 Open Access
 Total Downloads : 21
 Authors : Ritesh. B. Dhamoji, Harish Kumar, Raju. B. S , Manjunatha L H
 Paper ID : IJERTCONV3IS17015
 Volume & Issue : NCERAME – 2015 (Volume 3 – Issue 17)
 Published (First Online): 24042018
 ISSN (Online) : 22780181
 Publisher Name : IJERT
 License: This work is licensed under a Creative Commons Attribution 4.0 International License
Experimental Investigation of Static Pressure Distribution on the Flat Surface Due to Impingement of Air Jets
Ritesh. B. Dhamoji [1], Harish Kumar [2], Raju. B. S [3] , Manjunatha L H[4] [1]PG Student, School of Mechanical Engineering, REVA University, Bangalore.
[2] Assistant Professor, School of Mechanical Engineering, Reva University [3] Associate professor, school of Mechanical engineering, Reva University [4] Professor & Head, Mechanical Engineering, REVA ITM/RUAbstract: The promising technique of impingement cooling of such components needs primary attention as this technique is associated with nonuniform distribution of heat transfer coefficients. The designer sand researchers may choose suitable impingement cooling system based on the permissible nonuniformities of heat transfer rates depending on scheme of application. The present work is to address the issue of nonuniformity. This may be accomplished by obtaining the local heat transfer and fluid flow distributions due to various schemes of jet impingement cooling and there by quantify its degree of nonuniformity.
Hence, present work aims to investigate the local heat transfer and fluid flow characteristics due to impingement on flat surface which simulates leading edge of a typical gasturbine blade. Further it is aimed to investigate separately the influence of geometric parameters of jets and target surface on local distribution of heat transfer coefficients and wall static pressure distribution. Thin foil and Infrared radiometry technique used by Lytle and Webb [27] will be considered in the present study of local temperature measurement. The uncertainty analysis will be carried out for all the parameter estimation as detailed by Moffat
Key words: Air jets, Static Pressure distribution, Venturimeter

INTRODUTION
The jet impingement heat transfer is one of the well established high performance techniques for heating, cooling and drying of a surface. Such impinging flow devices allow for short flow paths on the surface with relatively high heat transfer rates .Interest and researching this topic continues unabated and may have even accelerated in recent years because of its high potential of local heat transfer enhancements. Applications of the impinging jets include drying of textile sand film; cooling of gas turbine components and the outer wall of combustors; and cooling of electronic equipment.
Single jet finds its application mostly where highly localized heating or cooling is necessary. However, when large surface areas require cooling or heating, multiple jet impingements are desirable. The proposed research work is to focus study on gas turbine blade cooling application which requires multiple jetimpingements.
The efficiency of the gasturbine engines depend primarily on the turbine in let gas temperatures. The metallurgical considerations of the gasturbine blade put a limit on the maximum in let gas temperature. Hence, an option to improve the engine efficiency could be to device an effective method to cool the turbine blades. A promising method of cooling turbine blades is to impinge cool air on the internal surfaces of the blades so that the gas turbine cycle may be operated at higher engine compression ratios with higher inlet gas temperatures for higher efficiencies and reduced fuel consumption. VanTreuren [1] reports a reduction of blade metal temperature of 40Â°C can improve blade life tenfold. Hanetal.[2] reports that turbine entry temperatures in some of the advanced gas turbines are far higher than the melting point of the blade material hence, the turbine blades need to be efficiently cooled using the relatively cool air bled from the compressor for improved performance.
Fig.1: cooling of gas turbine blade with a row of impinging jets
Hence, this region needs primary attention of efficient cooling method. The internal passage at the leading edge may be considered to have a semicircular concave surface and this region may be convectively cooled by a span wise row of impinging jets. In order to design and choose an effective cooling method, the knowledge of the local heat
transfer characteristics and wall static pressure distribution of the blade are essential.

