Performance Analysis of Double Pipe Heat Exchanger using Convergent - Divergent-Divergent Spring Turbulators

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IJIRST – International Journal for Innovative Research in Science & Technology| Volume 2 | Issue 02 | July 2015

ISSN (online): 2349-6010

All rights reserved by www.ijirst.org 224

Performance Analysis of Double Pipe Heat

Exchanger using Convergent – Divergent-

Divergent Spring Turbulators

Ayush Kumar M. Tech. Scholar

Department of Mechanical Engineering

U.I.E.T, Kurukshetra, India

Abstract

In recent years, many numerical and experimental studies on heat transfer have been discussed with different configurations.

Here CDD (convergent divergent spring tabulators) CDDSTs were placed in the inner tube of double pipe heat exchanger and

effect on heat enhancement and friction factor was experimentally investigated. CDDSTs at various pitches i.e (p=0,p=15,p=16)

were used for the different ranges of Reynolds number. For cold water its ranges between 9000to17000 and for hot water 18000

to 24000. Results from CDDSTs were compared with plane tube and results showed that Nusslet number increased while friction

factor decreased with increased in Reynolds number. Friction factor was increased by 287% while Nusslet no increased by 28%.

. However thermal performance factor was maximum for CDDSTs (p = 15) with value 0.319.Keywords: Effectiveness, Friction Factor, Nusselt Number, Spring Turbulator,CDD(convergent- divergent- divergent)

_______________________________________________________________________________________________________

I. INTRODUCTION

Here experimental studies on heat transfer have been discussed with different configurations. Mainly heat transfer (Nusselt no)

and friction factor (FF) have been studied in detail with respect to different geometrical parameters in various ranges of Reynolds

number (Re).M.KANNAN et al [13] 2012 performed experiment to compare different types heat enhancement techniques with

the help of simulation. From the data, it was found that annular method gave better heat transfer than other methods V.

Kongkaitpaiboon et al [14] 2010 performed experiment to determine the effect of perforated conical rings (PCR) on Nusselt

number, friction factor and thermal performance factor characteristic. The experiment was carried in the range of Re from 4000

to 20000. It was found that the PCR enhanced the heat transfer more efficiently than the typical CR on the basis of thermal

performance factor of around 0.92 at the same pumping power. A. Mehta Kushal K et al [15] performed experiment to determinethe effect of insert and delta winglet in tube in tube heat exchanger The range of Re number was from 1400 to 6500. From the

results, it was found that Nusselt number for 50 cm insert was enhanced by 53%, for 20 cm was enhanced by 27% in comparison

to plain tube. Pongjet Promvonge et al [16] 2010 performed experiment to predict the combined effect of ribs and winglet type

vortex generators (WVGs) on heat enhancement and friction factor for turbulent air flow. Nusselt number and friction factor was

more in case of both ribs and WVGs as compared to alone ribs and winglet. Oliver and Shoji [17] performed experiments by

inserting wire coils in a tube using a non- Newtonian fluid and found that heat transfer is enhanced by a factor of 4 and the

relative pressure drop caused by the wire coil is by factor of 5. ] M. N. Lokhande, Dr. V. M. Kruplani, experimentally studied the

results of Heat Transfer Enhancement Techniques Using Ribs and Baffles. Experimental investigations have been carried out in

the rectangular duct to study the effect of Diamond shaped baffle of tip angle 100 on heat transfer enhancement, friction factor.

The heat transfer in rectangular duct with Diamond shaped baffle of tip angle 100 is to be more as compared to without baffle.

The increase in heat transfer coefficient of air higher for flat baffle and for Diamond shaped baffle over when no baffles in duck

drop.

II.

EXPERIMENTAL SET UP

The block diagram of heat exchanger is shown in Figure 4.1 and 4.2 shows the photograph of double pipe heat exchanger setup

used for the work. DBHE used for this setup consisted of following:

-

Pressure drop and heat transfer test sections,

-

One mild steel hot water tank having capacity of 260 liters, and

-

Two cold water tanks one for collecting outlet warm water of 600 liters capacity and one for supplying in the cold water

of 300 liters capacity.



Performance Analysis of Double Pipe Heat Exchanger using Convergent – Divergent-Divergent Spring Turbulators (IJIRST/ Volume 2 / Issue 02/ 037)


Hot water tank has four 2KW heaters installed in it that can maintain a maximum constant temperature of 75 degree Celsius.

