heat transfer augmentation in double pipe heat exchanger …€¦ · · 2016-07-22heat transfer...

INTERNATIONAL JOURNAL OF R&D IN ENGINEERING, SCIENCE AND MANAGEMENT

Vol.4, Issue 3, July 2016, p.p.37-55, ISSN 2393-865X

Available at :www.rndpublications.com/journal Page 37 © R&D Publications

Heat Transfer Augmentation in Double Pipe Heat Exchanger using

Divergent-Plain Spring Turbulators Karan Gopal

1, Sunil Dhingra

2,Gurjeet Singh

3

1M.Tech Scholar, Mechanical Engineering Department, UIET, Kurukshetra University, Kurukshetra 136119

2Assistant professor, Mechanical Engineering Department, UIET, Kurukshetra University, Kurukshetra, 136119

3 Assistant Professor, Mechanical Engineering Department, PEC University of Technology, Chandigarh 160012

ABSTRACT

A Heat enhancement technique has been utilized for many decades by evaluating different parameters of heat exchangers. The

work presented here focuses on heat transfer augmentation by hybrid combination of (divergent-plain) spring turbulators the

enhancement device at ratio of (3:2). Aim of the present work is to find such an optimum pitch in different ratios at which the

augmentation in heat transfer is maximum and the amount of power consumption is minimum, so that the appropriate and an

economic design can be created with maximum thermal efficiency. So, we are introducing the concept of pitch variation in ratio

3:2. This is defined as the horizontal distance between two consecutive turbulators. It describes that, the lesser is the pitch the

more numbers of turbulators that can be inserted in inner pipe of double pipe heat exchanger, hence more will be the friction

factor. This technique increases convective ability of the heat transfer process from the surface of inner pipe. There is a certain

limit to which a pitch can be decreased, lesser the pitch more the pressure drop and friction factor and hence the more will be the

pumping power requirement to maintain a desired mass flow rate of hot water. A analysis of thermal factors such as Nusselts

number friction factors, with different pitches of divergent plain spring turbulators of circular cross-section 6, 3 and 0 cm in ratio

of 3:2 at Reynolds’s number ranging between 40000 < Re < 65,000 is done graphically.

Keywords: Spring turbulators , Reynolds’s number etc..

______________________________________________________________________________________

NOMENCLATURE D Pipe diameter

g Acceleration due to gravity

h Heat transfer coefficient

hf Head loss

K Thermal conductivity of fluid

L Characteristics length of pipe

Ln Natural logarithm

m˙ Mass flow rate of fluid

Nu Local Nusselt number based on bulk temperature of the fluid

Nut Convective heat transfer coefficient of tube

Nup Convective heat transfer coefficient of plain tube

p Pitch length/spacing

P Static pressure

Heat Transfer Augmentation in Double Pipe Heat Exchanger using Divergent-Plain Spring Turbulators

Page 38

Pr Prandtl number

ΔP Pressure drop over length L of pipe

r Pipe radius

Re Reynolds number

Ts Surface temperature

T∞ Bulk temperature

T Temperature Cp Specific heat capacity

Subscripts

L Characteristic length of pipe

M Mean

Max Maximum

min Minimum

PT Plain tube

t Turbulator

Theo Theoretical

Exp. Experimental

O Outlet, outer

i Inlet, inner

a Air

w Pipe wall

Greek symbols

λ Darcy friction factor

η Thermal performance

μ Fluid viscosity

ν Kinematic viscosity

ρ Fluid density

Δ Net change in quantity

τ Shear stress

β Coefficient of thermal expansion

Abbreviations

DPST-C Divergent Plain spring turbulator-circular

DPHE Double pipe heat exchanger

LPH Liter per hour

LPM Liter per minute

RTD Resistance temerture detector

1. INTRODUCTION

A heat exchanger is a device that is used to transfer thermal energy (enthalpy) between two or more fluids, between a

solid surface and a fluid, or between solid particulates and a fluid, at different temperatures and in thermal contact. In

heat exchangers, there are usually no external heat and work interactions. Typical applications involve heating or

cooling of a fluid stream of concern and evaporation or condensation of single or multi-component fluid streams.