LITERATURE REVIEW
The high heat transfer rates associated with impinging air jet is well recognized and documented for many years. Review of the experimental work on impinging jets is reported by Living ood and Hrycak [3], Martin [4], Jambunathanetal.[5] and Viskanta[6].
The available literature reveals that there have been some experimental investigations on flow and heat transfer characteristics of semicircular concave surface with arrow of impinging jets which typically simulate cooling of leading edge of gas turbine blades. One of the first investigations on the impingement of a row of circular jets on a concave surface is reported by Chuppetal.[7].Their configuration simulated cooling internal passages of leading edge of typical gasturbine blade. Jusionis [8], Metzgeretal.[9,10] and Dyban and Mazur[11] have studied the influence of Reynolds number(Re)and other geometric parameters on the average heat transfer characteristics in a semicircular concave surface impinged by single row of circular jets. Taslimet al. [1215], Taslimand Khanicheh[16] and Taslimand Bethka[17] made extensive experimental and numerical investigate on off low and heat transfer due to impingement on a smooth and ribroughened leading wall of gas turbine for constant z/d and s/d at different Reynolds number. Their study included average heat transfer characteristics for different flow conditions. They reported that the air mass flow rate through all the holes remains almost same for the cases of flow entering the supply channel from one end or both ends. Iacovidesetal.[18],studied experimentally the flow and thermal development of a row of cooling jets impinging on a rotating concave surface. Cooling fluid is injected from a row of five jet holes along the center line of the flat surface of the passage and strikes the concave surface. Craftetal [19] studied modeling of three dimensional jet array impingement and heat transfer on a concave surface. Fenotetal.[20] carried out experimental investigation of heat transfer due to arrow of air jets impinging on a concave semicylindrical surface. The jets are issued from round tubes and flow to the supply channel is normal to the concave surface. However most of the heat transfer studies are not well supported by the wall static pressure measurements on the cylindrical concave surface along both the longitudinal and circumferential directions .Tabakoff and Clevenger[21] studied the effect of surface curvature on wall static pressure distribution for slot jet impingement .They varied the ratio of diameter of flat surface to slot width between 5.0 and 20.0 at constant jettosurface distance of six times jet width. They found that, the wall static pressure decrease along the curvature at higher rate at lower ratios of diameter of flat surface to slot width. Florschuetzet al.[22]studied flow distribution characteristics for arrays of impinging jets on flat surface. They developed a theoretical model to predict
the rowbyrow flow distribution and compared their results with the experiments. Choietal[23] carried out LD A measurement of mean and fluctuating components of velocity in an experimental study with converging slot jet impinging on a concave surface. Recently, Ramakumarand Prasad [24, 25] reported experimental and computational results of the flow characteristics from multiple circular air jets impinging on a concave surface. They reported experimental results for the configuration of
D/d=30,s/d=5.4andz/d=1.0.However,their study does not explicitly correlate the influence of jettoplate distance on wall static pressure at a given Reynolds number. Bunker[26] reports about many and mostly unattended major thermal issues of turbine cooling as advanced engine design has allowed surpassing normal material temperature limits. One of the key issues includes uniformity of internal cooling of turbine blade passages.

CONCLUSIONS FROM LITERATURE SURVEY AND OBJECTIVE OF PRESENT PROPOSED
WORK:
The promising technique of impingement cooling of such components needs primary attention as this technique is associated with nonuniform distribution of heat transfer coefficients. The designer sand researchers may choose suitable impingement cooling system based on the permissible nonuniformities of heat transfer rates depending on scheme of application. The present work is to address the issue of nonuniformity. This may be accomplished by obtaining the local heat transfer and fluid flow distributions due to various schemes of jet impingement cooling and there by quantify its degree of nonuniformity.
Hence, present work aims to investigate the local heat transfer and fluid flow characteristics due to impingement on flat surface which simulates leading edge of a typical gasturbine blade. Further it is aimed to investigate separately the influence of geometric parameters of jets and target surface on local distribution of heat transfer coefficients and wall static pressure distribution. Thin foil and Infrared radiometry technique used by Lytle and Webb

will be considered in the present study of local temperature measurement. The uncertainty analysis will be carried out for all the parameter estimation as detailed by Moffat [28].


OBJECTIVES OF THE PRESENT WORK
Based on the literature review the following objectives are defined for the present work. The present work is to study experimentally the distribution of static pressure on the flat surface due to air jet impingement from long pipe circular nozzle.

Study the influence of Reynolds number of flow on the static pressure distribution on the flat surface. The experiment is conducted for the
Reynolds number of flow ranging from 12000 to
Minimum flow rate Qmin= 0.36
47000.

Study the influence of longitudinal distance(X) on the static pressure distribution on the flat surface
Maximum flow rate Qmax = 2.8

THROAT DIAMETER
(Rated by Manf.)
for various Reynolds number of flow and distance between target plate and nozzle (Z).