Hot water motor has power of 1HP and that for cold water has power of 0.5HP. This is so because hot water was required to

supply higher LPMs then cold water at certain specific stages.

Test section includes two pipes,

Fig. 1:

- Inner pipe (smooth) of copper, of 4m length and its U-bend is of 0.232m length,

-

Outer pipe is made up of G.I. and is approximately equal in length to that of inner pipe.

Two well calibrated rotameters were used where hot water rotameters is 0-2000LPH and cold water rotameter has a range of

0-500LPH. Two pressure gauges are used both of 0-5kg/cm range and have a ± 0.01 error.

To measure inlet and outlet temperatures of Hot water and cold water four pt-100 RTDs are used and to measure outside wall

temperature of inner copper tube four chip sensors were used. Readings of temperature were noted down from multi-point digital

temperature indicators.

Data ReductionA.

Equations which form the basis of such experimental investigation can be summed as follows:The steady state of the heat transfer rate is assumed to be equal to the heat loss from the test section which can be expressed as:

Qair = Qconv (3.1)

Where,

Qair = ṁC p,a (To-Ti) (3.2)

Qconv = hA(Tw – T b) (3.3)

Where,

T b = (To + Ti)/2 (3.4)

Tw = Σ Tw/N (3.5)

Where, N – Total number of chip sensors or resistance temperature detectors between inlet and exit of the test section and

evaluation is done at the outer wall surface of the inner tube.

The convective heat flux is assumed to be uniformly distributed over the heated wall tube and can be evaluated as:

Qconv = h*A (Tw-T b) (3.7)

Where, T b = (To + Ti)/2 (3.8)The averaged heat transfer coefficient, h and the mean Nusselts number, Nu are estimated as follows:

h = m * C p,a (To-Ti) / A( T w-T b) (3.9)

Where, V is mean velocity in the tube:

V = ṁ/ρ A

Results and discussion

Verification of plain tube

For verification of plain tube following equation were used

Gnielinski correlation: NuPT,Theo=()

() for Re > 104 (3.10)

Dittus-Boelter Equation: NuPT,Theo = 0.023 * Re0.8 * Pr 0.4; for Re > 104 (3.11)





Sieder-Tate Equation: NuPT,Theo = 0.023*Re0.8*Pr 0.4*(μ/μs)0.14; for Re > 104 (3.12)

(μ/μs)0.14 is known as viscosity correction factor and falls very close to 1.45, hence is taken 1.45 for all calculations.

Petukhov Equation: NuPT,Theo = ⁄

()() for Re > 104 (4.10)

Using these eq.s it was found that nusselt no

Fig. 2: Nusselt No. Verification for Plain Tube

Fig. 3: Fiction for plain tube

Heat Tr ansfer Characteri sticsB.

Fig. 4: Relationship between Nusselt number and Reynolds number





While contemplating the quantitative analysis the results concluded are that the percentage heat transfer rate of the tube with

CDDST of WoutP, P=15cm and P=16cm inserted, when compared vis-à-vis plain tube at Re = 26970 is 28%, 19%, 16%

respectively. The heat transfer rate of the tube inserted with CDDST of different pitches is found within the range of 1.2 – 1.3

times higher vis-à-vis the heat transfer rate of plain tube.

F luid F low Characteri sticsC.

Fig. 5: Fluid flow characteristics

From Fig. it was observed that friction factor decreased with increased in Reynolds number for plain tube and tube with

CDDSTs. For a particular Reynolds number, tube equipped with CDDSTs having less pitch led to a high friction factor than

plain tube and other CDDSTs. The reason behind this was that due to less pitch, more CDDSTs were used. Due to more

CDDSTs, there was more obstruction to the hot water stream and hence more turbulence induced resulting in larger pressure

drop and hence friction factor.

Thermal Performance FactorD.

At constant pumping power

(λ *Re3)PT= (λ *Re3) T

Fig. 6: Thermal performance factor

When graph is plotted between thermal performance factorand Reynolds number that at constant pumping power, with an

increase in Reynolds number there is a decrease in thermal performance factor. Also, it can be seen that, for same pumping

power CDDST with P=15cm proved to be most efficient, the reason being the least friction offered by this pitch of CDDST

springs.





III. CONCLUSION

1) The Nusselts number was found to be enhanced by 28%, 19% and 16% when CDDST without pitch, with pitch 15cm

and 16cm were inserted in inner plain tube of double pipe heat exchanger, vis-à-vis plain tube.