9

Temperature gradient is the factor facilitating heat transfer including a certainty that heat exchange occors in the

direction of decreasing temperature. He-at transfer theory explains itself by three peculiar modes of heat

transmission, radiation, conduction and convection. In Heat exchangers radiation phenomenon does take place, but

its role is rather insignificant vis-à-vis conduction and convection.

We know that Tubes with rough surfaces have much higher heat transfer coefficients than tubes with smooth

surfaces. Therefore, tube surfaces are often willfully roughened, corrugated, or finned in order to enhance the

convection heat transfer coefficient and thus the convection heat transfer rate. Heat transfer in turbulent flow in a

tube has been increased by as much as 400 % by roughening the surface [15]. Roughening the surface, of course,

also increases the friction factor and thus the power requirement for the pump or the fan. Designing a heat exchanger,

which suits majority of applications, is very difficult as w-ell important, as it always has the limitation of size and

fluid flow rate, resulting in a low heat transfer rate.

The augmentation of heat transfer is the ability to achieve high performance heat exchangers, leads to its size

reduction and high initial investment. So, the passive heat enhancement techniques can be applied by installing the

turbulence generators or turbulators, e.g. the insertion of twisted stripes and tapes [2–13], the insertion of coil wire

[10,17,21] and helical wire coil in the heat exchangers. The results of those studies show that although heat transfer

efficiencies were improved, the friction factor of pipes was considerably increased.

In recent years, thousands of numerical and experimental studies have been performed on heat transfer enhancement

techniques of different configuration. Mainly heat transfer and frictional characteristics have been studied in detail

with respect to different geometrical parameters in various ranges of Reynolds number. Further these studies have

been cross-verified with researches already performed in this field. It is investigated by studying various research

papers that there is a scope of design modification in heat transfer enhancement device which can significantly affect

the rate of heat transfer and friction factor. As it is clear, twisted tape gives better results which encourage new

researchers to further improve the design. So, in this research a new concept inspired by the literature which is

divergent plain spring turbulators of circular cross section (DPSTC).

The spring type design of DPST is influenced from experiments performed by Kumbhar et al. [11] as heat transfer

behavior in a tube with conical wire coil inserts. The conical shape and divergent-plain section is influenced from

experiments performed by Eiamsaard et al. [19] as enhancement of turbulent flow heat transfer in a tube by using

nozzle turbulators. Sufficient pitch (as zero pitch increased the friction factor and hence the pumping power re-

quirements) is provided in DPST which is analogous to experimental investigation of heat transfer and turbulent flow

friction in a tube fitted with perforated conical-rings by Kongkaitpaiboo et al. [22].

So far no work is reported on the study of spring turbulators of varying cross section, so it is a new design which is

influenced by twisted tapes and wire coil inserts. This research work is focused upon overcoming the limitations im-

posed previously So, the DPST-C design advantage is that it uses less material and also expected to have higher heat

transfer rate at the cost of lesser pressure drop.


0

Fig.1 Block Diagram of Double Pipe Heat Exchanger

2 METHODOLOGY AND EXPERIMENTAL SET UP

2.1 Objectives

The focus here is on the experimental, graphical analysis & study of the effect of divergent-convergent springs

having circular cross-sections, on the following factors & parameters:

1. Heat gain and heat drop

2. Friction factor

3. Nusselt number

4. Convective heat transfer coefficient

5. Overall heat transfer coefficient

6. Nusselt number versus Reynolds number for verification of Nusselt number of plain tube

7. Ratio of friction factors of DPST and that of plain tube versus Reynolds number.

8. Friction factor verification of plain tube.

9. Ratio of Nusselt number of DPST and that of plain tube versus Reynolds No.

10. Mass flow rate of hot water versus heat gain and and heat drop.

11. Thermal performance factor versus Reynolds number


Page 41

2.2 Assumptions

1. Flow is assumed to be steady.

2. Flow is non-uniform.

3. Flow is incompressible.

4. Isothermal conditions are maintained, though minor heat losses are neglected.

5. Coefficient of thermal expansion on inner side and coefficient of thermal contraction on outer side of

inner pipe each other, of concentric tube heat exchanger.

