We also found several experimental data sets for forced convective heat transfer during gas–liquid two-phase flow in vertical pipes, very limited data for horizontal pipes, and no data f
Trang 2CopyrightC Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457630903500809
e d i t o r i a l
Heat Transfer in Industrial
Applications—PRES 2008
JI ˇ R´I KLEME ˇS1and PETR STEHL´IK2
1Centre for Process Integration and Intensification (CPI2), Faculty of IT, University of Pannonia, Veszpr´em, Hungary
2Institute of Process and Environmental Engineering, Brno University of Technology, Brno, Czech Republic
This editorial provides an overview of a special issue dedicated to the 11th Conference on Process Integration, Modeling, and
Optimization for Energy Saving and Pollution Reduction—PRES 2008 Nine papers have been selected and peer-reviewed
covering important subjects of heat transfer engineering They focus on recent development of various features of heat
transfer equipment design and optimization This issue of Heat Transfer Engineering is the sixth special journal issue
dedicated to selected papers from PRES conferences [1–5].
INTRODUCTION
Issues of global warming and greenhouse gas emissions,
to-gether with other pollution and effluents, are increasingly one
of the major technological and also important societal and
po-litical challenges Because of the increasing urgency, various
conferences are being held to encourage closer collaboration
among people of many nations about the problems, and progress
in meeting these challenges A very important contribution to
successfully deal with those problems can be offered by heat
transfer engineering
The series of conferences on Process Integration, Modeling,
and Optimization for Energy Saving and Pollution Reduction
(PRES) is one such opportunity for cross-fertilization, running
now into its second decade It was established originally to
ad-dress issues relevant to process energy integration in connection
with the efficient heat transfer issues The organisers of the
PRES conferences are proud to continuously attract delegates
from numerous countries worldwide, providing a friendly and
highly collaborative platform for fast and efficient spreading of
novel ideas, processes, procedures, and energy-saving policies
PRES conferences have a comprehensive publication strategy:
Address correspondence to Prof Jiˇr´ı Klemeˇs, Centre for Process Integration
and Intensification (CPI 2 ), Research Institute of Chemical Technology and
Pro-cess Engineering, FIT, University of Pannonia, Egyetem u 10, 8200 Veszpr´em,
Hungary E-mail: klemes@cpi.uni-pannon.hu
see refs [1] to [8] This special issue is already the sixth
spe-cial issue of Heat Transfer Engineering dedicated to selected
contributions from PRES conferences
PRES 2008 was held, as it has been traditionally every ond year, in collaboration with the 18th International CongressCHISA 2008 in the heart of Europe—in Prague, the capital ofthe Czech Republic, 24–28 August 2008 This Central Euro-pean capital, known as a city of a thousand spires, welcomeddelegates from more than 55 countries; 987 authors submitted
sec-345 contributions They represented, beside traditional pean countries, Asia, Africa, Australia, and North and SouthAmerica
Euro-SELECTED CONTRIBUTIONS
For this special issue of Heat Transfer Engineering, nine
papers dealing with various aspects of heat transfer engineeringand related inputs are included They tackle various aspectsand levels of industrial implementations from two-phase flow,through compact heat exchangers and microwaves to total sites.The first paper presents a keynote lecture, “Importance ofNon-Boiling Two-Phase Flow Heat Transfer in Pipes for Indus-trial Applications,” authored by Afshin J Ghajar and Clement
C Tang from School of Mechanical and Aerospace ing, Oklahoma State University, Stillwater, Oklahoma, USA
Engineer-707
Trang 3They present extensive results of the recent developments in the
non-boiling two-phase heat transfer in air–water flow in
hor-izontal and inclined pipes conducted at their two-phase flow
heat transfer laboratory The validity and limitations of the
nu-merous two-phase non-boiling heat transfer correlations that
have been published in the literature over the past 50 years
are discussed Practical heat transfer correlations for a
vari-ety of gas–liquid flow patterns and pipe inclination angles are
recommended The application of these correlations in
engi-neering practice, and how they can influence the equipment
design and consequently the process design are discussed in
detail
In their future plans they stated that the overall objective
of their research has been to develop a heat transfer
corre-lation that is robust enough to span all or most of the fluid
combinations, flow patterns, flow regimes, and pipe
orienta-tions (vertical, inclined, and horizontal) They made a lot of
progress toward this goal However, to fully accomplish their
research objectives, a much better understanding of the heat
transfer mechanism in each flow pattern is needed They plan to
perform systematic heat transfer measurements to capture the
effect of several parameters that influence the heat transfer
re-sults They also plan to complement these measurements with
extensive flow visualizations They claim that the systematic
measurements would allow them to develop a complete database
for the development of their “general” two-phase heat transfer
correlation
The second paper presents a novel extension of heat
inte-gration methodology stressing an enhanced heat transfer It is
titled “Total Sites Integrating Renewables With Extended Heat
Transfer and Recovery,” authored by Petar Varbanov and Jiˇr´ı
Klemeˇs from the Centre for Process Integration and
Intensifi-cation (CPI2), Research Institute of Chemical Technology and
Process Engineering, Faculty of Information Technology
Uni-versity of Pannonia, Veszpr´em, Hungary The challenge of
in-creasing the share of renewables in the primary energy mix
could be met by integrating solar, wind, and biomass as well
as some types of waste with the fossil fuels Their work
ana-lyzed some of the most common heat transfer application at total
sites The energy demands, the local generation capacities, and
the efficient integration of renewables into the corresponding
total sites CHP (combined heat and power generation) energy
systems, based on efficient heat transfer, are optimized
mini-mizing heat waste and carbon footprint, and maximini-mizing
eco-nomic viability The inclusion of renewables with their changing
availability requires extensions of the traditional heat
integra-tion approach The problem becomes more complicated and has
several more dimensions Revisiting some previously developed
process integration tools and their further development enables
solving this extended problem Their contribution has been a
step in this direction, summarizing the problem and suggesting
some options for its solution A demonstration case study
illus-trated the heat-saving potential of integrating various users and
using heat storage Their future work progresses to developing
advanced software tools based on the suggested methodology
“Alternative Design Approach for Plate and Frame Heat changers Using Parameter Plots” by Mart´ın Pic´on-N´u˜nez, Gra-ham Thomas Polley, and Dionicio Jantes-Jaramillo, from theDepartment of Chemical Engineering, University of Guanaju-ato, in Mexico, follows their previous paper published withinthe PRES Conference series [9] and analyzes the simultaneousdesign and specification of heat exchangers of the plate andframe type They used a pictorial representation of the designspace to guide the designer toward selection of the geometrythat best meets the heat duty within the limitations of pressuredrop The design space was represented by a bar plot wherethe number of thermal plates is plotted for three conditions: (i)for fully meeting the required heat load, (ii) for fully absorb-ing the allowable pressure drop in the cold stream, and (iii) forfully absorbing the allowable pressure drop in the hot stream.This type of plot is suitable for representing the design space,given the discrete nature of the plate geometrical characteris-tics, such as effective plate length and plate width The authorsalso presented applications of the use of bypasses as a designstrategy
Ex-The fourth contribution, “Heat Transfer of Supercritical CO2
Flow in Natural Convection Circulation System,” comes fromHideki Tokanai, Yu Ohtomo, Hiro Horiguchi, Eiji Harada, andMasafumi Kuriyama from the Department of Chemistry andChemical Engineering, Yamagata University, in Japan, andpresents measurements of heat transfer to supercritical CO2flow in a natural convection circulation system that consists of
a closed-loop circular pipe Systematic data of heat transfer efficients are given for various pressures and pipe diameters.They found that heat transfer coefficients of supercritical CO2
co-flow were very much higher compared to those of usually countered fluid flow and expressed them by a nondimensionalcorrelation equation proposed in their work They also presentednumerical model calculations of the velocity and temperaturedistributions in supercritical CO2flow to elucidate the exceed-ingly high value of heat transfer coefficient They concludedthat the heat transfer enhancement of supercritical CO2resultedfrom the high speed flow near the pipe wall This strong flowshows steep velocity and temperature gradients to enhance therate of heat transfer in the vicinity of the pipe wall
en-Zdenˇek Jegla, Bohuslav Kilkovsk´y, and Petr Stehl´ık, fromthe Institute of Process and Environmental Engineering, BrnoUniversity of Technology, the Czech Republic, deal with “Cal-culation Tool for Particulate Fouling Prevention of TubularHeat Transfer Equipment.” They studied fouling of heat trans-fer equipment in incineration plants They found that the mainprocess stream in such plants produced a stream of flue gas,and its thermal and physical properties significantly influenceoperating, maintenance, and investment costs of installed equip-ment and its service life Their contribution deals with the issue
of fouling mechanism at the heat transfer area of tubular heattransfer equipment installed in plants like these They presented
a mathematical model developed for fouling tendency tion and for prevention in design and operation of tubular heattransfer equipment designed for applications in the field of wasteheat transfer engineering vol 31 no 9 2010
Trang 4predic-J KLEME ˇS AND P STEHL´IK 709
incineration Obtained results were compared with
experimen-tal data published in worldwide available literature and a very
good agreement was found Their model is suitable for
equip-ment fouling tendency prediction and for prevention in design