2)

Friction factor and pressure drop characteristics were also studied and evaluated. It revealed that with an increment in

CDDST pitch friction factor and pressure drop increases. CDDST offers a maximum of 20%, 54% and 207% friction

factor with 16cm, 15cm and without pitch respectively, vis-à-vis friction factor generated by plain tube. However,

friction factor and pressure drop followed expected plot trends.

3)

Effectiveness offered by CDDST with varying pitches was also studied and it was found that the CDDST with least pitch generates least amount of friction factor , which leads to a maximum effectiveness of 0.319. Effectiveness by

P=15cm was 3.24% more than CDDST with P=16cm and 7.5% more than CDDST without pitch and that too at same

pumping power.

4) At lowest mass flow rates of Hot water and cold water, the amount of time contact between hot water and cold water on

either side of pipe wall, at a certain cross-section, was more, leading to large heat drops from hot water side and large

heat gain at cold water side. This trend decreases with increase in mass flow rates and hence Reynolds number. The

maximum heat gain & heat drop in plan tube is 11.517 &10.289 . On 15 cm pitch maximum heat drop and heat gain is

15.287 and 18.3. On 16 cm pitch maximum heat drop and heat gain is 12.6 and 17.14. On without pitch, Maximum heat

drop & heat gain is 12.6 & 14.9.

5)

Overall heat transfer coefficient was analyzed against mass flow rates of hot water and cold water. It can be concluded

that with an increase in mass flow rate heat transfer coefficient increases. As was expected, as it has same units as that

for convective heat transfer coefficient which followed exactly same trend on graphs.

NomenclatureA.

-

f – Friction factor

- y – Twist ratio

-

k – Thermal conductivity KW/ m c

- Pr – Prandtl number

-

Re – Reynolds number

- C p – Specific heat KJ/ kg c

-

Nu – Nusselt number

-

ρ – Density (kg/m3)

- U – Overall heat transfer coefficient KW/ m2 c

SubscriptsB.

-

avg Average

- h Hot

-

in Inlet

-

out Outlet

- w Water

-

c Cold

R EFERENCE

[1] M.Kannan, S.Ramu, S.Santhanakrishgan, G.Arunkumar, Vivek.M “experimental and analytical comparison of heat transfer in double pipe heat exchanger,”

International Journal of Mechanical Engineering applications Research 2012.

[2]

V. Kongkaitpaiboon, K. Nanan, S. Eiamsa-ard “Experimental investigation of heat transfer and turbulent flow friction in a tube fitted with perforatedconical-rings,” International Communications in Heat and Mass Transfer 37 (2010) 560– 567.

[3]

A. Mehta Kushal K, B. Lakdawala A M “Comparison of Heat transfer Enhancement in Tube in Tube heat exchanger using Different Turbulent Generator,”Institute of Technology, Nirma University, Ahmedabad – 382 481, 8-10 DECEMBER, 2011.

[4]

Pongjet Promvonge , Teerapat Chompookham, Sutapat Kwankaomeng, Chinaruk Thianpong “enhanced heat transfer in a triangular ribbed channel with

longitudinal vortex generators,” Energy Conversion and Management 51 (2010) 1242– 1249.

[5] Oliver, D. R. and Shoji (1992). Y. Heat transfer enhancement in a round tubes using different tube inserts: non-newtonian fluids. J. Chern. Engng. Res. and

Des.Vol.70, PP.558-564.

[6]

B.Adrian and K. Allan D, Wiley – interscience (2003). Heat transfer enhancement. In Heat Transfer Handbook, Chapter 14, Page 1033 - 1101.[7]

Bergles, A.E. (1985). Techniques to augment heat transfer. In Handbook of Heat Transfer Applications (Ed.W.M. Rosenhow), Ch.3 (McGraw- Hill, New

York)

[8]

Champagne, P.R. and Bergles, (2001). A.E. "Development and testing of a novel, variable Roughness technique to enhance, on demand, heat transfer in asingle - phase heat exchanger”. Journal of Enhanced Heat Transfer 8, Vol 5 (2001) Page 341- 352, http://www.bestinnovativesource.com/wp-

content/uploads/2012/10.

Documents

Performance Analysis of Double Pipe Heat Exchanger using Convergent - Divergent-Divergent Spring Turbulators