Study the influence of longitudinal distance from point of impingement of jet (X) on the static pressure distribution on the flat surface for various Reynolds number of flow and distance between
Throat diameter is designed to get minimum 50mm
deflection in water manometer for Qmin Flow through Venturimeter is given by Qv = a1a2 2 g ha ..eq. (5)
a12a22
For 50mm deflection of water at 35 ha= 43.63 m of air
target plate and nozzle (Z).
0.36= 5.1104a2
2 9.81 43.63
60 (5.1104)2a22

Study the influence of distance between target plate and nozzle (Z) on the static pressure

a2 = 1.90*104 m2
d22 = 4 a2 = 4 1.90 104
distribution on the flat surface for different Reynolds number and along transverse axis(Z)


DESIGN OF VENTURIMETER
Discharge through a pipe is usually measured by
d2 = 0.01556 m = 15.56 mm
Considering d2= 16mm Inlet diameter d1 = 25.4mm Throat diameter d2= 16mm

Length of manometer:
For maximum flow rate length of manometer is
providing coaxial area contraction within the pipe and by
2.8= 5.11042.01104
2 9.81 ha eq. (6)
recording the pressure drop across the contraction.
A venturimeter is a device consisting of a short length of gradual convergence and a longer length of
60 (5.1104)2(2.01104)2
ha= 2321 m of air
h = haa = 23211.146 = 2.65 m of water
gradual divergence. Semi angle of convergence is 80 to 100 and the semi angle of convergence is 30 to 50. A pressure
w w
1000
tapping is provided at a location before the convergence commences and another pressure tapping is provided at the throat section of the venturimeter. The pressure difference

Design of venturimeter:
For convergent angle 1 = 100 tan1= d1d2 = 25.416 eq.(7)
(p1p2) between the two tappings is measured by means of a l
2l1
2l1
U tube manometer. The manometer may contain water or mercury as manometric fluid depending upon the pressure
1 = 26.65 mm 27 mm For divergent angle 2 = 50
tan2= d1d2 = 25.416
difference is expected. A flow nozzle is a device in which
2l2
2l2
the contraction of area is brought by nozzle. One of the pressure tappings is provided at a distance of one diameter upstream the nozzle plate and other at the nozzle exit.
Air blower AEG GM 600E (600W, 6 Speed, 016000rpm)

Minimum pressure rise at lower speed of blower = 4mm of Hg

Maximum pressure rise at maximum speed of blower=33mm of Hg

Density of air at 350C (a)
= P = 101.325103 = 1.146 .eq. (1)
l2 = 53.721 mm 54 mm
a RT
0.287308

Pressure head ()
h = mhm
= 13.61034103 = 47.47 eq.(2)
a a
1.146

Velocity of air through Orifice (Vo)
Vo=Cd 2 g ha = 0.622 9.81 47.47 = 18.92
.eq. (3)

Flow rate (Qo)
Ao= Area of orifice, d0 = Diameter of orifice.
2 3 2 2
2 3 2 2
Ao= d0 = (20 10 ) =3.14* m …eq. (4)
Fig.5.1: 2D Venturimeter
Venturimeter Dimensions: Inlet Diameter: d1= 25.4 mm
4
Qo= AoVo =
4
3.14*104 * 18.92 = 0.356
Throat Diameter: d2= 16 mm Convergent Angle: 1=100
Consider,
Divergent Angle: 2=50