6. Inner pipe’s inner surface is assumed to be smooth.

7. Sieder-tate equation takes into account the change in viscosity (μ and μs) due to temperature change

between the bulk fluid average temperature and the heat transfer surface temperature, respectively. The

viscosity factor will change as the Nusselt number changes.

Fig. 2 Photograph of double pipe heat exchanger

2.3 Experimental setup

The experimental set up used in the present work is shown in figs. 1 and 2 is discussed as follows:


Page 42

The experimental setup include a hot water tank consisting of four heaters 2KW each that can maintain a maximum

constatnt temperature of 75 °C. The setup includes two motors of capacity 1 HP for hot water and 0.5 HP for cold

water. This is so because hot water is required to supply higher L-PHs than cold water at certain specific stages. Test

section includes two pipes, Inner pipe (smooth) of copper, 4m length and its U-bend is of 0.232m length, and outer

pipe is made up of G.I. which is insulated and is approximately equal in length to that of inner pipe. Two well

calibrated rotameters of range 0–2000 and 0–1500 LPH are used for hot water and cold water respectively. Two

pressure gauges are used of range 0–5 kg/cm2 with ±0.01 error. To measure inlet and outlet temperatures of hot

water and cold water four pt-100 RTDs are used and to measure out-side wall temperature of inner copper tube four

chip sensors are used and all the experimental work taken under insulated environment. Readings of temperature are

noted down from multi-point digital temperature sensor indicators.

2.4 Augmetation technique used in current work

Aim of the work was to employ divergent-convergent turbulators (DPST) of circular (DPST-C) cross-section as

shown in Fig. 3, inside of copper tube of double pipe heat Exchanger. DPST-C s made of high carbon spring steel

with 10 cm of free length and an external diameter of 2.1 cm. It was only sufficiently large so that it could make an

interference fit with inner tube of DPHE. These were mounted at regular intervals on two thin cylindrical rods with

rod diameter equals to the minimum diameter or midsection of DPST-C. Three rods mounted with DPST-C are

inserted in the inner tube prior to attaching U-bend section to outer tube as shown in Fig. 3a–c. Inner tube contained

hot water flow with flow rate ranging from 700 to 2000 LPH and outer tube contained cold water with flow rate

varying between 500 and 1500 LPH. Rotameters were installed to measure the flow rate, and was controlled by flow

regulating valves. As described earlier this work utilized passive technique for heat transfer enhancement, which is

primarily aimed to generate a swirl/turbulent flow.

Fig.3 (a) DPST mounted on rod in ratio 3:2 without pitch. (b) DPST mounted on rod in ratio 3:2 p =

3cm. (c_ DPST mounted on rod in ratio 3:2 p = 6cm

A number of DCST-C were mounted on a brass rod of 8 mm diameter and 200 cm of length, by brazing to eliminate

any undue movement of DCST-C, inserted in inner pipe of copper and flow was initiated. Three variations in pitch

ratios were provided viz. 0, 3 and 6 cm as shown in Fig. 3a–c. The spring pitch was kept constant in all of the DCST-

C. As it can be observed from Fig. 3a–c that lower the pitch results in more number of springs that can be mounted

on a single rod. The term “pitch” used throughout the paper refers to the distance between two consecutive DCST-C

when mounted on brass rod. The concept behind varying the pitch was that lower is the pitch the more are the


Page 43

DCSTC that can be mounted on a single rod and more will be the obstruction they will cause to hot water when

inserted in inner copper pipe and more will be the friction factor produced leading to more fluid mixing, breaking of

boundary layer, swirl production and consequently higher will be the heat transfer.

Fig. 4 Validation of Nusselt number for plain tube with literature

Fig. 5 a Effect of pitch on HTE for DPST with plain tube, Pitch P = 6, 3, 0 cm in ratio of (3:2).

100

150

200

250

300

350

400

40000 45000 50000 55000 60000 65000 70000

Nu

Re

Nu (TH)

Nu DB

Nu ST

Nu exp

Nu PT

0

50

100

150

200

250

300

350

400

41,279.67 49,355.64 56,957.50 64,192.12 64,746.90

Nu

Re

Nu, Plain Tube

Nu, p 0cm (3:2)

Nu, p 3cm (3:2)

Nu, p 6cm (3:2)


Page 44

Fig. 5 b Effect of pitch on HTE for DPST with P = 0, 3 and 6 cm in ratio (3:2) NuT/NuPT versus Re

2.5 Experimental procedure

2.5.1 Step 1: Rotameter and RTDs calibration

Two buckets of 25 L were used to collect water for 3 min which flows through cold water and hot water side

rotameter (shown in Tables 1, 2). Further, four RTDs were calibrated by dipping in water troughs one after another

and observed readings were compared with the referenced set by premeasured RTD value in Table 3.