and operating of tubular heat transfer equipment designed for
applications in waste incinerating plants The application for
design of the economizer demonstrates the contribution of a
de-veloped extended mathematical model to a complex analysis
The results of the developed extended model together with
tech-nical and economic analysis can contribute to selecting the most
suitable design alternative that can successfully satisfy
require-ments from several different points of view, such as fouling,
design, operation, and economics
The sixth paper comes from the University of Ottawa,
Canada, and its title is “Effect of High-Temperature Microwave
Irradiation on Municipal Thickened Waste Activated Sludge
Solubilization.” The authors are Isil Toreci, Kevin J Kennedy,
and Ronald L Droste They deal with sludge digestion and
stabilization Increasing hydrolysis by implementing
pretreat-ment prior to digestion can increase the digestion efficiency
They studied microwave pretreatment (MWP) as an alternative
to conventional thermal pretreatment They stated that MWP
above the boiling point has not been studied yet for sludge
solubilization and digestion Their paper provides preliminary
results on the effect of MWP conditions such as high
tempera-ture (110–175◦C), MWP intensity of 1.25 and 3.75◦C/min, and
sludge concentration of 6 and 11.85% on solubilization
The next paper deals with “Improvement of a Combustion
Unit Based on a Grate Furnace for Granular Dry Solid
Biofu-els Using CFD Methods.” The authors, Christian Jordan and
Michael Harasek, come from the Institute of Chemical
Engi-neering, Vienna University of Technology, in Austria They
studied the design and construction of an improved small-scale
combustion unit for various biofuels: wood, straw pellets, and
especially grain Using computational fluid dynamics (CFD)
methods and measurement data from a pilot unit, this study
contributes to the continuous enhancement of biomass firing
technology by addressing the commonly known problems
re-garding emissions and ash melting Based on the calculated
results, improvements for the existing prototype geometry have
been suggested and will be included in the design of a new
1.5-MW pilot-scale grate firing unit that was planned to start
operation by the beginning of 2009 Their future work will
deal with the detailed design of the prototype Plans for 2009
also included setting up a new grate furnace at a production
facility by Polytechnik GmbH and starting continuous
opera-tion by mid 2009 Detailed fuel analyses will be carried out
to close the mass and energy balances This will be followed
by further measurements for longer periods of stable
opera-tion and will provide a more reliable foundaopera-tion for validaopera-tion
of the simulation Additional CFD simulations will be done
for other fuels (e.g., grain) The introduction of a soot model,
fuel NOx, and a more detailed bed combustion model will be
considered
The eighth paper comes from the State University of NewYork College at Buffalo, New York, USA The authors, David J.Kukulka, Holly Czechowski, and Peter D Kukulka, evaluate thefeasibility of using surface coatings on commonly used processsurfaces to minimize/delay the effect of fouling They exploredstainless steel and copper with AgION and Xylan coatings Theyplaced sample plates vertically in test tanks and then exposedthem to untreated lake water for various time periods Theirresults compare surface roughness over time Additional resultsshow transient deposit weight gain The progressive change
in surface appearance with increasing immersion times is alsopresented and gives a visual representation of the surface at aspecific time Their review includes observations on the fouling
of coated process surfaces All coated samples showed somedeposit accumulation with no change in surface appearance forthe periods of immersion considered The authors summarizedresults of the material coatings for surfaces that are commonlyused in designs where fouling may be a concern Fouling rates,transient surface roughness values, and transient photographs
of the frontal surfaces of the materials were given for typicalconditions
The last paper, prepared by Zoe Anxionnaz, Michel sud, Christophe Gourdon, and Patrice Tochon, from ChemicalEngineering Laboratory, University of Toulouse/INPT, France,and Atomic Energy Commission–GRETh, Grenoble, France,has the title “Transposition of an Exothermic Reaction From
Cabas-a BCabas-atch ReCabas-actor to Cabas-an Intensified Continuous One.” The plementation of chemical syntheses in a batch or semi-batchreactor is generally limited by the removal or the supply of heat
im-A way to enhance thermal performances is to develop tifunctional devices like heat exchanger/reactors The authorsanalyzed a novel heat exchanger/reactor characterized in terms
mul-of residence time, pressure drop, and thermal behavior in order
to estimate capacities to perform an exothermic reaction: theoxidation of sodium thiosulfate by hydrogen peroxide Theirexperimental results highlighted the performances of the heatexchanger/reactor in terms of intensification, which allows theimplementation of the oxidation reaction at extreme operatingconditions They compared these conditions with a classicalbatch reactor The studied ShimTec reactor was a good example
of intensified unit and sustainable technology By combiningreaction and heat transfer, the process became safer, more envi-ronment friendly, and cheaper The future work will be aimed
at setting up reliable control system, design, scale-up, and mization procedures and safety studies
heat transfer engineering vol 31 no 9 2010
Trang 5[1] Klemeˇs, J., and Stehlik, P., PRES Conference—Challenges in
Complex Process Heat Transfer, Heat Transfer Engineering, vol.
23, pp 1–2, 2002
[2] Stehl´ık, P., and Klemeˇs, J., Selected Papers from the PRES 2002
Conference, Heat Transfer Engineering, vol 25, pp 1–3, 2004.
[3] Klemeˇs, J., and Stehl´ık, P., Selected Papers from the PRES 2003
Conference, Heat Transfer Engineering, vol 26, pp 1–3, 2005.
[4] Stehl´ık, P., and Klemeˇs, J., Recent Advances on Heat
Trans-fer Equipment Design and Optimization—Selected Papers from
PRES 2004 Conference, Heat Transfer Engineering, vol 27, pp.
1–3, 2006
[5] Stehl´ık, P., and Klemeˇs, J., Achievements in Applied Heat
Transfer—PRES 2006, Heat Transfer Engineering, vol 29, pp.
503–505, 2008
[6] Klemeˇs, J., and Pierucci, S., Emission Reduction by Process
In-tensification, Integration, P-Graphs, Micro CHP, Heat Pumps and
Advanced Case Studies, Applied Thermal Engineering, vol 28,
pp 2005–2010, 2008
[7] Klemeˇs, J., and Huisingh, D., Economic Use of Renewable
Re-sources, LCA, Cleaner Batch Processes and Minimising
Emis-sions and Wastewater, Journal of Cleaner Production, vol 16, pp.
159–163, 2008
[8] Bulatov, I., and Klemeˇs, J., Towards Cleaner Technologies:
Emis-sions Reduction, Energy and Waste Minimisation, Industrial
Im-plementation, Clean Technologies and Environmental Policy, vol.
Transac-Jiˇr´ı Klemeˇs is a P´olya Professor and EC Marie Curie
Chair Holder (EXC), Head of the Centre for Process Integration and Intensification (CPI 2 ) at the Univer- sity of Pannonia, Veszpr´em, in Hungary Previously
he worked for nearly 20 years in the Department of Process Integration and the Centre for Process Inte- gration at UMIST and after the merge at the Univer- sity of Manchester, UK, as a senior project officer and honorary reader He has many years of research and industrial experience In 1998 he founded and has been since the President of the International Conference “Process Integration, Mathematical Modeling, and Optimization for Energy Saving and Pollution Reduction—PRES.”
Petr Stehl´ık is a professor of process engineering
at the Brno University of Technology (UPEI—VUT Brno) and a director of the Institute of Process and Environmental Engineering He is also a member of the Presidium of the Czech Society of Chemical Engi- neers, and a member of renowned foreign engineering societies He had several years of experience in en- gineering practice before joining the university His research interests involve applied heat transfer, pro- cess design, mathematical modeling, energy saving, and environmental problems He is the author of numerous publications.
heat transfer engineering vol 31 no 9 2010
Trang 6CopyrightC Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457630903500833
Importance of Non-Boiling Two-Phase Flow Heat Transfer in Pipes for
Industrial Applications
AFSHIN J GHAJAR and CLEMENT C TANG
School of Mechanical and Aerospace Engineering, Oklahoma State University, Stillwater, Oklahoma, USA
The validity and limitations of the numerous two-phase non-boiling heat transfer correlations that have been published in the
literature over the past 50 years are discussed The extensive results of the recent developments in the non-boiling two-phase
heat transfer in air–water flow in horizontal and inclined pipes conducted at Oklahoma State University’s two-phase flow
heat transfer laboratory are presented Practical heat transfer correlations for a variety of gas–liquid flow patterns and
pipe inclination angles are recommended The application of these correlations in engineering practice and how they can
influence the equipment design and consequently the process design are discussed.
INTRODUCTION
In many industrial applications, such as the flow of oil and
natural gas in flow lines and well bores, the knowledge of
non-boiling two-phase, two-component (liquid and permanent gas)
heat transfer is required During the production of two-phase
hydrocarbon fluids from an oil reservoir to the surface, the
tem-perature of the hydrocarbon fluids changes due to the difference
in temperatures of the oil reservoir and the surface The change
in temperature results in heat transfer between the hydrocarbon
fluids and the earth surrounding the oil well, and the ability to
estimate the flowing temperature profile is necessary to address
several design problems in petroleum production engineering
[1]
In subsea oil and natural gas production, hydrocarbon fluids
may leave the reservoir with a temperature of 75◦C and flow in
subsea surrounding of 4◦C [2] As a result of the temperature
This is an extended version of the keynote paper presented at the 11th
Con-ference on Process Integration, Modeling and Optimization for Energy Saving
and Pollution Reduction (PRES2008), Prague, Czech Republic, August 24–28,
2008.
Generous contributions in equipment and software made by National
Instru-ments are gratefully acknowledged Sincere thanks are offered to Micro Motion
for generously donating one of the Coriolis flow meters and providing a
sub-stantial discount on the other one Thanks are also due to Martin Mabry for his
assistance in procuring these meters.