Calibration of venturimeter
Venturimeter is a device used to measure flow rates. The
mmax= 2.81.1425 = 0.053316 kg/s
60
Reynolds mber for minimum flow rate of air
basic principle on which venturiemeter works is that by nu 3
reducing the crosssectional area of the passage, a pressure
Remin=Vd1 =4mmin = 46.8810 =
d1 0.02541.983105
difference is created and measurement of pressure difference ensures the discharge through the pipe. Since
16846.94 .eq. (10)
Reynolds number for maximum flow rate of air
the crosssectional area of throat is smaller than the cross
Remax=4mmax= 40.053316
= 134775.59eq.
sectional area of the inlet section, because of which the velocity of flow at throat become greater than at the inlet section. The increase in velocity of flow at throat resists decrease in pressure at this section. The pressure difference between these two sections is determined by connecting a differential manometer between the taps provided at these sections; measurement of pressure difference enables the rate of flow to be calculated.
The actual discharge (Qact) is calculated by
d1 0.02541.983105
(11)
Venturimeter calibration with water at average temperature 300c
w= 8.315*104 kg/ms Re = Vd1 = 4mw
w d1w
1a2
1a2
a
4[ 2ghw]w
a2a2
t
t
Qact= Ah
Re=
1 2 =
d1w
IfAoandA2be the crosssectional areas of the inlet and
a
1a2
1a2
4[ 2gh
m]
thoat sections respectively and His the difference between pressure head sand g acceleration due to gravity then the theoretical discharge (Qth ) is given by
a2a2
1 2
1 2
d1w
mw w
Q = a1a2
2 g (w 1) ha..eq. (8)
d1 = Pipe diameter = 0.0254 m
th a12a22 a
The coefficient of discharge is given by
C =Qact eq. (9)
d2 = Throat diameter = 0.016 m
1
1
a = Area of pipe = 12 = 0.02542 = 5.06*104m2
4 4
d Qthe
a = Area of throat = d22 = 0.0162= 2.016*104 m2
The diameter of the inlet and outlet are measured and their crosssectional area is calculated. A differential
2 4
4[ 5.0671042.0106104
4
29.81h
13.6 ]1000
manometer is connected at the inlet and throat
4 2
4 2
m1000
(5.06710 ) (2.010610 )
sections and water is allowed to pass through the
Re =
0.02548.315104
venturiemeter. The dimensions of the collecting tank are noted, the difference between the two limbs of the manometer are noted. The flow is then varied and above procedure is repeated for different Reynolds
number, Reynolds numbers are varied from 10000 100000 trials for particular Reynolds number is done
4[2.18910416.33hm]1000
R =
R =
e 6.635105
Re = 21.5514*104*hm
hm= m
hm= m
Re
21.55104
For minimum Reynolds number Remin= 16846
and the results are tabulated. Actual and theoretical
h = ( 16846
2
) = 6.1108*103 m
values of discharge are calculated, and coefficient of
m 21.55104
discharge determined by the ratio of actual /theoretical
For maximum Reynolds number Re
= 134775
discharge values.
2
hm = ( 134775 ) = 0.3911 m
hm = ( 134775 ) = 0.3911 m
21.55104
max
Venturimeter is calibrated between Re number 10000 10000
Deflections of mercury in differential manometer are:
If Re = 10000
21.55104
21.55104
hm = .eq. (13)
h = ( Re )2= ( 10000
2
) = 2.1533*103 m = 2.1533
Fig. 5.2: venturiemeter
m 21.55104
mm
21.55104
Calculations:
Blower: – AEG GM 600E
If Re =12500
Re 2 12500 2
Minimum volume flow rate of air for given air blower
hm = (
4) = (
4) = 3.364*103 m = 3.364 mm
Qmin= 0.36 m3/min Minimum mass flow rate:
21.5510 21.5510
If Re =15000
4 4
4 4
Re 2 15000 2 3
mmin= 0.361.1425 = 6.88*103 kg/s
hm = ( ) = ( ) = 4.845*10
21.5510 21.5510
m = 4.845 mm
Maximu
60
lume flow rate of air for given air blower:
If Re=17500
m vo
R 2 17500 2
Qmax= 2.8 m3/min Maximum mass flow rate:
hm = ( e ) = (
4
4
21.5510
21.55104
) = 6.594*103 m = 6.594 mm
Trial 1 :
1.2
1.2
Reynolds number
Deflection in mm
Actual deflection in mm
Corresponding Re
Q actual
Q the.
cd
10000
2.36
2
9311.432
8.69E05
1.54E04
0.564
12500
3.604
4
13168.35
1.78E04
2.18E04
0.816
15000
5.19
5
14722.67
2.31E04
2.44E04
0.949
17500
7.06
7
17420.09
2.60E04
2.88E04
0.9
20000
9.226
9
19752.53
2.96E04
3.27E04
0.905
22500
11.67
12
22808.26
3.38E04
3.77E04
0.895
25000
14.41
14
24635.73
3.74E04
4.08E04
0.917
27500
17.44
17
27147.26
4.10E04
4.49E04
0.914
30000
20.76
21
30172.49
4.85E04
4.99E04
0.971
32500
24.36
24
32255.75
5.03E04
5.34E04
0.94
35000
28.35
27
34212.39
5.29E04
5.66E04
0.935
37500
32.43
29
35456.89
5.50E04
5.86E04
0.93
Standard deviation
0.023917451
Avg. Cd
0.9256
% Error
2.583994234
Reynolds number
Deflection in mm
Actual deflection in mm
Corresponding Re
Q actual
Q the.
cd
10000
2.36
2
9311.432
8.69E05
1.54E04
0.564
12500
3.604
4
13168.35
1.78E04
2.18E04
0.816
15000
5.19
5
14722.67
2.31E04
2.44E04
0.949
17500
7.06
7
17420.09
