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