Address correspondence to Professor Afshin J Ghajar, School of Mechanical
and Aerospace Engineering, Oklahoma State University, Stillwater, OK 74078,
USA E-mail: afshin.ghajar@okstate.edu
gradient between the reservoir and the surrounding, the edge of heat transfer is critical to prevent gas hydrate and waxdeposition blockages [3] Wax deposition can result in problems,including reduction of inner pipe diameter causing blockage, in-creased surface roughness of pipe leading to restricted flow linepressure, decrease in production, and various mechanical prob-lems [4] Some examples of the economical losses caused bythe wax deposition blockages include: direct cost of removingthe blockage from a subsea pipeline was $5 million, productiondowntime loss in 40 days was $25 million [5], and the cost ofoil platform abandonment by Lasmo Company (UK) was $100million [6]
knowl-In situations where low-velocity flow is necessary whilehigh heat transfer rates are desirable, heat transfer enhance-ment schemes such as the coil-spring wire insert, twisted tapeinsert, and helical ribs are used to promote turbulence, thusenhancing heat transfer Although these heat transfer enhance-ment schemes have been proven to be effective, they do comewith drawbacks, such as fouling, increase in pressure drop, andsometimes even blockage Celata et al [7] presented an alterna-tive approach to enhance heat transfer in pipe flow, by injectinggas into liquid to promote turbulence In the experimental studyperformed by Celata et al [7], a uniformly heated vertical pipewas internally cooled by water, while heat transfer coefficientswith and without air injection were measured The introduction
of low air flow rate into the water flow resulted in increase ofthe heat transfer coefficient up to 20–40% for forced convection,and even larger heat transfer enhancement for mixed convection[7]
711
Trang 7Two-phase flow can also occur in various situations related
to ongoing and planned space operations, and the understanding
of heat transfer characteristics is important for designing piping
systems for space operations limited by size constraints [8] To
investigate heat transfer in two-phase slug and annular flows
under reduced gravity conditions, Fore et al [8, 9] conducted
heat transfer measurements for air–water and air–50% aqueous
glycerin aboard NASA’s Zero-G KC-135 aircraft
Due to limited studies available in the literature, Wang et al
[10] investigated forced convection heat transfer on the shell side
of a TEMA-F horizontal heat exchanger using a 60% aqueous
glycerin and air mixture Their work resulted in
recommenda-tion of correlarecommenda-tions for two-phase heat transfer coefficient in
stratified, intermittent, and annular flows in shell-and-tube heat
exchangers
In this article, an overview of our ongoing research on this
topic that has been conducted at our heat transfer laboratory
over the past several years is presented Our extensive literature
search revealed that numerous heat transfer coefficient
correla-tions have been published over the past 50 years We also found
several experimental data sets for forced convective heat transfer
during gas–liquid two-phase flow in vertical pipes, very limited
data for horizontal pipes, and no data for inclined pipes
How-ever, the available correlations for two-phase convective heat
transfer were developed based on limited experimental data and
are only applicable to certain flow patterns and fluid
combina-tions
The overall objective of our research has been to develop a
heat transfer correlation that is robust enough to span all or most
of the fluid combinations, flow patterns, flow regimes, and pipe
orientations (vertical, inclined, and horizontal) To this end, we
have constructed a state-of-the-art experimental facility for
sys-tematic heat transfer data collection in horizontal and inclined
positions (up to 7◦) The experimental setup is also capable of
producing a variety of flow patterns and is equipped with two
transparent sections at the inlet and exit of the test section for
in-depth flow visualization In this article we present the
high-lights of our extensive literature search, the development of our
proposed heat transfer correlation and its application to
experi-mental data in horizontal, inclined, and vertical pipes, a detailed
description of our experimental setup, the flow visualization
results for different flow patterns, the experimental results for
various flow patterns, and our proposed heat transfer correlation
for various flow patterns and pipe orientations
COMPARISON OF 20 TWO-PHASE HEAT TRANSFER
CORRELATIONS WITH SEVEN SETS OF
EXPERIMENTAL DATA
Numerous heat transfer correlations and experimental data
for forced convective heat transfer during gas–liquid two-phase
flow in vertical and horizontal pipes have been published over
the past 50 years In a study published by Kim et al [11], a
comprehensive literature search was carried out and a total of
38 two-phase flow heat transfer correlations were identified.The validity of these correlations and their ranges of applica-bility have been documented by the original authors In mostcases, the identified heat transfer correlations were based on asmall set of experimental data with a limited range of variablesand gas–liquid combinations In order to assess the validity ofthose correlations, they were compared against seven extensivesets of two-phase flow heat transfer experimental data availablefrom the literature, for vertical and horizontal tubes and dif-ferent flow patterns and fluids For consistency, the validity ofthe identified heat transfer correlations were based on the com-parison between the predicted and experimental two-phase heattransfer coefficients meeting the±30% criterion
In total, 524 data points from the five available experimentalstudies [12–16] were used for these comparisons (see Table 1).The experimental data included five different gas–liquid com-binations (air–water, air–glycerin, air–silicone, helium–water,Freon 12–water), and covered a wide range of variables, includ-ing liquid and gas flow rates and properties, flow patterns, pipesizes, and pipe inclination Five of these experimental data setsare concerned with a wide variety of flow patterns in verticalpipes and the other two data sets are for limited flow patterns(slug and annular) within horizontal pipes
Table 2 shows 20 of the 38 heat transfer correlations [14,16–35] that were identified and reported by Kim et al [11].Eighteen of the two-phase flow heat transfer correlations werenot tested, since the required information for those correlationswas not available through the identified experimental studies
In assessing the ability of the 20 identified heat transfer lations, their predictions were compared with the experimentaldata from the sources listed in Table 1, both with and withoutconsidering the restrictions on ReSLand VSG/VSLaccompany-ing the correlations The results from comparing the 20 heattransfer correlations and the experimental data are summarized
corre-in Table 3 for major flow patterns corre-in vertical pipes
There were no remarkable differences for the tions of the heat transfer correlations based on the results withand without the restrictions on ReSLand VSG/VSL, except for thecorrelations of Chu and Jones [18] and Ravipudi and Godbold[25], as applied to the air–water experimental data of Vijay [12].Details of this discussion can be found in Kim et al [11].Based on the results without the authors’ restrictions on
recommenda-ReSLand VSG/VSL, the correlation of Chu and Jones [18] was
Table 1 The experimental data used in Kim et al [11]
Source Orientation Fluids Number of data points
Rezkallah [13] Vertical Air–silicone 162 Aggour [14] Vertical Helium–water 53 Aggour [14] Vertical Freon 12–water 44 Pletcher [15] Horizontal Air–water 48
heat transfer engineering vol 31 no 9 2010
Trang 9Table 3 Recommended correlations for vertical pipes, Kim et al [11]
Air–water Air–glycerin Air–silicone Helium–water Freon 12–water Air–water
Note.√= Recommended correlation with and without restrictions Shaded cells indicate the correlations that best satisfied the ±30% two-phase heat transfercoefficient criterion A = annular, B = bubbly, C = churn, F = froth, S = slug.
recommended for only annular, bubbly-froth, slug-annular, and
froth-annular flow patterns of air–water in vertical pipes While
the correlation of Ravipudi and Godbold [25] was recommended
for only annular, slug-annular, and froth-annular flow patterns
of air–water in vertical pipes
However, when considering the ReSL and VSG/VSL
restric-tions by the authors, the correlation of Chu and Jones [18] was
recommended for all vertical pipe air–water flow patterns
in-cluding transitional flow patterns, except the annular-mist flow
pattern While the correlation of Ravipudi and Godbold [25]
was recommended for slug, froth, and annular flow patterns and
for all of the transitional flow patterns of the vertical air–water
experimental data
All of the correlations just recommended have the following
important parameters in common: ReSL, PrL, µB/µW, and either
void fraction (α) or superficial velocity ratio (VSG/VSL) It
ap-pears that void fraction and superficial velocity ratio, although
not directly related, may serve the same function in two-phase
flow heat transfer correlations
From the comprehensive literature search, Kim et al [11]
found that there is no single correlation capable of predicting the
flow for all fluid combinations in vertical pipes In the following
section, the effort of Kim et al [36] in developing a heat transfer
correlation that is robust enough to span all or most of the fluid
combinations and flow patterns for vertical pipes is highlighted
Kim et al [36] developed a correlation that is capable of
predicting heat transfer coefficient in two-phase flow regardless
of fluid combinations and flow patterns The correlation uses
a carefully derived heat transfer model that takes into account
the appropriate contributions of both the liquid and gas phases
using the respective cross-sectional areas occupied by the two
phases
DEVELOPMENT OF THE HEAT TRANSFER
CORRELATION FOR VERTICAL PIPES
The void fraction (α) is defined as the ratio of the
gas-flow cross-sectional area (AG) to the total cross-sectional area,
A (= AG+ AL):
AG+ AL
(1)The actual gas velocity VGcan be calculated from
Based upon this correlation, the single-phase heat transfercoefficients in Eq (4), hLand hG, can be modeled as functions
of Reynolds number, Prandtl number, and the ratio of bulk towall viscosities Thus, Eq (4) can be expressed as:
(µB/µW)L
(6)heat transfer engineering vol 31 no 9 2010
Trang 10A J GHAJAR AND C C TANG 715
Substituting the definition of Reynolds number (Re =
ρVD/µB) for the gas (ReG) and liquid (ReL) yields
(µB/µW)L
(7)Rearranging yields
(µW)G
(8)
where the assumption has been made that the bulk viscosity ratio
in the Reynolds number term of Eq (7) is exactly canceled by
the last term in Eq (7), which includes the same bulk viscosity
ratio Substituting Eq (1) for the ratio of gas-to-liquid diameters
(DG/DL) in Eq (8) and based upon practical considerations
assuming that the ratio of liquid-to-gas viscosities evaluated at
the wall temperature [(µW)L/(µW)G] is comparable to the ratio
of those viscosities evaluated at the bulk temperature (µL/µG),
Further simplifying Eq (9), combine Eqs (2) and (3) for VG
(gas velocity) and VL(liquid velocity) to get the ratio of VG/VL
and substitute into Eq (9) to get
hTP= (1 − α)hL
1+ fctn
x
1− x
,
α
Assuming that two-phase heat transfer coefficient can be
expressed using a power-law relationship on the individual
pa-rameters that appear in Eq (10), then it can be expressed as:
hTP = (1 − α)hL
1+ C
x
1− x
mα
as commonly used in the correlations of the available literature[11]:
ReL=
ρVDµ
The values of the void fraction (α) used in Eq (11) eitherwere taken directly from the original experimental data sets (ifavailable) or were calculated based on the equation provided byChisholm [37], which can be expressed as
Table 4 Results of the predictions for available two-phase heat transfer experimental data using Eq (11), Kim et al [36]
Values of constant and exponents RMS Mean Number of Range of parameters
deviation deviation data
127,000
14 to 209,000
9.99 × 10 −3to
137 × 10 −3 3.64× 10 −3to
23.7 × 10 −3
heat transfer engineering vol 31 no 9 2010
Trang 11Figure 1 Comparison of the predictions by Eq (11) with the experimental
data for vertical flow (255 data points), Kim et al [36].
HEAT TRANSFER CORRELATION FOR GAS–LIQUID
FLOW IN VERTICAL PIPES
To determine the values of leading coefficient and the
expo-nents in Eq (11), four sets of experimental data (see the first
column in Table 4) for vertical pipe flow were used The ranges
of these four sets of experimental data can be found in Kim et al
[11] The experimental data (a total of 255 data points) included
four different gas–liquid combinations (air–water, air–silicone,
helium–water, Freon 12–water) and covered a wide range of
variables, including liquid and gas flow rates, properties, and
flow patterns
The selected experimental data were only for turbulent
two-phase heat transfer data in which the superficial Reynolds
num-bers of the liquid (ReSL) were all greater than 4000 Table 4 and
Figure 1 provide the details of the correlation and how well the
proposed correlation predicted the experimental data
The two-phase heat transfer correlation, Eq (11), predicted
the heat transfer coefficients of 255 experimental data points for
vertical flow with an overall mean deviation of about 2.5% and
a root-mean-square deviation of about 12.8% About 83% of
the data (212 data points) were predicted with less than±15%
deviation, and about 96% of the data (245 data points) were
predicted with less than ±30% deviation The results clearly
show that the proposed heat transfer correlation is robust and
can be applied to turbulent gas–liquid flow in vertical pipes with
different flow patterns and fluid combinations
A GENERAL TWO-PHASE HEAT TRANSFER
CORRELATION FOR VARIOUS FLOW PATTERNS AND
PIPE INCLINATIONS
The heat transfer correlation developed by Kim et al [36],
Eq (11), was meant for predicting heat transfer rate in
two-phase flow in vertical pipes In order to handle the effects of
Gas-Liquid Interface at Equilibrium State
Realistic Gas-Liquid Interface
SL ,e
S
L
Figure 2 Gas–liquid interfaces and wetted perimeters.