2.5.2 Step 2: Standardization and verification

Plain tube readings were obtained for pressure drop and hence friction factor characteristics at normal temperature of

water. The main motive behind this study was to insert DCST of varying pitches in inner plain tube which could

make its surface rough, and hence could alter its readings. After obtaining readings, the collected data was compared

with calculated theoretical value.

2.5.3 Step 3: Initiating and executing plain tube experimentation

After obtaining the readings as mentioned in step-II, 260 L of water was heated to 75 °C which took approximately

1.5 h to reach this stage. Then the hot and cold water motors were started at same LPH and steady fl flow was

obtained in 15 min. LPH of hot water was maintained constant for six variations in LPH of cold water and this

procedure was obtained for a range of 500–1500 LPH of cold water and 700–2000 LPH range of hot water. The cold

water and hot water followed counter flow directions at their respective flow rates. When constant and stable values

at temperature display panel were obtained, then the inlet–outlet and inner wall temperatures were measured. After

obtaining the results for plain tube the Nusselt number verification for plain tube is done by comparing with results

obtained from Eq. (14).

2.5.4 Step 4: Preparing DCST for inner tube insertion

0

0.2

0.4

0.6

0.8

1

1.2

1.4

41,279.67 49,355.64 56,957.50 63,657.78 64,746.90

Nu

T/N

uP

T

Re

NuT/NuPT for p 0cm (3:2) NuT/NuPT for p 3cm (3:2) NuT/NuPT for p 6cm (3:2)


Page 45

DCST of length 18 cm were mounted and brazed on 2 m brass rods at a pitch of 15, 10 and 5 cm after wards. A pair

for each pitch set was made as shown in Fig. 4a–c. The U-bend of pipes were given a detachable flange coupling

joint due to which it became fairly convenient to insert DCSTs in and out. The same procedure to initiate the

experiment was followed as mentioned for plain tube. The steady state was achieved within 20 min after insertion of

DCSTs. Friction factor and pressure drop results were collected and compared with plain tube results. Same

procedure was followed for all DCSTs with varying pitches.

2.5.5 Step 5: Thermal performance results and repeatability

The procedure was re-initiated for thermal performance and repeatability check (shown in Tables 4, 5). For a

constant pumping power equivalent Reynolds number was calculated and water was kept at 75 °C.

2.5.6 Step 6: For heat transfer coefficient calculation

Hot water at 75 °C is allowed to pass through the inner pipe of heat exchanger at 1000 LPH (mh = 0.2715 kg/s).

Cold water is now allowed to pass through the outer pipe of heat exchanger in counter current direction at 1000 LPH

(mc = 0.2715 kg/s). The inlet and outlet temperatures for both hot water and cold water (T1–T4) are recorded only

after temperature of both the fluids attains a constant value. The procedure was repeated for different cold water flow

rates.

2.6 Data reduction

Heat exchangers with working fluid water was taken in all of the experiments mentioned in previous section, with

parametric study of effects of Reynolds number varying from 40000 to 65,000. Aspiration behind the variation of

pitch ratios was basically to enhance the friction factor which varies in an inverse proportion with all these

parameters. But, increasing friction factor i.e. decrease in pitch ratio or an increase in the number of turbulators

employed, puts a direct impact on pressure drop and hence on the required pumping power. But, to maintain higher

levels of Reynolds number or high LPMs, a high and constant pumping power is desired. Hence, it is required to

maintain optimum conditions such that a balance can be created between pitch ratio, friction factor, pressure loss and

hence pumping power. Equations which form the basis of such experimental investigation can be summed as

follows. Since Reynolds number represents the ratio of momentum to viscous forces the relative magnitudes of Gr

and Re are an indication of the relative importance of natural and forced convection in determining heat transfer.

