlet and exit mach number, as well as off design

condition as shown below. This procedure includes the

following assumptions:

Neglect the thickness of the vanes,

Neglect the boundary layer displacement thickness,

Assume the isentropic efficiency of the rotor

: 7 er x 1 277, , 17

,

is the stage isentropic efficiency.

Using the

isentropic relation the static

density ratio can be

written as:

P2 Po2

P Poi

[1 M

1+Y] M

L2

Y - l

5 )

In the case

of no preswirl condition

the stagnation pressure

ratio can be written in terms

of the blade velocity at the

impeller tip and

the absolute tangential velocity at the impeller

exit. This can be achieved by using the isentropic

relation and

the definition of the power,

as will as the rotor efficiency.

c

There are some formulas that can be used to calculate the

P 02

141 77,

U2

Fez

1 r-I

The oneli factor which was discussed by Wiesner [4] and

6

modified by Hill and Peterson [

2], is a representative formula

Poi

2

r

RTo

U2

since it has been validated experimentally. It takes into

consideration number

of the impeller blades and the exit

condition and configuration. Using their formula and

introducing the definition of the exit radial relative velocity

yields:

The rotor efficiency

is modeled from the experimental

results done by Krain et al

. [5]. These

results were selected

since they are close to the operating conditions

of the desired

test facility of the compressor Also, because Krain et al. [5]

study

was chosen as a reference to validate

four different

advanced codes which indicates the confident in their results

see Eisenlohr et al. [6 ]. The chosen eff iciency

model for

validation was:

77,. = 0.8545*exp -IE-06*rpm)

C e2

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This efficiency

equation will be introduced in the code to

account for the stage loses

so the total stage pressure and

temperature ratios can be obtained

.

The main purpose of

pressure ratio validation is to check whether the code can

predict the compressor performance accurately as shown later.

Mach number at the impeller inlet can be written as:

M, _

C

2(7

C, )

YR T01-2C

r

Substituting

for C,

in term of the angular velocity and the

inlet angle and then square the Mach numb er leads to:

yR

2

f .

tan /3,

z

0, Qr.

t a n 1 3

o

2C

r

Similarly, the square exit Mach number can be written as:

z

M

2

=

C2

C

YR (T02 - )

n

8 )

9 )

where

C, is the absolute velocity at

the impeller exit as

illustrated

in Fig. 2.

It can be written as:

2

C2= U2-

Wr

2

ttanQ2)_

Wr

1 0 )

The derived Eqs., 4 and 6, can be substituted into Eq. 10 to

obtain the density at the impeller exit.

2

y ) RT

o

x

1t^^

^

Input Power

Assuming there is no presw irl at the inlet, CO

, is equal to

zero so the input pow er can be written as:

Pw=

y

., 6sU22

1 2 )

Substituting into Eq. 12 for the slip factor and write blade

velocity at the impeller tip as a function of the angular velocity,

the power can be obtained as:

Pw =M'

1- cosfl2)

b

z

x 1

mom, m

tan/i2)(2r2)

2P2 n2 r2 b(0 )

IMPLEMENTATIONS

OF THE ANALYSIS

1 3 )

A computer code was developed

, using MATLAB

software

to calculate compressor performance parameters.

Figure 3 is a flow chart for the code. The input data are:

Impeller radiuses,

Impeller configuration (angles and number of blades),

Inlet conditions, total pressure and temperature, as

well as flow factor,

Gas properties ideal gas only).

Three types of output are configured to present the

compressor performance as the following:

The first type produces plots for input power, input torque.

total pressure and temperature ratios, separately vs. mass flow

rate for different rpm. This will serve in validating the code.

The second type produces plots for input power, input

torque,

total pressure and temperature ratios and mass flow rate.

separately, vs. rpm for different impeller sizes and exit

tangential

velocities. This will aid in selecting proper

instrumentation range and determine

the required power input

for a given impeller geometry.

The third type produces plots of the total pressure ratio vs.

impeller size for different input power and rpm. This output is

the distinguish one which will aid in designing compressor for

a given system, as well as aid in the design of a drive unit for

different impeller sizes and configurations. Further details are

given in the example set in results and discussion section.

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Input

geometry

, size, gas properties,

inlet stagn ation

conditions, efficiency, flow factor, rpm, bulk flow correction.

p2 = pt , as a first iteration for p2

Calculate: slip factor

, W,2. C2, C 02 ,

Power

Torque T a Mi Mt

efficiency over the flow range for a given rpm. Figure 4

illustrates

clearly the validation of the efficiency model used in

the present code for the given case. Input to the code is the

same as provided in the literature. The impeller data are shown

in Fig. 5. Figure 5 is a comparison between the experimental

results and the present code output for total pressure ratio. It is

clear that the present code predicts total pressure ratio very well

up to the running speed of 40,000 rpm with negligible error,

specifically near the maximum efficiency. At higher rpm the

present code over predicts the total pressure ratio. The

deviation increases with rpm and as operating condition

deviates from maximum efficiency point. However, the

maximum deviation

is less

than 11% at 50, 000 rpm at the

highest efficiency. Surge cannot be predicted by the present

code, which is the usual case for many codes.