various flow patterns and inclination angles on the two-phaseheat transfer data with only one correlation, Ghajar and Kim[38] and Kim and Ghajar [39] introduced the flow pattern factor(FP) and the inclination factor (I)
The void fraction (α), which is the volume fraction of thegas phase in the tube cross-sectional area, does not reflect theactual wetted perimeter (SL) in the tube with respect to the cor-responding flow pattern For instance, the void fraction and thenondimensionalized wetted perimeter of annular flow both ap-proach unity, but in the case of plug flow the void fraction is nearzero and the wetted perimeter is near unity However, the esti-mation of the actual wetted perimeter is very difficult due to thecontinuous interaction of the two phases in the tube Therefore,instead of estimating the actual wetted perimeter, modeling theeffective wetted perimeter is a more practical approach In theirmodel, Ghajar and his co-workers have ignored the influence ofthe surface tension and the contact angle of each phase on theeffective wetted perimeter The wetted perimeter at the equilib-rium state, which can be calculated from the void fraction, is
of the equilibrium wetted perimeter, Eq (15), is proposed:
as the shape factor, and in essence is a modified and normalizedheat transfer engineering vol 31 no 9 2010
Trang 12A J GHAJAR AND C C TANG 717
Froude number The shape factor (FS) is defined as
≥ 1, which is common in gas–liquid flow, and
represents the shape changes of the gas–liquid interface by the
force acting on the interface due to the relative momentum and
gravitational forces
Due to the density difference between gas and liquid, the
liquid phase is much more affected by the orientation of the
pipe (inclination) A detailed discussion of the inclination effect
on the two-phase heat transfer is available in Ghajar and Tang
[40] In order to account for the effect of inclination, Ghajar and
Kim [38] proposed the inclination factor
I= 1 + g D
ρL− ρG
sin θ
ρLV2 SL
(18)
where the term [g D (ρ L− ρG ) sin(θ)]/[ρ L V2
SL] represents therelative force acting on the liquid phase in the flow direction due
to the momentum and the buoyancy forces
Now, introduce the two proposed factors for the flow pattern
(FP) and inclination (I) effects into our heat transfer correlation,
Eq (11) Substituting (FP) for (1−α), which is the leading
co-efficient of (hL), and introducing (I) as an additional power-law
term in Eq (11), the two-phase heat transfer correlation becomes
where (hL) comes from the Sieder and Tate [35] correlation for
turbulent flow [see Eq (12)] For the Reynolds number needed
in the (hL) calculation, Eq (13), presented and discussed earlier,
was used The values of the void fraction (α) used in Eqs (13),
(16), and (19) were calculated based on the correlation provided
by Woldesemayat and Ghajar [41], which can be expressed as
C0(VSG+ VSL)+ uGM
(20)where the distribution parameter (C0) and the drift velocity of
gas (uGM) are given as
uGM = 2.9(1.22 + 1.22 sin θ)(P atm/P sys )
×
gDσ (1+ cos θ)ρL− ρG
ρ2 L
0.25
Note that the leading constant value of 2.9 in the preceding
equation for the drift flux velocity (uGM) carries a unit of m−0.25,and Eq (20) should be used with SI units
Other void fraction correlations could also be used in place
of the Woldesemayat and Ghajar [41] correlation Tang andGhajar [42] showed that Eq (19) has such robustness that itcan be applied with different void fraction correlations Thedifference resulting from the use of different correlations will beabsorbed during the determination of the values of the constantand exponents of Eq (19)
The two-phase heat transfer correlation, Eq (19), was dated with a total of 763 experimental data points for differentflow patterns and inclination angles [39, 42, 43] Overall, thecorrelation, Eq (19), has successfully predicted over 85% of theexperimental data points to within±30% for 0◦, 2◦, 5◦, and 7◦
vali-pipe orientations
However, upon revisiting the two-phase heat transfer relation, Eq (19), along with the equations for flow patternfactor (FP), Eq (16), and inclination factor (I), Eq (18), it wasrealized that the correlation has not accounted for the surfacetension force Since surface tension is a variable that can affectthe hydrodynamics of gas–liquid two-phase flow, it is sensible
cor-to include the surface tension incor-to the correlation To do that,the equation for the inclination factor (I), Eq (18), is modified.The modified inclination factor takes on the following form:
hTP= FPhL
1+ C
x
of the constant and exponents are discussed in a later section.heat transfer engineering vol 31 no 9 2010
Trang 13Table 5 Summary of experimental database sources, Woldesemayat and Ghajar [41]
Source Physical flow configuration/characteristics Mixture considered Measurement technique Number of data points
Beggs [45] Horizontal, uphill, and vertical, D = 25.4 mm,
and 38 1 mm
Spedding and Nguyen [46] Horizontal, uphill, and vertical, D = 45.5 mm Air–water Quick-closing valves 1383
Mukherjee [47] Horizontal, uphill, and vertical, D = 38.1 mm Air–kerosene Capacitance probes 558
Minami and Brill [48] Horizontal, D = 77.93 mm Air–water and air–kerosene Quick-closing valves 54 and 57
COMPARISON OF VOID FRACTION CORRELATIONS
FOR DIFFERENT FLOW PATTERNS AND PIPE
INCLINATIONS
Due to the importance of void fraction in influencing the
characteristics of two-phase flow in pipes, Woldesemayat and
Ghajar [41] conducted a very extensive comparison of 68 void
fraction correlations available in the open literature against 2845
experimental data points The experimental data points were
compiled from various sources with different experimental
fa-cilities [44–51] Out of the 2845 experimental data points, 900
were for horizontal, 1542 for inclined, and 403 for vertical pipe
orientations (see Table 5)
Based on the comparison with experimental data, six void
fraction correlations [52–57] were recommended for
accept-ably predicting void fraction for horizontal, upward inclined,
and vertical pipe orientations regardless of flow patterns The
percentage of data points correctly predicted for the 2845
exper-imental data points within three error bands for each correlation
is summarized in Table 6
Among the six void fraction correlations listed in Table 6,
Dix [53] showed better performance The correlation by Dix
[53] has the following expression:
C0(VSG+ VSL)+ uGM
(24)
Table 6 Number and percentage of data points correctly predicted by the
six recommended void fraction correlations and Eq (20) for the entire
experimental database summarized in Table 5, Woldesemayat and Ghajar
0.25
Figure 3 shows the performance of the void fraction tion by Dix [53], Eq (24) Woldesemayat and Ghajar [41] pro-posed an improved void fraction correlation, Eq (20), that givesbetter predictions when compared with available experimentaldata The performance of Eq (20) on the 2845 experimental datapoints in comparison with the recommended six void fractioncorrelations is also summarized in Table 6
correla-As shown in Table 6, the void fraction correlation, Eq (20),introduced by Woldesemayat and Ghajar [41] gives noticeableimprovements over the other six correlations The results of thecomparison for Eq (20) with the 2845 experimental data pointsare also illustrated in Figure 4 Both Table 6 and Figure 4 show
0.0 0.2 0.4 0.6 0.8 1.0
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Figure 3 Comparison of void fraction correlation by Dix [53], Eq (24), with 2845 experimental data points summarized in Table 5, Woldesemayat and Ghajar [41].
heat transfer engineering vol 31 no 9 2010
Trang 14A J GHAJAR AND C C TANG 719
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Figure 4 Comparison of void fraction correlation by Woldesemayat and
Ghajar [41], Eq (20), with 2845 experimental data points summarized in Table
5, Woldesemayat and Ghajar [41].
the capability and robustness of Eq (20) to successfully predict
void fraction for various pipe sizes, inclinations, and two-phase
fluid mixtures from various sources with different experimental
facilities The benefit of comparing with experimental data from
different facilities is the minimization of sample bias
EXPERIMENTAL SETUP AND DATA REDUCTION FOR
HORIZONTAL AND SLIGHTLY UPWARD INCLINED
PIPE FLOW
A schematic diagram of the overall experimental setup for
heat transfer measurements is shown in Figure 5 The test section
is a 27.9 mm inner diameter (I.D.) straight standard stainless
steel schedule 10S pipe with a length to diameter ratio of 95 The
setup rests atop a 9 m long aluminum I-beam that is supported
by a pivoting foot and a stationary foot that incorporates a small
electric screw jack
In order to apply uniform wall heat flux boundary condition
to the test section, copper plates were silver soldered to the
inlet and exit of the test section The uniform wall heat flux
boundary condition was maintained by a Lincoln SA-750 welder
for ReSL > 2000 and a Miller Maxtron 450 DC welder for
ReSL<2000 The entire length of the test section was wrapped
using fiberglass pipe wrap insulation, followed by a thin polymer
vapor seal to prevent moisture penetration The calming section
(clear polycarbonate pipe with 25.4 mm I.D and L/D= 88)
served as a flow developing and turbulence reduction device
and flow pattern observation section
T-type thermocouple wires were cemented with Omegabond
101, an epoxy adhesive with high thermal conductivity and
electrical resistivity, on the outside wall of the stainless steel
test section as shown in Figure 6 Thermocouples were placed
on the outer surface of the pipe wall at uniform intervals of
254 mm from the entrance to the exit of the test section There
were 10 thermocouple stations in the test section (refer to ure 6) All the thermocouples were monitored with a NationalInstruments data acquisition system The average system sta-bilization time period was from 30 to 60 min after the systemattained steady state The inlet liquid and gas temperatures andthe exit bulk temperature were measured by Omega TMQSS-125U-6 thermocouple probes Calibration of thermocouples andthermocouple probes showed that they were accurate to within
Fig-±0.5◦C The operating pressures inside the experimental setup
were monitored with a pressure transducer To ensure a uniformfluid bulk temperature at the inlet and exit of the test section,
a mixing well of alternating polypropylene baffle type staticmixer for both gas and liquid phases was utilized The outletbulk temperature was measured immediately after the mixingwell
The fluids used in the test loop are air and water The water
is distilled and stored in a 55-gal cylindrical polyethylene tank
A Bell & Gosset series 1535 coupled centrifugal pump wasused to pump the water through an Aqua-Pure AP12T water fil-ter An ITT Standard model BCF 4063 one-shell and two-tubepass heat exchanger removes the pump heat and the heat addedduring the test to maintain a constant inlet water temperature.From the heat exchanger, the water passes through a Micro Mo-tion Coriolis flow meter (model CMF100) connected to a digitalField-Mount Transmitter (model RFT9739) that conditions theflow information for the data acquisition system From the Cori-olis flow meter it then flows into the test section Air is suppliedvia an Ingersoll-Rand T30 (model 2545) industrial air compres-sor The air passes through a copper coil submerged in a vessel
of water to lower the temperature of the air to room temperature.The air is then filtered and condensation is removed in a coalesc-ing filter The air flow is measured by a Micro Motion Coriolisflow meter (model CMF025) connected to a digital Field-MountTransmitter (model RFT9739) and regulated by a needle valve.Air is delivered to the test section by flexible tubing The waterand air mixture is returned to the reservoir, where it is separatedand the water is recycled
The heat transfer measurements at uniform wall heat fluxboundary condition were carried out by measuring the localoutside wall temperatures at 10 stations along the axis of thepipe and the inlet and outlet bulk temperatures, in addition toother measurements such as the flow rates of gas and liquid,room temperature, voltage drop across the test section, and cur-rent carried by the test section A National Instruments dataacquisition system was used to record and store the data mea-sured during these experiments The computer interface used torecord the data is a LabVIEW Virtual Instrument (VI) programwritten for this specific application
The peripheral heat transfer coefficient (local average) wascalculated based on the knowledge of the pipe inside wall sur-face temperature and inside wall heat flux obtained from a datareduction program developed exclusively for this type of exper-iment [58] The local average peripheral values for inside walltemperature, inside wall heat flux, and heat transfer coefficientwere then obtained by averaging all the appropriate individualheat transfer engineering vol 31 no 9 2010
Trang 15Figure 5 Schematic of experimental setup.