Forced convection effects are usually insignificant when Gr/Re2 >> 1 and conversely natural convection effects may

be neglected when Gr/Re2 << 1. When the ratio is of the order of one, combined effects of natural and forced

convection have to be taken into account. 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 (1)

Where,

Qair = m.Cp,a(T0 − Ti) (2)

Qconv = hA(Tw − Tb) (3)

Where

Tb = (To + Ti)/2 (4)


Page 46

Tw = Σ Tw/N (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 averaged heat transfer coefficient, h and the mean Nusselts number, Nu are estimated as follows:

h = m * Cp,a (To-Ti) / A( Tw-Tb) (6)

Nu = hd / k (7)

2.7 Standard equations

2.7.1 Friction factor calculations

Darcy-Weisbach Friction Factor: λ = (Δ P/L)∗D (8)

(ρ V2/2)

Blasius Equation: λ = 0.316 (9)

Re0.25

Colburn’s Equation: λ = 0.046 (10)

Re0.2

2.7.2 Heat transfer calculations

Dittus–Boelter Equation [11,20]:

NuPT,Theo = 0.023 * Re0.8

* Pr0.3

; for Re > 104

Sieder-Tate Equation [12, 21]:

NuPT,Theo = 0.023 * Re0.8

* Pr0.4

* ( μ/μs)0.14

; for Re > 104

(μ/μs)0.14

is known as viscosity correction factor and falls very close to 1, hence is taken unity for all calculations.

When the ratio was measured for wide range of Reynolds number (especially at lower values), the experimental

readings obtained for plain tube tends to bend greater towards deviations of more than 90 % for which the

simultaneous action of natural and forced convection are held accountable. It is worthy to note that natural

convection phenomena is more pronounced and dominating at lower values of Reynolds number whereas forced

convection ruled the upper limits of Reynolds number.

2.7.3 Thermal performance calculations

At constant pumping power

(λ*Re3)PT= (λ*Re

3) T (13)

(14)


Page 47

3.RESULTS AND DISCUSSION

3.1 Heat transfer analysis

3.1.1 Plain tube

The first step before starting the experiments that had any inclusions of varying pitch turbulators was to measure

plain tube readings keeping several parameters in mind to be calculated. Nusselt number is one of such parameters

which, sequentially is thought to be measured under an unvarying or constant condition of heat flux. Then compared

the obtained results for convective heat transfer coefficient and Nusselt number vis-à-vis the results that were

obtained from the fundamental equations given by Dittus–Boelter, as mentioned in Eq. (11). The main motive behind

conducting the plain tube experiments was the experimental validation of plain tube. From Fig. 4, it can be

concluded that the results obtained from plain tube experiments, for heat transfer i.e., the trend followed by the graph

representing the variation of Nusselt number with Reynolds number lies well within the agreement depicted by the

graphical trends formed by Eq. (11).

3.1.2 DPST of different pitches

When DPST with different pitches (P = 0, 3 and 6) were inserted in internal copper tube of double pipe heat

exchanger in ratio (3:2), it exhibited different trends when the graph between the obtained experimental values of

Nusselt number and Reynolds number was plotted, which is depicted in Fig. 5a, b. While analyzing the graph, it is

clear that the rate of heat transfer is significantly enhanced when the plain tube and tube with DPST-C with various

pitches and different ratio are compared for a given fixed value of Reynolds number. Hence, obstruction, or

resistance to flow, caused by DPST-C is the phenomenon which is accountable for such enhancement, which signifi-

cantly increases with decreasing pitch. This obstruction intrigues the thermal boundary layer destruction adjacent to

the inner tube wall leading to swirl flow and local turbulent zones thereby augmenting the heat transfer and heat

transfer ratio with increase in Reynolds number. This, on the other side, agitates the entirety of thermal boundary

layer thereby increasing the value of heat transfer coefficient. While contemplating the quantitative analysis the

results concluded that the heat transfer rate of the tube having DPST-C at different pitches and different ratio is

found within the range of 1.10–1.25 times higher vis-à-vis the heat transfer rate for plain tube.