K r a i n e t a l [ 6 ]

0.85

Calculate: p2rew

u

0.83

N d

V

0.81

CL m

0.79

- - - -1

U

N

0.77

0.75

25,000

35,000

rpm

, 0.07 s.-G.tt2r.-0.93

63se.

q.. -310.8

0-36 b. 010T

Fig. 4: Efficiency

modeling validation.

If U2> 6 00, break

the loop rpm loop)

Calculate final

Power

Torque

, Pa/Pui, T(Mrrn,

3

Show

the output

Fig. 3: General flow chart of the code.

RESULTS AND DISCUSSION OF THE ANALYSIS

Code validation

The present code and analysis are to be validated through

comparison with experimental results of Krain et al. [5]. The

validation of the code is based on the comparison of the total

pressure ratio. So, efficiency model of the compressor must be

first validated with experimental results of Krain et al. [5].

Experimental results shown in Fig. 4 are based on the average

Efficiency m odel

45,000 55.000

sower;'' ,

116 . 1S2

_2S3

Anu6wr

0* )

Fig. 5: Code validation,

total pressure

r a t io vs .

flow rate for different

rotational speeds

Example

Dimensions

of the impeller given in

table I are for the

reference

size in the example

. A multiplication factor, Z, is

used to

alter the dimensions

of the reference impeller without

effecting

geometrical angles and number

of blades.

The working fluid is

air with inlet total pressure and

temperature

of 101.3 kPa and 27 C, respectively. Both high

and low mass

flow rate impellers are considered

each as a case

as shown in table

I (,(3l = 45 and 67). Figures 6 to 9 are the

first type of output for

the present

code for the basic dimension

of impeller (Z=1). This type of output is useful for code

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validation with experimental work for further development and

modification. It can also aid in the selection of instrumentation

for a given impeller, as well as predicting the required drive

power and speed range. In addition, it shows the effect of

changing impeller configuration, such as impeller inlet angle.

The second type of output is illustrated in Figs. 10-14.

Here,

mass

flow rate, the input power and torque, as well as

total temperature and pressure ratios are provided for different

impeller sizes vs. rpm on the same plot. Moreover, impellers tip

speeds are indicated. The information provided will indicate the

set of instrumentation required for the whole test rig. The first

design constrain is the available input power (say 500 kW, for

example). This will provide the limit for the impellers tip

speed. For example, if the tested impeller is of a 3Z size, this

will lead to a tip speed limit of about 530 m/s at (0, =1). The

second design constrains is the maximum output rpm for the

drive unit (say 100,000 rpm as an example). This would also

put a limit for tip speed of small impellers as shown in Fig. 10.

The third type of output is illustrated in Figs. 15. a and b.

The figures indicate the total pressure ratio vs. impeller size for

two different angles

. Iso-lines of input power and rpm provide a

useful tool not only for experimental drive unit setup but also

for selecting impeller size for a given compression system or

gas turbine engine. For example, if the power limit is 200 kW,

speed limit is 100,000 rpm and inlet impeller angle is 67, then

the maximum total pressure ratio that can be tested is 2.9 for

impeller size of (Z= 3) due to power limit. Also, the maximum

total pressure ratio for an impeller size of (Z=1) is 2.8 due to

speed limit. However, the maximum possible total pressure

ratio that can be obtained for 200 kW at 100,000 rpm is 8.2

with impeller size corresponding to (Z= 1.56). The same

outputs of the code may be obtained for any configuration of

impeller for the design purpose of a test facility.

Table 1: Impeller characteristics of the reference size Z

Impeller characteristics Value

Inlet im pe

l ler raduis r , m)

0.025

H u

p

im p

eller raduis ri, (m)

0.008

Exit impeller raduis

r2 m)

0.0381

Im el ler width

m ) 0.005

Im eller inlet an gle deg)

67,45

Im

p

eller exit angle (deg)

25

Impeller blades

number

7+ 7

a t

o

0.05... 0.1015..... 02...025 03... 035 04

A6W.1 mess Aow rep tkgfs)

Fig. 6: Input power vs. actual mass flow rate

for different rpm and a given size, Z=1.