2.60E04
2.88E04
0.9
20000
9.226
9
19752.53
2.96E04
3.27E04
0.905
22500
11.67
12
22808.26
3.38E04
3.77E04
0.895
25000
14.41
14
24635.73
3.74E04
4.08E04
0.917
27500
17.44
17
27147.26
4.10E04
4.49E04
0.914
30000
20.76
21
30172.49
4.85E04
4.99E04
0.971
32500
24.36
24
32255.75
5.03E04
5.34E04
0.94
35000
28.35
27
34212.39
5.29E04
5.66E04
0.935
37500
32.43
29
35456.89
5.50E04
5.86E04
0.93
Standard deviation
0.023917451
Avg. Cd
0.9256
% Error
2.583994234
Table No 1: Venturimeter Calibration
Cd vs Corr. Re
Cd vs Corr. Re
power to the water bath is switched off and temperature of water is allowed to drop.During cooling, mill volt meter readings, temperature readings from the calibrated thermometer, RTD and the K type thermocouple are noted at every 50Cdecrease in thermometer reading
Fig.5.4: Thermocouple arrangement
0.949 0.9 0.9050.8950.9170.9140.971 0.94 0.935 0.93
0.949 0.9 0.9050.8950.9170.9140.971 0.94 0.935 0.93
1)Thermos flask 2)Water bath 3)Mill voltmeter 4)Tutor 5)Calibrated thermometer 6)K type thermocouple wire 7) Test tube Experimental setup for calibration of thermocouple .A linear fit is obtained (Fig—)for the variation of emf with average temperature valuesfrom95oCto25oC .And the calibration equation is T(0C) = 23.188*v+3.8439
Table No 3: Calibration of Thermocouple
Corresponding Re
Corresponding Re
10.816
0.8
0.6
0.4
0.2
0
10.816
0.8
0.6
0.4
0.2
0
Cd
Cd
Temperture
voltmeter reading
25.6
0.94
30.1
1.13
34.6
1.33
39.8
1.55
44.9
1.77
49.8
1.98
54.7
2.19
59.7
2.41
64.6
2.62
69.7
2.84
75
3.07
79.8
3.28
84.6
3.48
89.7
3.7
94.7
3.92
Temperture
voltmeter reading
25.6
0.94
30.1
1.13
34.6
1.33
39.8
1.55
44.9
1.77
49.8
1.98
54.7
2.19
59.7
2.41
64.6
2.62
69.7
2.84
75
3.07
79.8
3.28
84.6
3.48
89.7
3.7
94.7
3.92
Fig.5.3: Graph showing Cd vs Re
The thermo couple used to measure jet temperature is calibrated. The calibration procedure is as explained below. Two junctions are formed by the K type thermocouple wire. One of the junctions (cold junction) is maintained at 00C b y dipping it in a test tube enclosed in a thermos flask whose lid is provided with a hole to insert the wire in to the flask. The test tube containing mercury and water is surrounded by ice and water in equilibrium.Thecoldjunctionismaintainedatthemercurywater interface.Thelidof the thermos flask is closed tightly for minimizing the heat transfer into the flask. The other Junction (hot junction) is dipped in a water bath. Standards are in contact with the water .Calibrated RTD, thermocouple and thermometer are chosen as standards and their average readings are used to calibrate the present thermocouple. A milli volt meter is connected between the two junctions to give the emf developed in the circuit.
The arrangement is as shown in Fig.5. Initially, water in the water bath is heated to the temperature of about95oC (i.e., the temperature more than the maximum temperature obtained during the experimentation).Then, the electrical
y = 23.188x + 3.8439
RÂ² = 1
y = 23.188x + 3.8439
RÂ² = 1
2
2
Pressure above XX in the left limb = 1045 x 9.81 x Hb
..eq.(16)
100
90
80
70
60
50
40
30
20
100
90
80
70
60
50
40
30
20
2
2
Pressure above XX in the right limb = 1000 x 9.81 x Hw
..eq. (17)
3.92, 94.7
3.7, 89.7
3.48, 84.6
3.28, 79.8
3.07, 75
2.84, 69.7
2.62, 64.6
2.41, 59.7
2.19, 54.7
1.98, 49.8
1.77, 44.9
1.55, 39.8
1.33, 34.6
1.13, 30.1
0.94, 25.6
3.92, 94.7
3.7, 89.7
3.48, 84.6
3.28, 79.8
3.07, 75
2.84, 69.7
2.62, 64.6
2.41, 59.7
2.19, 54.7
1.98, 49.8
1.77, 44.9
1.55, 39.8
1.33, 34.6
1.13, 30.1
0.94, 25.6
Equating the two pressures, we get 1045 x 9.81 x Hb= 1000 x 9.81 x Hw Hw = 1.045 Hb(1)eq.(18)
temperature
temperature
When pressure head over the surface in C increased by 1mm of water, let the separation level falls by an amount Z. Then YY becomes the new separation level.
Now fall in surface level of C multiplied by Cross sectional area of bulb C must be equal to the fall in separation level multiplied by cross sectional area of the limb.
Therefore,
= Fall in seperation level a A
= x 7.854×10^3 7.143×10^3
= 90.95
Also,
Rise in surface level B= .
10
0
10
0
90.95
0
0
1
1
2
2
3
3
4
4
5
5
2
2
The pressure of 1mm of water = g h = 1000 x 9.81 x 0.001 = 9.81
voltmeter reading
voltmeter reading
Pressure above Y Y in the left limb = 1000 x 9.81[ +
Hb + 90.95
] ..eq.(19)
Pressure above Y Y in the right limb = 1000 x 9.81[ +
Hw 90.95
] + 9.81eq. (20)
Fig.5.5: Graph showing Temperature vs. mili voltmeter reading.
Equating the two pressures, we get
Magnification Factor
1000 x 9.81[ + Hb +
90.95
] = 1000 x 9.81[ + Hw
90.95
] + 9.81 eq. (21)
Solving and substituting result 1, we get Z = 14.59mm