3.2 Friction factor analysis

3.2.1 Plain tube

Friction Factor is the first parameter that is measured before commencing any experiments related to DPST-C, i.e.,

when the apparatus is freshly fabricated, so as to validate and verify the plain inner copper tube for pressure drop and

friction factor by comparing to the standard data under optimal experimenting conditions. The obtained results of

friction factor were then compared vis-à-vis the results obtained from the fundamental equations given by Blasius

and Colburn as mentioned in Eqs. (12) and (13). From Fig. 6, it can be concluded that the results obtained from plain

tube experiments, for friction factor i.e., the trend followed by the graph representing the variation of friction factor

with Reynolds number lies well within i.e., ±10 to ±15 % agreement depicted by the graphical trends formed by Eqs.

(11) and (12). Darcy–Weisbach equation was used for calculating experimental values of friction factor.


Page 48

Fig. 6 Verification of friction factor for plain tube

3.2.2 DPST of different pitches

Experiments were performed under unvarying or constant conditions of heat flux for measuring the effect of

inserting DPST-C of different pitches viz. 0, 3 and 6 cm in ratio of (3:2). To obtain the friction factor and to plot the

trend followed, calculations were made when friction factor varied with the varying Reynolds number. The results

hence obtained were compared with that obtained while using plain inner copper tube. In Fig. 7 a, the trend indicates

an increment in friction factor with a decrease in DPST-C pitches. The reason behind this phenomenon was that the

less is the distance between two consecutive DPST-C the more are the DPSTC that can be mounted on the

cylindrical rod and hence inserted in inner copper tube, consequently causing more obstruction to the hot water

stream, and hence more is the turbulence induced resulting in larger pressure drop and hence increase in friction

factor.

Fig. 7 Friction factor comparison of DPST and plain tube; (fT/fPT vs. Re)

0

0.005

0.01

0.015

0.02

0.025

0.03

0.035

0.04

0.045

19000 24000 29000 34000

f

Re

f,DWEF

f,Petu

f,Blausius

f,colburn

0

0.5

1

1.5

2

18000 23000 28000 33000

fT/f

PT

Re

R1 = fp/fpt (3:2)

R2 = fp/fpt (3:2)

R3 = fp/fpt (3:2)


Page 49

3.3 Thermal performance factor

At this stage, experimentation involved DPST with varying pitches in ratio of (3:2), which includes an augmentation

of heat transfer rate, though with a simultaneous increment in friction factor with decrease in pitch. The criteria of

decreasing pitches comes at a price, that is, every extra turbulators inserted cause an extra resistance to flowing water

and hence leading to an increment in friction factor. The more is the obstruction caused the more will be the effort

required by pump to sustain a constant mass flow rate and hence more will be the pumping power required which at

a certain level proves to be uneconomical. Consequently, a mutual agreement or an optimum state has to be

concluded between the effectiveness of DPST in heat transfer augmentation and the increase in friction factor it

causes. This problem is to be judged from performance evaluation criteria. As depicted in Fig. 8, when graph is

plotted between thermal performance factor and 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 DPST with P = 0 cm in ratio of (3:2) proved to be most efficient, the reason being the least friction offered by

this pitch of DPST springs. In present work, Nusselt number (Nu) is increased by only 11.78-28. 58.1% in ratio of

(3:2) increase in friction factor for Reynolds no. range (40000–65,000), whereas maximum increase in Nusselt

number is claimed by Kongkaitpaiboon et al. [22]. So, in present work a wide range of Reynolds number is used and

at higher Reynolds number increase in friction factor is in well acceptable range which is much lower as compared

with the previous studies.

Fig. 8 Thermal performance factor versus Reynolds number

0

0.4

0.8

1.2

1.6

41,279.67 49,355.64 56,957.50 64,192.12 64,746.90

η

Re

η for p 0cm (3:2)

η for p 3cm (3:2)

η for p 6cm (3:2)


Page 50

3.4 Comparision of performance parameters of different mechanical turbulators

R.NO MAJOR CONCLUSION NAME & IMAGE OF INSERT

01

HT (TTI)>HT(WTTI)

FF (TTI)>FF(WTTI)

FLUID: WATER

Re: 7000-25000

TWISTED TAPE INSERT.