00005 0 04602025

Acted mesa now W e tkp0 s )

i39R rt ka^

o _ _ _ _

03504

Fig. 7: Input torque vs. actual mass flow rate for

different

rpm and a given

size, Z=1.

+flrn k,afi

0 050t 0

t5 02 .._..0.25

Ac tual m.s1 flow rate l 51)

0 .3 0 . 3 5 04

Fig. 8: Stagnation temperature ratio vs. actual mass

flow rate

for different

rpm and a given

size, Z=1.

6

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1I .. 60k rpm

12

... _., .... _.,._.

00060AOti 02026. 0

6OJ6Oa

M1* maea now rue ftW

Fig. 9: Stagnation

pressure ratio

vs. actual mass flow rate

for different rpm and a given size, Z=1.

U -000.M

7 ..............., .... _...

2762

:,- 32

5w

2 757

6 6 0 x , . ^ .

5

a l e _ 2 262

4 6 Q,

2 2

62

1.7

z M o .

.

%W tb

2

.7 15z

I

Fig. 10: Input power vs. rpm for different

impeller sizes

and exit tangential velocities.

120-

100.

60 .

252

GO 400 ..-1 2165 ..

2Z

a

1762

20 ti 152

Zk m wp.[.

1la.

t26Z

Z

TSZ

00O1Y6. 2

I n

Fig. 11: Input torque

vs. rpm for different

impeller sizes and

impeller exit tangential velocities.

25

10

Fig. 13: Stagnation pressure ratio vs. rpm for different impeller

sizes and impeller exit tangential velocities.

0

0611622S

151510

Fig. 14: Actual

mass

flow rate vs. rpm for different impeller

sizes and impeller exit tangential

velocities.

^,na.i.,.u.12r

Fig. 15.a: Pressure ratio versus impeller size for different

rotational speed and input power where 6, = 670.

Fig. 15.b

Pressure ratio versus im peller size for different

Fg 12: Stagnation Temperature ratiovs rpmfor different o

impeller sizes and impeller exit tangential velocities. rotational speed and input power where

,5I = 45.

7

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DRIVER UNIT SELECTION

The type of driver unit used to drive a compressor has to

satisfy many criteria. For instance, the driver has to be flexible

to test different types of compressors with good, control system

on the rotational speed, especially near the surge line. In

general, the driver has to be:

Flexible (

can accumulates

several sizes of compressor),

Reliable,

Excellent control on the rpm and torque,

Safe,

Easy to use,

Environment friendly,

Economical.

Comparison Between The Electric Motor,

The Combustion

Gas Turbine And The Blow

-

Down Facility Wth A

Turbocharger

The choice of an electric motor type depends on the

application. One of the examples of the electric motor selection

criteria was discussed by Clarkson et al. [7]. In case of selection

a motor driver for driving a high-speed centrifugal compressor,

over 30,000 rpm, usually a gearbox is used since majority of

motors in the market cannot reach a very high speed, especially

in the case of high power requirements (>250 hp). Recently,

considerable efforts are made to

overcome this problem for the

case

of the low power, Soong et al. [8] and Yuri and Smith, [9].

Most of these motors

are still under development phase, so

availability and reliability is questionable at the present. Most

of the

experimental

investigations for centrifugal compressors

with high rotational speed and power are accompany with a

gearbox, Kim et al. [101.

As a drive unit for centrifugal compressors,

gas turbines

have many drawbacks compared to electric motors. In general,

the efficiency of the gas turbine is low, except at its design

operation point. Normally, the gas turbine is affected by the

change in the ambient temperature, operation at partial loads,

filtration of inlet air and blade fouling of the compressor,

Saxena [11]. It is also difficult to control the turbine especially

if it is operating at off-design condition. Actually, the control

unit of the turbine engine alone needs speed, temperature, flame

detection and vibration inputs. These inputs are important to

control the turbine engine during startup and shutdown, steady

state operation and for the turbine protection, Boyce [12].

Moreover, gas turbines require a number of auxiliaries, such as,

the electrical starter, the main oil pump and the fuel pump. All

of these auxiliaries require control. These supplementary

devices increase the maintenance difficulty and price, as well as

complexity of control system. The flexibility to test many

different sizes of compressors is also reduced due to

performance characteristics of gas turbines. In addition, special

type of fuel and filtration need to be used to decrease air

pollution. Gas turbines are also considered to be unsafe

machine. Sometimes the available turbine needs some

development before it can be used to drive the centrifugal

compressor for the purpose of testing; see Turner et al. [13] as

an example. In their case, any considerable change in

compressor size and characteristics may require some

modifications, such as, a special gearbox. This means this type

of drive is not suitable for testing wide range of compressors

sizes and configurations.