local peripheral values at each axial location The variation in
the circumferential wall temperature distribution, which is
typ-ical for two-phase gas–liquid flow in horizontal pipes, leads
to different heat transfer coefficients depending on which
cir-cumferential wall temperature is selected for the calculations
In two-phase heat transfer experiments, in order to overcome
the unbalanced circumferential heat transfer coefficients and to
get a representative heat transfer coefficient for a test run, the
following equation was used to calculate an overall two-phase
heat transfer coefficient (hTP EXP) for each test run:
hTPEXP = 1
L
¯h dz= 1L
where L is the length of the test section, and ¯h, ¯˙q , ¯Tw, and TB
are the local mean heat transfer coefficient, the local mean heat
flux, the local mean wall temperature, and the bulk temperature
at a thermocouple station, respectively; k is the index of thethermocouple stations, NST is the number of the thermocouple
stations, z is the axial coordinate, and z is the element length
of each thermocouple station The data reduction program used
a finite-difference formulation to determine the inside wall perature and the inside wall heat flux from measurements of theoutside wall temperature, the heat generation within the pipewall, and the thermophysical properties of the pipe material(electrical resistivity and thermal conductivity)
tem-The reliability of the flow circulation system and of the perimental procedures was checked by making several single-phase calibration runs with distilled water The single-phaseheat transfer experimental data were checked against the well-established single-phase heat transfer correlations [59] in theReynolds number range from 3000 to 30,000 In most instances,the majority of the experimental results were well within±10%
ex-of the predicted results [59, 60]
The uncertainty analysis of the overall experimental dures using the method of Kline and McClintock [61] showedthat there is a maximum of 11.5% uncertainty for heat transfercoefficient calculations Experiments under the same conditionsheat transfer engineering vol 31 no 9 2010
Trang 16proce-A J GHAJAR AND C C TANG 721
Tail of Flow Direction
1727 cm
Pressure Tap Hole
Figure 6 Test section.
were conducted periodically to ensure the repeatability of the
results The maximum difference between the duplicated
exper-imental runs was within±10%
FLOW PATTERNS
The various interpretations accorded to the multitude of flow
patterns by different investigators are subjective; no uniform
procedure exists at present for describing and classifying them
In this study, the flow pattern identification for the experimental
data was based on the procedures suggested by Taitel and Dukler
[62] and by Kim and Ghajar [59], and on visual observations
as deemed appropriate All observations for the flow pattern
judgments were made at the clear polycarbonate observation
sections before and after the stainless steel test section (see
Figure 5) By fixing the water flow rate, flow patterns were
observed by varying air flow rates
Flow pattern data were obtained at isothermal condition with
the pipe in horizontal position and at 2◦, 5◦, and 7◦ inclined
positions These experimental data were plotted and compared
using their corresponding values of ReSG and ReSL and the
flow patterns Representative digital images of each flow pattern
were taken using a Nikon D50 digital camera with Nikkor 50
mm f/1.8D lens Figure 7 shows the flow map for horizontalflow with the representative photographs of the various flowpatterns The various flow patterns for horizontal flow depicted
in Figure 7 show the capability of our experimental setup tocover a multitude of flow patterns The shaded regions representthe transition boundaries of the observed flow patterns.The influence of small inclination angles of 2◦, 5◦, and 7◦onthe observed flow patterns is shown in Figure 8 As shown in this
Figure 7 Flow map for horizontal flow with representative photographs of flow patterns.
heat transfer engineering vol 31 no 9 2010
Trang 17Figure 8 Change of flow pattern transition boundaries as pipe inclined
up-ward from horizontal position.
figure, the flow pattern transition boundaries for horizontal flow
were found to be quite different from the flow pattern transition
boundaries for inclined flow when slight inclinations of 2◦, 5◦,
and 7◦were introduced The changes in the flow pattern
tran-sition boundaries from horizontal to slightly inclined flow are
the transition boundaries for stratified flow and slug/wavy flow
When the pipe was inclined from horizontal to slight inclination
angles of 2◦, 5◦, and 7◦, the stratified flow region was replaced
by slug flow and slug/wavy flow for ReSG < 4000 and 4000 <
ReSG <10000, respectively
Other shifts in the flow pattern transition boundaries were
ob-served in the plug-to-slug boundary and the slug-to-slug/bubbly
boundary In these two cases, the flow pattern transition
bound-aries were observed to be shifted slightly to the upper left
direc-tion as inclinadirec-tion angles were slightly increased from horizontal
to 7◦ For slightly inclined flow of 2◦, 5◦, and 7◦, there were no
drastic changes in the flow pattern transition boundaries
For verification of the flow pattern map, flow patterns data
from Barnea et al [63] were used and compared with the flow
pattern maps for horizontal and 2◦ inclined pipe Using flow
pattern data from Barnea et al [63] for air–water flow in 25.5
mm horizontal pipe, the data points plotted on the flow map for
horizontal flow (see Figure 7) are illustrated in Figure 9
The comparison between the data points from Barnea et al
[63] and the flow pattern map for horizontal flow showed very
satisfactory agreement, especially among the distinctive major
flow patterns such as annular, slug, and stratified It should be
noted that Barnea et al [63] had successfully compared their
horizontal flow pattern data with the flow map proposed by
Mandhane et al [64]
In a similar manner, using flow pattern data from Barnea
et al [63] for air–water flow in 25.5 mm 2◦ inclined pipe, the
data points plotted on the flow map for 2◦ inclined flow (see
1000 10000
Annular Elongated bubble Slug Stratified smooth Stratified wavy
Wavy
Annular
Slug Plug
of the comparable flow patterns in the horizontal position Forexample, it was observed that the slug flow patterns in the in-clined positions of 5◦ and 7◦ have reverse flow between slugsdue to the gravitational force, which can have a significant effect
on the heat transfer To understand the influence of flow patterns
on heat transfer, systematic measurement of heat transfer datawere conducted Table 7 and Figure 11 illustrate the number
of two-phase heat transfer data points systematically measuredfor different flow patterns and test section orientations Heat
1000 10000
Annular Elongated bubble Slug Stratified wavy
Slug/Wavy Plug
Wavy Annular/Bubbly/Slug
Figure 10 Flow patterns data points from Barnea et al [63] plotted on the flow map for 2 ◦inclined flow (see Figure 8).
heat transfer engineering vol 31 no 9 2010
Trang 18A J GHAJAR AND C C TANG 723
Table 7 Number of two-phase heat transfer data points measured for different
flow patterns and pipe orientations
Test section orientation Flow patterns Horizontal 2 ◦inclined 5◦inclined 7◦inclined
transfer data at low air and water flow rates (ReSG <500 and
ReSL <700) were not collected At such low air and water
flow rates, there exists the possibility of local boiling or dry-out,
which could potentially damage the heated test section
SYSTEMATIC INVESTIGATION ON TWO-PHASE
GAS–LIQUID HEAT TRANSFER IN HORIZONTAL AND
SLIGHTLY UPWARD INCLINED PIPE FLOWS
In this section, an overview of the different trends that have
been observed in the heat transfer behavior of the two-phase
air–water flow in horizontal and inclined pipes for various flow
patterns is presented The non-boiling two-phase heat transfer
data were obtained by systematically varying the air and water
flow rates and the pipe inclination angle The summary of the
experimental conditions and measured heat transfer coefficients
are tabulated in Table 8 Detailed discussions on the complete
experimental results are documented by Ghajar and Tang [40]
Figures 12 and 13 provide an overview of the pronounced
influence of the flow pattern, superficial liquid Reynolds number
(water flow rate) and superficial gas Reynolds number (air flow
rate) on the two-phase heat transfer coefficient in horizontal flow
The results presented in Figure 12 clearly show that two-phase
heat transfer coefficient is strongly influenced by the superficial
liquid Reynolds number (ReSL)
As shown in Figure 12, the heat transfer coefficient increases
proportionally as ReSLincreases In addition, for a fixed ReSL,
Table 8 Summary of experimental conditions and measured two-phase heat
COMPARISON OF GENERAL HEAT TRANSFER CORRELATION WITH EXPERIMENTAL RESULTS FOR VARIOUS FLOW PATTERNS AND PIPE INCLINATIONS
The two-phase heat transfer correlation, Eq (19), was idated with a total of 763 experimental data points for differ-ent flow patterns in horizontal and slightly inclined air–watertwo-phase pipe flows [39, 42, 43] Equation (19) performed rel-atively well by predicting over 85% of the experimental datapoints to within±30% for 0◦, 2◦, 5◦, and 7◦pipe orientations.
val-Recently, Franca et al [65] compared their mechanistic modeldeveloped for convective heat transfer in gas–liquid intermittent(slug) flows with the general heat transfer correlation proposed
in this study For void fraction, Franca et al [65] used their ownexperimental data, which were obtained for air–water flow in a
15 m long, 25.4 mm inside diameter copper pipe When paring their mechanistic model with Eq (19), the agreement iswithin±15%, which is considered to be excellent
com-However, when comparing the heat transfer correlation, Eq.(19), with data from vertical pipes and different gas–liquid com-binations, Eq (19) has shown some inadequacy in its perfor-mance Equation (19) was validated with 986 experimental datapoints for different flow patterns, inclination angles, and gas–liquid combinations The 986 experimental data points werecompiled from various sources with different experimental fa-cilities (see Table 9) with a wide range of superficial gas andliquid Reynolds numbers (750 ≤ ReSL ≤ 127,000 and 14 ≤
ReSG ≤ 209,000) and inclination angles (0◦ ≤ θ ≤ 90◦).