EXPERIMENTAL SETUP AND
METHODOLOGY
Fig.5.6: Magnification Factor
3
3
Diameter of each bulb = D =95.37mm Diameter of the limb = d =10mm Density of Benzyl Alcohol = 1045
3
3
Density of Water = 1000
Cross sectional area of the Bulb =D2 = (0.09537)2 =
4 4
7.143×10^3 m2 eq. (14)
Cross sectional area of the limb =d2 = (0.01)2 =
4 4
7.854×10^3 m2 .eq. (15)
Let,
XX be the initial separation level
Hb= height of the benzyl Alcohol above XX Hw= Height of the water above XX
The schematic layout of the experimental setup is shown in Fig. —.The experimental set up consists of Air blower of 600 watts capacity (Make AEG GM 600E), 6 speed having minimum and maximum flow rate 0.36 m3/min and

m3/min. The venturimeter is designed for this flow range and throat and inlet diameters are 16 mm and 25.4 mm. The venturimeter is calibrated with water for the Reynolds No. 10000100000. Cd of venturimeteris found to be 0.92355Â±2%
The flow rate is controlled by flow control valves, the Reynolds number is set by adjusting the flow rate with the calibrated venturimeter. The temperature of the air is measured by using ktype thermocouple placed near the nozzle exit. It is calibrated with RTD and the relation between temperature and mv obtained is t=23.188v+3.8439 with RÂ² = 1.
In addition, all experiments were performed under a steady state condition so that accurate temperature data could be obtained
Fig.5.6: Schematic diagram of experimental setup
Acrylic cylinder of inner diameter 50mm and thickness 5mm is used as flat surface and static pressure difference is measured by double bulb, two fluid micromanometer using Benzyl alcohol and water as manometer fluids, with magnificaion factor 14.59.