02

HT (MTC)>HT(STC)

BY 30%

FF (MTC)>FF(STC)

(2.4-17.9)

ηHT(MTC)> ηHT(STC)

(2.7-4.2)

FLUID: WATER

Re:3000-60000

THREE-START SPIRALLY CORRUGATED

TUBES COMBINED WITH FIVE TWISTED

TAPE INSERT

(IMAGE NOT AVAILABLE)

04

HT(RCTTI)>HT(PT)

2.3-2.9 Times

FF(RCTTI)>FF(PT)

39-80%

10000<Re<19000

FLUID: WATER

RECTANGULAR TWISTED TAPE INSERT.

05

HT(C-CCTTI)>HT(TT)

12.8-41.9%

FF(C-CCTTI)>FF(TT)

2.44-3.59 Times

HT(C-CCTTI)>HT(PT)

12.5-41.5%

ALTERNATE CLOCKWISE AND COUNTER


Page 51

FF(C-CCTTI)>FF(TT)

3.42-5.10 Times

HTEI (C-CCTTI)>HTEI(TT)

1.28-1.19 (PP=CONST)

3000<Re<27000

FLUID: WATER

CLOCKWISE TWISTED TAPE INSERT.

06

HT(CRT)>HT(PT)

By up to 17%

ηEE(CRT)> ηEE(PT)

0.86-1.16 Times

Re: 5000-25000(PP=C)

FLUID: WATER

CONICAL RING TURBULATOR.

07

HT(DCT)>HT(PT)

11.46 % < Nu < 26.76 %

FF(DCT)>FF(PT)

20.79–66.87 %

Re: 5000-40000

FLUID: WATER

DIVERGENT CONVERGENT SPRING TURBUL-

ATORS

08

Present work

HT(DPST) (3:2)>HT(PT)

11.78%<Nu<28.57%

FF(DPST)(3:2)>FF(PT)

48.71-58.1%

Re: 40000-65000

FLUID: WATER

DIVERGENT PLAIN SPRING TURBULATORS


Page 52

4. CONCLUSION

The objectives mentioned are successfully performed in the range of cold water varying from 500 to 1500 LPH and

hot water ranging from 500 to 2000 LPH, thereby obtaining a wide range of Reynolds number from as low as 40,000

to as high as 65,000. The effects of DPST-C with varying pitch in different ratios when inserted in inner plain tube of

double pipe heat exchanger are studied including their role in heat transfer augmentation. The most significant

conclusions that are drawn after performing this experiment are as:

(a) Keeping the experimentation conditions approximately identical, Heat transfer augmentation, is best provided by

using DPST-C with P = 0 cm in ratio (3:2). It depicts an inverse trend where Nusselt number ratio (thermal effi-

ciency) decreases with increase in Reynolds number. It can be comprehended as the relative effect of plain tube and

tubes with DPST-C installed, on heat transfer augmentation. Hence, it can be concluded that DPST-C (best results

given by DPST-C with P = 0 cm in ratio (3:2)) performs great at relatively lower values of Reynolds number and

their effect, not sharply but gradually decreases with increment in Reynolds number.

(b) The Nusselt number is found to be enhanced by 11, 14.88 and 28.76 % with DPST-C pitches P = 6, 3 and 0 cm in

ratio (3:2), vis-à-vis plain tube.

(c) Friction factor and pressure drop characteristics were also studied and evaluated. It reveals that, with an

increment in DPST-C pitch in different ratio, friction factor and pressure drop increases. DPST-C offers a maximum

of 48.71, 56.41 and 58.97% friction factor with 6, 3 and 0 cm in ratio (3:2) respectively, vis-à-vis friction factor

generated by plain tube.

(d) Thermal performance factor offered by DPST-C with varying pitches is also studied and it is found that the

DPST-C with high pitch generates least amount of friction factor vis-à-vis DPST-C with P = 0 cm in ratio (3:2) and

P = 0 cm in ratio (1:1), which leads to a maximum thermal performance factor of 1.346044 in ratio (1:1) and

1.300545 in ratio (3:2). Thermal performance factor for P = 0 cm was 12.16% more than DPST with P = 3 cm and

2.92 % more than DPST-C with P = 6 cm in ratio (3:2) and that too at same pumping power. The experiment for

augmentation of heat transfer is successfully performed with DPST-C arrangement in double pipe heat exchanger

and heat transfer (Nusselt no) is enhanced by 48.71 %, Friction factor is increased maximum of 58.97 % for DPST-C

pitch 0 cm in ratio (3:2) vis-à-vis plain tube.