The third possible driver is the cold air turbine, such as in

turbochargers. Blow-down facility is usually required to

provide compressed air to the turbocharger to rotate the radial

turbine that is coupled directly with the centrifugal compressor

through a common shaft. This type of experiment is not

expensive especially if the blow-down facility is available.

However, it needs some auxiliaries for the purpose of control,

which leads to larger size and more complexity in construction

and maintenance compared to the electric motor. The facility is

considered to be safe, has no negative effect on the

environment but its major drawback is the limited flexibility for

the purpose of testing different sizes of compressors.

The main objective of this study is to develop a test

facility that can be used to test different compressor sizes with a

good controlling system up to a speed limit of 100,000 rpm.

Table 2 compares the characteristics of the three driver units

discussed. From this table the selection of the electrical motor

as a driver unit is the best choice. The main reasons of that are

the simplicity of control and flexibility to test different

compressor sizes. In addition, it is easier to maintain and the

overall system is smaller compared to the other two drive

systems.

Comparison

Between

Two Electrical Motors,

With And

Without Vacuum System

There are two methods commonly used to satisfy dynamic

similarity when testing centrifugal compressors driven by

electric motors. First method includes a vacuum environment

system and the second one is without vacuum. By using the

vacuum system almost all practical conditions can be tested and

this is done mainly by simulating the pressure to match the

Reynolds and Mach numbers while maintaining low rpm.

However, this type of driving system is relatively complex. It

requires an advanced sealing system; a closed loop that needs a

heat exchanger to cool the output air from the compressor; and

flow straightener and equalizer to improve the flow quality at

the compressor inlet, Shirley [14]. Also, the test section of the

vacuumed type is more difficult to move, maintain, control

(inlet conditions) and has larger size compared to non-vacuum

one. However, it has lower power requirement for the motor

due to low rpm and vacuum. On the other hand, driving the

impeller without the vacuum system is relatively simple

compared to the vacuum assisted system. It does not need a

closed loop or advanced sealing system but requires higher

power and rpm drive unit, which might require a gearbox in

some cases Table 2 shows the characteristics of the two type of

electrical system.

To sum up, the non-vacuumed type is the most suitable to

meet the objectives of the present design. However, future

developments can be made by adding a vacuum loop to

broaden the test condition, Re and Mach numbers. The

suggested drive unit will be composed of electric motor,

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gearbox and variable speed torque converter to meet variation

of impeller size, see Fig. 16.

Table 2: Characteristics of the electric motor gas turbine and

bow down facility with a turbocharger as centrifugal

compressor driver units.

Drive

Motor no Motor

with Blow-down

vacuum

a v acu u m Gas Turb ine

facility with

Characte r is t ics

Sys.)

system

turbocharger

Cost

Expe nsive

Ex pen sive V ex pe nsive Medium

Size

Medium Medium

Medium La r ge

Specification

Isolated -La rg er E xtra

Blow-down

base - Isolated

vent i l a t ion facility

requirement

Al loca t ion

Diff icul t difficult Possible

Possible

flexibili

Energy High

Relatively

Fuel

Compressed air

re q

uirement

low

Technology S p e c i a l

Advanced

V . h i g h

V . h i g h

gearbox

scal ing S y

s

Excellent

Contro l

Difficult

at

Diff icu l t a t

rpm, power

Excellent

Difficult

flow

surge

surge

conditions)

Ma in te n a n c e

Easy

Difficult Difficult

Medium

Environmenta l

Excellent

Excellent

Poor Excellent

issue

Bi-directional

OK

OK

Special

No

g e a rbox

S a f e

t y

H i

gh

High Low High

Flexibility of

V. good

Excel lent Low

Low

compressors

sizes

C

I- Centrifugal compressor, 2- Gearbox

3- Torque converter, 4- Electric motor

Fig. 16: General facility layout of the centrifugal compressor,

CONCLUSION

A performance analysis for centrifugal Compressors of

high-pressure ratio was developed and incorporated into a

numerical code. The analysis was validated with experimental

results. A special output configuration was made for

compressor perfomance in which total pressure ratio was

plotted against both power requirements and operating speeds

for different impeller sizes. This proved to be vital tool on the

selection process of the drive unit for testing varies impeller

sizes. Moreover, the results of the analysis will guide the

selection of the instrumentation

range

. Further improvement of

the code is needed in terms of efficiency prediction.

Alternatives of drive units were studied carefully. As a result,

eclectic motor in combination with a variable speed torque

converter and a gearbox was selected to meet present

requirements. In addition a vacuum loop can be integrated with

the facility to broaden test conditions.

ACKNOWLEDGMENTS

The authors would like to acknowledge KFUPM for the

support in the preparation

of this

research.

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