Figure 14 shows the comparison of Eq (19) with all 986experimental data points for different inclination angles andgas–liquid combinations
Figure 14 shows that Eq (19) performed well for two-phaseflow with heat transfer coefficient between 1000 W/m2-K and
5000 W/m2-K However, Eq (19) has shown some inadequacy
in predicting two-phase flow with heat transfer coefficients low 1000 W/m2-K and above 5000 W/m2-K Overall, Eq (19)successfully predicted 83% of the 986 experimental data pointswithin ±30% agreement (see Table 9) The results shown inTable 9 and Figure 14 prompted further investigation and im-provements were made on Eq (19)
be-As discussed previously, improvements on Eq (19) weremade by modifying the inclination factor (I), Eq (18) Themodified inclination factor (I∗), Eq (21), which includes theE¨otv¨os number (Eo) to represent the hydrodynamic interactionheat transfer engineering vol 31 no 9 2010
Trang 19Figure 11 Flow maps for horizontal, 2 ◦, 5◦, and 7◦inclined flows with distribution of heat transfer data collected.
Figure 12 Variation of two-phase heat transfer coefficient with superficial
liquid Reynolds number in horizontal flow.
Figure 13 Variation of two-phase heat transfer coefficient with superficial gas Reynolds number in horizontal flow.
heat transfer engineering vol 31 no 9 2010
Trang 20A J GHAJAR AND C C TANG 725
Table 9 Results of the predictions for 986 experimental heat transfer data points with different gas–liquid combinations and inclination angles by using Eq (19)
deviation data points data points data points deviation
All 986 data points,
0 ◦ ≤ θ ≤ 90 ◦ 33.1 649 (66%) 746 (76%) 817 (83%) −16.9 to 30.8 750 to 127,000 14 to 209,000 9.99 × 10 −3
to 148 × 10 −3 3.64× 10 −3to
26.3 × 10 −3 Air–water (θ = 0 ◦), 160
data points [40]
43.4 124 (67%) 137 (74%) 150 (82%) −15.9 to 64.5 780 to 26,000 600 to 48,000 Air–water (θ = 7 ◦), 187
data points [40]
44.7 110 (59%) 132 (71%) 149 (80%) −16.3 to 74.7 770 to 26,000 560 to 47,000 Air–water (θ = 90 ◦),
105 data points [12]
25.0 67 (64%) 79 (75%) 85 (81%) −22.3 to 2.4 4000 to 127,000 43 to 154,000 Air–silicone (θ = 90 ◦),
56 data points [13]
5.9 56 (100%) 56 (100%) 56 (100%) −4.6 to 6.1 8400 to 21,000 52 to 42,000 Helium–water (θ = 90 ◦),
50 data points [14]
25.4 22 (44%) 31 (62%) 37 (74%) −25.9 to 6.9 4000 to 126,000 14 to 13,000 Freon 12–water (θ =
90 ◦), 44 data points
[14]
39.1 16 (36%) 17 (39%) 18 (41%) −33.3 to 0 4200 to 55,000 860 to 209,000
Note Values of constant and exponents: C= 0.82, m = 0.08, n = 0.39, p = 0.03, q = 0.01, and r = 0.40.
of buoyancy and surface tension forces, replaced the inclination
factor (I) and resulted in a generalized two-phase heat transfer
correlation for various pipe inclinations and gas–liquid
combi-nations, Eq (23)
With the proposed constant and exponents, C= 0.55, m =
0.1, n= 0.4, and p = q = r = 0.25, Eq (23) was
success-fully validated with a total of 986 experimental data points for
different flow patterns, inclination angles, and gas–liquid
com-binations The 986 experimental data points were compiled from
various sources with different experimental facilities (see Table
10) with a wide range of superficial gas and liquid Reynolds
Figure 14 Comparison of the predictions by Eq (19) with all 986
experi-mental data points for different inclination angles and gas–liquid combinations
(see Table 9).
numbers (750≤ ReSL ≤ 127,000 and 14 ≤ ReSG ≤ 209,000)and inclination angles (0◦≤ θ ≤ 90◦).
As summarized in Table 10, the comparison of the predictions
by the general two-phase heat transfer correlation, Eq (23), firmed that the correlation is adequately robust Of all the 986experimental data points, Eq (23) has successfully predicted90% of the data points within±25% agreement with the exper-imental results Overall, the prediction by Eq (23) has a root-mean-square deviation of 18.4% from the experimental data.Figure 15 shows the comparison of the calculated hTPvaluesfrom the general heat transfer correlation, Eq (23), with all 986
con-Figure 15 Comparison of the predictions by Eq (23) with all 986 mental data points for different inclination angles and gas–liquid combinations (see Table 10).
experi-heat transfer engineering vol 31 no 9 2010
Trang 21Table 10 Results of the predictions for 986 experimental heat transfer data points with different gas–liquid combinations and inclination angles by using Eq (23)
deviation data points data points data points deviation
All 986 data points,
0 ◦ ≤ θ ≤ 90 ◦ 18.4 793 (80%) 884 (90%) 922 (94%) −15.3 to 12.5 750 to 127,000 14 to 209,000 9.99 × 10 −3
to 148 × 10 −3 3.64× 10 −3to
26.3 × 10 −3 Air–water (θ = 0 ◦), 160
data points [40]
12.1 154 (84%) 169 (92%) 174 (95%) −7.7 to 11.8 780 to 26,000 600 to 48,000 Air–water (θ = 7 ◦), 187
data points [40]
12.3 164 (88%) 174 (93%) 176 (94%) −10.3 to 9.5 770 to 26,000 560 to 47,000 Air–water (θ = 90 ◦),
105 data points [12]
23.8 79 (75%) 92 (88%) 95 (90%) −24.5 to 11.4 4000 to 127,000 43 to 154,000 Air–silicone (θ = 90 ◦),
56 data points [13]
10.3 37 (66%) 42 (75%) 47 (84%) −1.7 to 9.4 8400 to 21,000 52 to 42,000 Helium–water (θ = 90 ◦),
50 data points [14]
28.3 41 (82%) 42 (84%) 46 (92%) −25.9 to 17.6 4000 to 126,000 14 to 13,000 Freon 12–water (θ =
90 ◦), 44 data points
[14]
29.8 30 (68%) 35 (80%) 36 (82%) −24.9 to 4.0 4200 to 55,000 860 to 209,000
Note Values of constant and exponents: C= 0.55, m = 0.1, n = 0.4, and p = q = r = 0.25.
experimental data points for different inclination angles and
gas–liquid combinations The comparison of the predictions by
Eq (23) with experimental data for air–water horizontal flow is
shown in Figure 16 The results illustrated in Figure 16 show
that the introduction of the flow pattern factor, Eq (16), into the
general heat transfer correlation, Eq (23), provides the needed
capability to handle different flow patterns
Figure 17 shows the comparison of the predictions by Eq
(23) with experimental data for air–water in slightly inclined
pipes (2◦, 5◦, and 7◦) Finally, as illustrated in Figure 18, the
comparison of the predictions by Eq (23) with experimental
data for various gas–liquid combinations in vertical pipes shows
Figure 16 Comparison of the predictions by Eq (23) with experimental data
for air–water horizontal pipe flow (see Table 10).
that the modified inclination factor (I∗)—see Eq (21)—has equately accounted for the inclination effects
ad-PRACTICAL ILLUSTRATIONS OF USING THE GENERAL TWO-PHASE HEAT TRANSFER CORRELATION
The general two-phase heat transfer correlation, Eq (23),
is applicable for estimating heat transfer coefficients for boiling two-phase, two-component (liquid and permanent gas)
non-Figure 17 Comparison of the predictions by Eq (23) with experimental data for air–water in slightly inclined pipes (see Table 10).
heat transfer engineering vol 31 no 9 2010
Trang 22A J GHAJAR AND C C TANG 727
Figure 18 Comparison of the predictions by Eq (23) with experimental data
for various gas–liquid combinations in vertical pipes (see Table 10).