EXPERIMENTATION
The experiment carried out for different (Z/D) ratios. The (Z/D) ratio ranging from 0.5 to 4.5
Observation table for (Z/D)=0.5
Table No:7.1
observed for all Z/D ratio. Hence further analysis is carried out for one representative Reynolds Number(Re=30000
Cp vs (X/d) for Re=30000
2
1.5
1
0.5
0
(X/D)
Cp vs (X/d) for Re=30000
2
1.5
1
0.5
0
(X/D)
Cp
Cp
0
0.06451
0.129
0.1935
0.258
0.3225
0.387
0.4516
0.5161
0.5806
0.6451
0.7096
0.7741
0.8387
0.9032
0
0.06451
0.129
0.1935
0.258
0.3225
0.387
0.4516
0.5161
0.5806
0.6451
0.7096
0.7741
0.8387
0.9032
Fig.8.1: Influence of longitudinal distance from point of impingement(X/D) on CP for
various Z/D ratio for Re=30000
1From the graph of Cpvs x/d graph, the values of Cp is higher for lower x/d ratio. They decrease gradually up to X/D=0.13 and then appreciable decrease of Cp is observed for further increase in X/D ratio. The atmospheric pressure is reached on the concave surface at X/D 1. This indicates that atmospheric condition is reached on the concave surface due to impingement of jet at a longitudinal distance equal to the diameter of jet
2
Cp
Cp
1.5
1
0.5
0
Cp
0.5 1 1.5 2 2.5 3 3.5 4 4.5
Z/D

RESULTS AND DISCUSSION
An experiment is conducted on the concave flat surface to determine coefficient of static pressure (C =p/0.5AV 2)
Fig.8.2: Influence of longitudinal distance from point of impingement(X/D) on CP for
various Z/D ratio
From the graph of CP0 (Stagnation pressure coefficient) v Z/D ratio, CP0 decreases as Z/D ratio increases up to 1. The
p j value of CP0 remains more or lessuniform in the range of
by impinging air jet by the circular straight nozzle at steady state. Experiments are conducted for different Reynolds number ranging from 12000 to 47000 for various circumferential angles (=035Â°), longitudinal distance from point of impingement of jet (X=0 14 mm), and distance between target plate and nozzle (Z=7.75124 mm). However the dimensionless distance x/d and Z/d are considered for the analysis, where d=15.5mm, is the diameter of circular straight nozzle.
From the graph of Cp v , and Cp v X/D, for varios Reynolds number, it is observed that, coefficient of
static pressure (Cp) is independent of Reynolds number as the curves overlaps to each other and same trend is
Z/D=1 to3. This may be because of target plate located within the potential core of free jet. Then the appreciable decrease ofCP0 is observed for further increase of Z/D ratio. As the distance from the nozzle increases the velocity goes on decreasing monotonically due to spreading of jet.

CONCLUSION
The effect of Circumferential angle of concave flat surface from point of impingement of air jet, the longitudinal distance X/D and distance of target surface form nozzle Z/D on Coefficient of static pressure Cp is experimentally investigated for different Reynolds number of flow at
steady state. The followings are the main conclusions that may be drawn from this study.

The static pressure distribution on the target surface due to impingement of jet is independent of Reynolds Number of flow.

The values of static pressure Coefficient Cp0 at stagnation points are higher due to higher centerline velocities at stagnation.

The values of static pressure Coefficient Cp on the concave flat surface are almost uniform up to curvature angle of 5Â°, and decrease appreciably for higher values of .

The values of static pressure Coefficient Cp on the concave flat surface are higher for lower X/d ratio, they decrease gradually up to X/D=0.13, and appreciable decrease of Cp is observed for further increase of X/D ratio.
The potential core of free jet is observed for the Z/D ratio between 1and 3. The velocity decay is minimum for this range of Z.


REFERENCES

Vantreurn K.W Impingement flow heat transfer measurements of turbine blades using a jet array, Ph D Thesis, University of Oxford, Trinity Term (1994).

Han J.C Dutta Ekkad S.V Gas turbines heat transfer and cooling technology, Taylor and Francis,(2000) pp.2.

Livingood J.N.B and Hrycak P 1970 Impingement heat transfer from turbulent air jets to flat plates.A Literature Survey, NASA Technical Memorandum NASA TM X2778.

Martin H 1977 Heat and mass transfer between impinging gas jets and solid surfaces Adv heat transfer 13, pp160.

Jambunathan K., Lai E., Moss M.A., Button B.L., 1992 A Review of Heat Transfer Data for Single Circular Jet Impingement,
Int.J.Heat Fluid Flow, 13, pp 106115

Viskanta R., 1993, Heat Transfer and Fluid Science, 6, pp 111134. [7]Chupp R.E.,Helms H.E.,Mc Fadden P.W., Brown T.R.,1969,
Evaluation of Internal Heat transfer coefficient for impingement cooled Turbine Air foils ,J Aircraft,6(3),pp 203208.

Jusionis V.J., 1970, Heat Transfer from Impinging Gas Jets on an Enclosed Concave Surfaces ;J. Aircraft, JanFeb, pp 8788.