Appendix See Tables 1, 2, 3, 4 and 5.

Table 1

For Large Rotameter

LPH m

(kg/s) T(s)

Observation

1 (kg)

m1

(ks/S)

Observation

2 (kg)

m2

(kg/s)

Observation

3 (kg)

m3

(kg/s) Mavg % Error

900 0.25 180 40.1 0.2388 41.2 0.2386 41.1 0.2386 0.2386 4.541481

1000 0.2778 180 43.7 0.2656 44 0.2656 44 0.2656 0.2656 4.394667

1100 0.3056 180 49 0.2919 49.5 0.2918 49.3 0.2919 0.2919 4.483636

1200 0.3333 180 54.2 0.3182 54.8 0.3181 54.8 0.3181 0.3181 4.557778

1300 0.3611 180 60.7 0.3442 61 0.3442 61 0.3442 0.3442 4.689231


Page 53

Table 2

For Small Rotameter

LPH m

(kg/s) T(s)

Observation

1 (kg)

m1

(ks/S)

Observation

2 (kg)

m2

(kg/s)

Observation

3 (kg)

m3

(kg/s) Mavg % Error

300 0.0833 180 15 0.07917 14.7 0.0793 14.9 0.0792 0.0792 4.955556

400 0.1111 180 20.3 0.10547 20 0.1056 20.1 0.1055 0.1055 5.033333

500 0.1389 180 25.3 0.13186 25.2 0.1319 25.2 0.1319 0.1319 5.046667

600 0.1667 180 30.4 0.15822 30.1 0.1583 30.1 0.1583 0.1583 5.033333

700 0.1944 180 35.3 0.18464 35 0.1847 35 0.1847 0.1847 5.014286

Table 3

RTD Calibration

T2 T6 T7 T8

Obs. 1 27.1 27.1 27.1 27.1

Obs. 2 27.1 26.8 27.1 26.8

Obs. 3 28.1 27.3 28.1 28.1

Obs. 4 27.2 27.2 27.2 27.2

Obs. 5 27.1 27.1 27.1 27.1

Obs. 6 26.9 27.1 26.9 26.9

Obs. 7 27.1 28.1 27.1 27.1

Obs. 8 27.1 27.1 27.1 27.3

Calibration ±1 ±1 ±1 ±1


Page 54

Table 4

Heat transfer versus Re for DCST having pitch 0 cm in ratio (3:2)

Mc

(kg/s)

T8 =

Tci

T6 =

Tco

Mh

(kg/S) Re HW

T2 =

Thi

T7 =

Tho

Trial 1

Nu exp

Trial 2

Nu exp % diff.

0.333 27.4 48 0.277 40550.44 79.3 56.6 373.9634 377.6404 -3.67699

0.333 27.4 48.1 0.333 48412.24 77 57.9 379.0766 375.0164 4.060209

0.333 27.5 48 0.388 55628.87 74.7 58.4 377.371 370.2797 7.091367

0.333 27.5 47.3 0.444 62324.6 71.7 58.3 355.7402 358.701 -2.9608

0.333 27.6 46.4 0.458 62833.94 69.4 57.4 328.0932 330.0805 -1.98733

Table 5

Heat transfer versus Re for DCST having pitch 3 cm in ratio (3:2)

Mc

(kg/s)

T8 =

Tci

T6 =

Tco

Mh

(kg/S) Re HW

T2 =

Thi

T7 =

Tho

Trial 1

Nu exp

Trial 2

Nu exp % diff.

0.333 29.1 48.5 0.277 40047.79 78.4 55.8 298.6421 301.2075 -2.5653

0.333 29.2 48.6 0.333 47976.68 76.1 57.6 307.8319 313.207 -5.3751

0.333 29.2 48.7 0.388 55395.63 74.2 58.2 308.8471 311.3151 -2.4680

0.333 29.3 48.2 0.444 62324.6 71.8 58.2 328.227 333.6538 -5.4267

0.333 29.3 47.6 0.458 64198.53 69.7 57.7 321.1593 324.9512 -3.7919

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