flow in pipes In this section, three illustrations of using the
general two-phase heat transfer correlation, Eq (23), are
dis-cussed The first illustration is about the application of the
corre-lation on air and gas–oil flow in a vertical pipe with gas-to-liquid
volume ratio of approximately two The second illustration is
with air and silicone (Dow Corning 200 Fluid, 5 cs) in a
ver-tical pipe with liquid-to-gas volume ratio of approximately 90
Finally, the third illustration is an application of the correlation
on air and water pipe flow in microgravity condition
Application in Air and Gas–Oil Flow
Dorresteijn [22] conducted an experimental study of heat
transfer in non-boiling two-phase flow of air and gas–oil through
a 70 mm diameter vertical tube The liquid phase consists of
do-mestic grade gas–oil with kinematic viscosity (νL) of 4.7 ×
10−6m2/s and Prandtl number (PrL) of approximately 60 [22]
In the conditions at which VSG = 8 m/s, VSL = 3.16 m/s,
ρG = 2.5 kg/m3, ρL = 835 kg/m3, and α= 0.67, Dorresteijn
[22] measured a value of 1.65 for hTP/hL The following
ex-ample calculation illustrates the use of the general two-phase
heat transfer correlation, Eq (23), to predict the hTP/hL value
measured by Dorresteijn [22]
From the measured superficial gas and liquid velocities, and
void fraction, the gas and liquid velocities are found to be
VG= VSG
α = 11.9 m/s and VL= VSL
1− α= 9.58 m/sThe gas and liquid mass flow rates are calculated as
20 to 30 N/m [66] Using the general two-phase heat transfercorrelation, Eq (23), the value for hTP/hLis estimated to be
Application in Air and Silicone Flow
Liquid silicone such as Dow Corning 200 Fluid, 5 cs, is usedprimarily as an ingredient in cosmetic and personal care productsdue to its high spreadability, low surface tension (σ= 19.7 N/m),nongreasy, soft feel, and subtle skin lubricity characteristics Atwo-phase flow of air and silicone (Dow Corning 200 Fluid,
5 cs) with ˙mL = 0.907 kg/s, x = 2.08 × 10−5, ρ
G = 1.19kg/m3, ρL = 913 kg/m3, µG = 18.4 × 10−6 Pa-s, µ
Trang 23From the gas and liquid mass flow rates, the superficial gas
and liquid velocities can be calculated:
VSG= m˙G
ρGA = 0.149 m/s and VSL= m˙L
ρLA = 9.24 m/s
Using the superficial velocities and void fraction, the gas and
liquid velocities are found to be
VG= VSG
α = 13.5 m/s and VL= VSL
1− α = 9.34 m/sEquations (17) and (16) are then used for calculating the flow
Using Eqs (22) and (21), the inclination factor (I∗) for
verti-cal tube (θ= 90◦) is calculated to be
Finally, with the general two-phase heat transfer correlation,
Eq (23), the value for hTPis estimated to be
hTP= hLFP
1+ 0.55
x
coefficient of 3480 W/(m2-K) by Rezkallah [13] in similar flow
conditions, the general two-phase heat transfer correlation, Eq
(23), overpredicted the measured value by 2%
Application in Microgravity Condition
An air–water slug flow heat transfer coefficient in
micro-gravity condition (less than 1% of earth’s normal micro-gravity) was
measured by Witte et al [67] in a 25.4-mm-diameter horizontal
tube In the conditions at which VSG = 0.3 m/s, VSL = 0.544
al [67] measured a value of 3169 W/(m2-K) for the two-phase
heat transfer coefficient (hTP) The following example
calcula-tion illustrates the use of the general two-phase heat transfer
correlation, Eq (23), to predict the hTPvalue measured by Witte
et al [67]
From the measured superficial gas and liquid velocities, and
void fraction, the gas and liquid velocities are found to be
hTP= hLFP
1+ 0.55
x
SUMMARY
The work documented in this article initiated with the tivation to understand, in both fundamentals and industrial ap-plications, the importance of non-boiling two-phase flow heattransfer in pipes Through the survey of literature and tracing thevalidity and limitations of the numerous two-phase non-boilingheat transfer correlations that have been published over the past
mo-50 years, it was established that there is no single correlationcapable of predicting the two-phase flow heat transfer for allfluid combinations in vertical pipes [11]
The results from the literature survey prompted the ment of a two-phase non-boiling heat transfer correlation that
develop-is robust and applicable to turbulent gas–liquid flow in verticalpipes with different flow patterns and fluid combinations [36].heat transfer engineering vol 31 no 9 2010
Trang 24A J GHAJAR AND C C TANG 729
Since the development of the two-phase non-boiling heat
trans-fer correlation for vertical pipes by Kim et al [36], extensive
efforts have been invested in the development of the general
two-phase heat transfer correlation, Eq (23) When compared with
experimental data from horizontal, slightly inclined, and vertical
pipes with various fluid combinations and flow patterns, the
gen-eral two-phase heat transfer correlation successfully predicted
90% of the data points within±25% agreement with the
ex-perimental data and has a root-mean-square deviation of 18.4%
from the experimental data In addition, practical illustrations of
using the general two-phase heat transfer correlation were also
discussed
In the efforts of investigating non-boiling two-phase flow
heat transfer in pipes, a significant amount of work has also been
done on understanding void fraction A very extensive
compar-ison of 68 void fraction correlations available in the literature
against 2845 experimental data points was conducted by
Wold-esemayat and Ghajar [41] From this work an improved void
fraction correlation, Eq (20), was proposed The improved void
fraction correlation gives noticeable improvements over other
correlations when compared with 2845 experimental data points
of various pipe sizes, inclinations, and two-phase fluid mixtures
from various sources with different experimental facilities
FUTURE PLANS
As pointed out in the introduction, the overall objective of our
research has been to develop a heat transfer correlation that is
robust enough to span all or most of the fluid combinations, flow
patterns, flow regimes, and pipe orientations (vertical, inclined,
and horizontal) As presented in this article, we have made a
lot of progress toward this goal However, we still have a long
way to go In order to accomplish our research objective, we
need to have a much better understanding of the heat transfer
mechanism in each flow pattern and perform systematic heat
transfer measurements to capture the effect of several parameters
that influence the heat transfer results We will complement these
measurements with extensive flow visualizations
We also plan to take systematic isothermal pressure drop
measurements in the same regions where we will obtain or
have obtained heat transfer data We will then use the pressure
drop data through “modified Reynolds analogy” to back out
heat transfer data By comparing the predicted heat transfer
results against our experimental heat transfer results, we would
be able to establish the correct form of the “modified Reynolds
analogy.” Once the correct relationship has been established, it
will be used to obtain two-phase heat transfer data for the regions
where, due to limitations of our experimental setup, we did not
collect heat transfer data The additional task at this stage would
be collection of isothermal pressure drop in these regions
At the present stage, the general two-phase heat transfer
correlation, Eq (23), has been validated with experimental data
for horizontal, slightly inclined, and vertical pipes; however,
its performance for pipe inclination angles between 7◦and 90◦
Table 11 Comparison of capabilities of the current and new experimental setups
Current experimental New
Test section I.D 2.54 cm (1 inch) 1.27 cm (0.5 inch) Heat transfer section with
flow observation sections
Test section orientation 0 ◦to 7◦ 0◦to±90 ◦
has yet to be validated Hence, we have recently constructed arobust experimental setup that is equipped for measuring heattransfer, pressure drop, and void fraction and also conductingflow visualization in air–water flow for all major flow patternsand inclination angles from 0◦ (horizontal) to±90◦(vertical).
A comparison between the capabilities of the current and newexperimental setups is summarized in Table 11
The new experimental setup consists of two test sections.One test section is a stainless-steel pipe and will be used forheat transfer and heated pressure drop measurements The othertest section is a clear polycarbonate pipe and will be used forisothermal pressure drop and void fraction measurements andflow visualization The capabilities of the new experimentalsetup allow an undertaking that combines the study of heat trans-fer, flow patterns, pressure drop, void fraction, and inclinationeffects Such combination of study has not been documented yet.The already-mentioned systematic measurements will allow
us to develop a complete database for the development of our
“general” two-phase heat transfer correlation
NOMENCLATURE
A cross-sectional area, m2
C constant value of the leading coefficient in Eqs (11),
(19), and (23), dimensionless
C0 distribution parameter, dimensionless
c specific heat at constant pressure, kJ/(kg-K)
D pipe inside diameter, m
Eo E¨otv¨os number, dimensionless
FP flow pattern factor, Eq (16), dimensionless
FS shape factor, Eq (17), dimensionless
Gt mass velocity of total flow, ρV, kg/(s-m2)
g gravitational acceleration, m/s2
h heat transfer coefficient, W/(m2-K)
I inclination factor, Eq (18), dimensionlessheat transfer engineering vol 31 no 9 2010
Trang 25I modified inclination factor, Eq (21), dimensionless
K slip ratio, dimensionless
m mass flow rate, kg/s or kg/min
Nu Nusselt number, hD/k, dimensionless
NST number of thermocouple stations, Eq (25),
dimen-sionless
n exponent in Eqs (11), (19), and (23), dimensionless
P mean system pressure, Pa
Pa atmospheric pressure, Pa
P/L total pressure drop per unit length, Pa/m
p exponent on the Prandtl number ratio term in Eqs
(11), (19), and (23), dimensionless
Pr Prandtl number, cµ/k, dimensionless
Q volumetric flow rate, m3/s
q exponent on the viscosity ratio term in Eqs (11),
(19), and (23), dimensionless
˙q heat flux, W/m2
r exponent on the inclination factor in Eqs (19), and
(23), dimensionless
Re Reynolds number, ρVD/µB, dimensionless
ReL liquid in-situ Reynolds number, 4 ˙m/
(π µ L D√
1− α), dimensionless
ReM mixture Reynolds number, dimensionless
ReTP two-phase flow Reynolds number, dimensionless
RL liquid holdup, 1− α, dimensionless
SL wetted perimeter, m
uGM drift velocity for gas, m/s
XTT Martinelli parameter, dimensionless
x flow quality, ˙mG/( ˙mG+ ˙mL), dimensionless
z axial coordinate, Eq (25), m
z element length of each thermocouple station, Eq
(25), m
Greek Symbols
α void fraction, AG/(AG+AL), dimensionless
µ dynamic viscosity, Pa-s
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Afshin J Ghajar is a Regents Professor and Director
of Graduate Studies in the School of Mechanical and Aerospace Engineering at Oklahoma State University Stillwater, and a Honorary Professor of Xi’an Jiao- tong University, Xi’an, China He received his B.S., M.S., and Ph.D all in mechanical engineering from Oklahoma State University His expertise is in ex- perimental and computational heat transfer and fluid mechanics Dr Ghajar has been a summer research fellow at Wright Patterson AFB (Dayton, OH) and Dow Chemical Company (Freeport, TX) He and his co-workers have pub- lished over 150 reviewed research papers He has received several outstanding teaching/service awards, such as the Regents Distinguished Teaching Award; Halliburton Excellent Teaching Award; Mechanical Engineering Outstanding Faculty Award for Excellence in Teaching and Research; Golden Torch Faculty Award for Outstanding Scholarship, Leadership, and Service by the Oklahoma State University/National Mortar Board Honor Society; and recently the Col- lege of Engineering Outstanding Advisor Award Dr Ghajar is a fellow of
the American Society of Mechanical Engineers (ASME), Heat Transfer Series
Editor for Taylor & Francis/CRC Press, and editor-in-chief of Heat Transfer Engineering He is also the co-author of the fourth edition of Cengel and Gha-
jar, Heat and Mass Transfer—Fundamentals and Applications, McGraw-Hill,
2010.
Clement C Tang is a Ph.D candidate in the
School of Mechanical and Aerospace Engineering at Oklahoma State University, Stillwater He received his B.S and M.S degrees in mechanical engineering from Oklahoma State University His areas of spe- cialty are single-phase flow in mini and microtubes and two-phase flow heat transfer.
heat transfer engineering vol 31 no 9 2010
Trang 28CopyrightC Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457630903500858
Total Sites Integrating Renewables
With Extended Heat Transfer and
Recovery
PETAR SABEV VARBANOV and JI ˘ R´I JAROM´IR KLEME ˇS
MC Chair (EXC) “INEMAGLOW”, Centre for Process Integration and Intensification (CPI2), Research Institute of Chemical
Technology and Process Engineering, Faculty of Information Technology, University of Pannonia, Veszpr´em, Hungary
The majority of industrial, residential, service, and business customers, as well as agriculture farms, are still dominated by
fossil fuels as primary energy sources They are mostly equipped with steam and/or gas turbines, steam boilers, and water
heaters (running on electricity or gas) for conversion units The challenge to increase the share of renewables in the primary
energy mix could be met by integrating solar, wind, and biomass as well as some types of waste with the fossil fuels This
work analyzes some of the most common heat transfer applications at total sites comprising users of the types just mentioned.
The energy demands, the local generation capacities, and the efficient integration of renewables into the corresponding total
site CHP (combined heat and power) energy systems, based on efficient heat transfer, are optimized, minimizing heat waste
and carbon footprint, and maximizing economic viability.