Metzger D.E., Yamashita T., Jenkins C.W., 1969, Impinging Cooling of Concave Surfaces with line of circular Air Jets, J Engg.for power, July,pp 149158.

Metzger D.E.Baltzer R.T.,Jenkins C.W.,1972, Impingement Cooling Performance in Gas Turbine Airfoils including Effects of leading Edge Sharpness, J. Engg. For Power, July, pp 219225.

Dyban Y.P.and Mazur A.I., 1970, HeattransferSoviet Research, 2(3), pp 1520.

Taslim, M.E., Pan Y and spring S.D., 2001, An Experimental study of Impingement on Roughened Airfoil LeadingEdge walls with film Holes ASME J.Turbomach., 123, October, pp, 766773.

Taslim, M.E., Setayeshgar, L., and spring S.D., 2001. An Experimental evaluation of advanced leading Edge Impingement Cooling Concepts, ASME J.Turbomach., 123, 2, pp.147153

Taslim, M.E., Pan Y.and Bakhtari K., 2002, Experimental Racetrack Shaped Jet Impingement on a Roughened LeadingEdge Wall with Film Holes,. Paper No. GT 200230477.

Taslim, M.E., Bakhtari K., and Liu,H.,2003, Experimental and Numerical Investigation of Impingement on a Rib Roughened LeadingWall, ASME J.Turbomach.,125,pp.682691.

Taslim, M.E., and Khanicheh , A.,2006 Experimental and Numerical Study of Impingement on an Airfoil LeadingEdge with
and without Showerhead and Gill Film Holes, ASME J. Turbomach.,128,2,pp.310320.

Taslim, M.E.,and Bethka,D.,2009, Experimental and Numerical in an Airfoil LeadingEdge Cooling channel with crossflow,
ASME J.Turbomach.131,1,PP.01102117

Iacovides H.,Kounadis D.,Launder B.E.,Li J.and Xu Z., 2005, Experimental Study of the Flow and Thermal Development of a Row of Cooling Jets Impinging on a Row Rotating Concave Surfaces, ASME J. Turbomach.,127,1,pp.222229

Craft T.J, Iacovides H., and Mostafa N.A.,2008, Modeling of three dimensional jet array impingement and heat transfer on a concave surfaces,Int.J. Heat transfer on a concave surfaces, Int.J.Heat Fluid flow, 29, pp 687702.

Fenot M.,Doringnac E., Vullierme JJ.,2008, An Experimental study on hot round jets impinging a concave surface,Int. J .Heat Fluid Flow, doi:10.1016/j.ij heat fluid flow.2008.03.015.

Tabak off , W and Clevenger ,W.,1972 Gas Turbine Blade Heat Transfer Augmentation by impingement of Air Jets Having Various Configurations, J.Engg.for power,94,pp 5160

Flors chuetz L.W., Metzger D.E., and Truman .R., 1981, Jet Array Impingement with cross flowCorrelation of Stream wise Resolved Flow and Heat Transfer Distributions, NASA Contractor Report 3373.

Choi,M.,Yoo,H.S.,Yang,J.S.,and Sohn,D.K.,2000 Measurements of Impinging Jet Flow and Heat Transfer on a SemiCircular Concave Surface,Int.J. Heat Mass Transfer, 43, pp18111822.

Ramakumar .V.N.and Prasad B.V.S.S.S., 2006 Experimental and Computational Study of Multiple Circular Jets Impinging on a Concave Surface Proceedings of 33rd National and 3rd International Conferences on Fluid Mechanics and Fluid Power, IT Bombay, India.

Ramakumar B.V.N.and Prasad B.V.S.S.S., 2007, Computational Flow and Heat Transfer of a Row of Circular Jets Impinging on a concave Surface, Heat Mass Transfer (Springer Verlag),DOI 10.1007/s0023100702743

Bunker R.S., Gas turbine heat transfer: Ten remaining hot gas path challenges, J.Turbomachinery, 129(2007) 193201.

Lytle D.and Webb B.W., Air Jet Impingement Heat Transfer at low Nozzle Plate Spacings, Int.J.Heat Mass Transfer, 37(1994) 1687 1697.

Moffat R.J., Describing the Uncertainties in experimental results, Experimental Thermal and Fluid Science, 1(1988) 317.

K L Kumar, Engineering Fluid Mechanics, 8121901006, 2007 edition.