INTRODUCTION
Renewable resources are usually available on smaller scale
distributed over a given area Their availability (with the
ex-ception of biomass) is usually well below 100% The resource
availability varies significantly with time and location This is
caused by the changing weather and geographic conditions The
energy demands (heating, cooling, and power) of the considered
sites vary significantly with time of the day and period of the
year The variations of the renewable supplies and the demands
are partly predictable and some are not changing in very
regu-lar time intervals—day and night for soregu-lar energy, for instance
However, the availability of other renewables, such as
wind-generated energy, can be less predictable
For this reason, optimizing the design of energy conversion
systems using renewable resources is more complex than when
using just fossil fuels By combining the supply and demand
streams of the individual users, such systems may serve
indus-trial plants as well as residential customers and the service sector
Financial support from the EC project Marie Curie Chair (EXC)
MEXC-CT-2003–042618 INEMAGLOW is gratefully acknowledged.
Address correspondence to Dr Petar Sabev Varbanov, Centre for Prossess
Integration and Intensification CPI2, Research Institute of Chemical Technology
and Process Engineering, Faculty of Information Technology, University of
Pan-nonia, Egyetem u 10, Veszpr´em, H-8200, Hungary E-mail:
varbanov@cpi.uni-pannon.hu
(hotel complexes, hospitals) They are typically using variousenergy carriers, and the task is to account for the variability onboth the demand and supply sides
The advanced process integration methodology using time
as another problem dimension is a potential solution to dealwith this problem A basic methodology has been developedpreviously for heat integration of batch processes: time sliceand time average composite curves [1, 2] This methodologyhas been recently revisited by Foo et al [3]
A novel approach has been extending this methodology toheat integration of renewables An important step in this direc-tion has been taken by Perry et al [4] They considered theintegration of waste and renewables into local energy sectorsfor a given steady state of supply and demand
ISSUES AND CONCEPTS FOR RESOLVING THEM Classification of Energy Demands
Energy demands vary with the various types of end users aswell as with the time schedules Industrial sites mostly require:
• Heating in a wide range starting from 100◦C up to 400◦C andeven to very high temperatures close to 1000◦C
733
Trang 29Table 1 Energy flow demand variability
Electricity Peak/off-peak Main shift/other sifts,
• Cooling in the range 20◦C to 50◦C, and chilling in the range
0◦C to 10◦C
• Refrigeration to temperatures reaching−100◦C and lower.
A special class of applications is farming and
agricul-tural production There are various examples of using the low
potential waste heat and renewables for supplying greenhouse
demands, e.g., the work by Kondili and Kaldellis [5]
Residential sites (dwellings and their complexes in the case
of district heating) feature demands for:
• Moderate-temperature heating of space and hot water
• Air conditioning
• Direct electricity consumption for lighting, cooking,
refriger-ators, and other household appliances
• Electricity for heat pumps
The energy demands for the service industry and for building
complexes (hotels, hospitals, schools and universities, banks,
entertainment premises, governmental complexes) are generally
similar in structure to residential sites Some specific features
are:
• A part of the heating demand can be at a temperature in the
range 90◦C to 150◦C For example, steam can be used in
hotels for cooking and in hospitals for sterilizing bedding and
other appliances
• The share of air conditioning may be significantly higher
compared with residential homes
• The specific resource consumption per person in the service
industry (as hotels and hospitals) is generally higher than in
residential homes because of the overheads for running
addi-tional services, infrastructures, and facilities such as
restau-rants, bars, and entertainment facilities
Classification of Energy Sources
The sources of energy for the considered users are for the
most part common:
• Fossil fuels Currently dominate the energy markets They
can be used in all three site categories—residential, industrial
sites, and service building complexes
• Solar radiation Can be captured into thermal energy carriers(water, steam, antifreeze, etc.) or used directly to generateelectricity A combination of both is possible as well, but this
is not very much developed so far
• Wind This is used mostly for electricity generation, withfuture potential for generating H2for the hydrogen economy
• Waste biomass and energy crop biomass They can be rectly utilized on-site for larger consumers as industrial sites,building complexes, and farms or in district heating plants
di-• Hydropower This is harnessed for electricity generation.Micro-hydropower technologies are available, but they aremainly suitable for remote locations, which generally implyless energy integration
• Geothermal energy This is harnessed at the locations where
it is available or close by
• Ground heat or cold Heat pumps are considered as renewablesources of energy by most classifications
Demand and Supply Characteristics
Both the supply and the demand for energy vary with timeand location To simplify the initial analysis, a given fixed a set
of locations is assumed
Variability of Demands
The time variations of energy demands have been subject
to research in both industrial and residential contexts Table
1 shows the types of temporal variations in energy demandstypical for the various users
An example is a study investigating the variation of tial energy consumption for heating, electricity and hot water[6] The results show two types of trends: hourly variations dur-ing each day, and seasonal variations during the year For thehourly variations (Figure 1) there are nearly steady periods dur-ing the usual office hours and two consumption peak intervals
residen-in the mornresiden-ing and residen-in the evenresiden-ing The seasonal variations arerelatively smooth, with more substantial space heating demandsfrom October until April
Demand variations are mostly predictable and feature nor uncertainties—mainly in the timing of the consumption.The picture is slightly different for buildings, industrial sitesheat transfer engineering vol 31 no 9 2010
Trang 30mi-P S VARBANOV AND J J KLEME ˇS 735
h , e m i T
Piecewise approximation
Figure 1 Typical residential electricity demands within a 24-hour cycle [6].
and farms A similar situation occurs in the other types of
building complexes—service buildings such as hotels and
hos-pitals, where the demand levels will obviously depend on the
occupancy rate and some less predictable features
Table 2 presents energy flow characteristics that point out
some parts of a site that can have a domineering effect
Nat-urally, large industrial plants are usually dominating the site
However, some sites can be well balanced—for example, a pulp
and paper plant and district heating for a town This type of
qualitative analysis can be very useful in deciding how to
com-bine various processes in a total site Another interesting feature
is that the power-to-heat ratio of the energy demands varies
within wide intervals, where the variation for the industry is the
smallest
Variability of Renewable Resources
For their efficient exploitation, it is necessary to assess
re-newables’ overall availability and variability with time Some
of them are close to the performance of fossil fuels and can
be well stored for continuous energy generation An
exam-ple is biomass, where the supply varies by year seasons and
by bio-waste availability However, sufficient storage could be
made available The availability of other renewable sources
such as wind and solar varies more rapidly—in hours and even
minutes
These types of variation present an integration challenge
where the time horizons of the changes are diverse From the
given examples, for biomass, the time slices needed would last
on the order of months and at smallest—weeks For wind and
solar energy, the time slice durations will obviously be much
shorter This brings a necessity to extend the total site ology [7, 8] to deal with the described variations
method-Integrating the Total Sites Including Renewables
Basic Heat Integration (Pinch Technology)
The heat integration methodology has traditionally dealtwith industrial plants, where the most effort has been made
to achieve stable steady-state production [9] It was rized by Smith [10] and more recently by Kemp [11] Someproduction plants are just run in several-month campaigns—
summa-as in sugar plants in European conditions A special type ofprocessing has been batch operation performed for some spe-cific production—biotechnologies (including breweries), phar-maceuticals, and dyestuff production [1]
The second step has been to integrate more industrial plantsinto total sites [7, 8] The integration into total sites frequentlyincludes more than just industrial plants The extension into inte-gration of residential complexes and service buildings has beenthe subject of recent works [4] The presented results indicate asignificant potential for saving even more energy compared tointegrating industrial processes only
Handling the Variability
Demands are generally imposed to the energy conversionsystems and do not belong in the degrees of freedom Short-term fluctuations are modeled using time-differential equations.For the longer interval variations, piecewise approximations ofthe demands are used
The piecewise representation of the demands can be ded within a total site formulation to model the changes in thedemand over longer periods Typical examples are campaigns
embed-in the sugar embed-industry, and first shift and the other shifts embed-in theindustrial plants, where the second and third shift could featureconsiderably lower energy consumption or might not be covered
at all Approximating winter and summer demands especiallyfor the residential buildings is another example
The maximum availability of renewables is usually limitedand cannot be controlled by the energy conversion system Tointegrate renewables as fixed and not belonging to the degrees
of freedom is obviously not correct What can be used as a
Table 2 Energy flow demand characteristics (all are considerably size dependent)
Indicators House/dwelling Industrial site Service or building complex Farms/agriculture
Electricity (E) Tens of kW One to hundreds MW Hundreds of kW to MW Hundreds of kW to MW
Power (shaft work) One to hundreds MW Hundreds of kW to MW Hundreds of kW to MW
Heating (H) Tens of kW Hundreds of kW to MW Hundreds of kW to MW Hundreds of kW to MW
Cooling Tens of kW One to hundreds MW (in the case
Trang 31Figure 2 A total site integrating renewables [4].
degree of freedom is the fraction of the renewable resources to
be harvested, compared with their overall availability
INTEGRATION APPROACH
To account for the variation of the demands, the renewables
availability and simultaneously maximizing the heat recovery,
it is necessary to apply total site integration and when needed to
consider heat storage possibilities
This is performed by drawing site profiles and site composite
curves for a set of time intervals, referred to as “time slices,”
and maximizing the heat recovery within each time slice This
formulation extends the methodology developed previously for
batch processes [1]
A first step is to identify the possible degrees of freedom
Important degrees of freedom are:
(i) Integration of several unit processes and/or consumers into
total sites Examples are the industrial total sites [8, 12]
District heating systems are also a kind of integrating of
residential customers into larger total sites with regard to
heat Integrating many users with different temporal energy
consumption patterns provides the opportunity for more
efficient utilization of the primary resources as well as for
heat exchange for better recovery
(ii) Selection of the degree of utilization of the available
renew-able resources—solar, wind, biomass (including waste),
geothermal, ground heat pumping
(iii) Storing excess waste heat for utilization in a future timeslice
A total site representation arrangement for users and pliers for locally integrated energy systems (LIES), as shown
sup-in Figure 2, at constant demand and supply rates was recentlyanalyzed and presented by Perry et al [4] Advanced analysis in-cluding the availability of renewables, considering heat storageoptions, is needed
Potential Tools to Be Used for Handling Renewables Supply Variability
Various heat integration tools dealing with the variability ofthe demand side for batch processes have been developed in thepast [1, 2] The demand to cope with changing processes wasreflected by the research of Kotjabasakis and Linnhoff [13] andincluding the time by Wang and Smith [14] However, theseworks have not been considerably revisited from that time.Kemp in 2007 [11] (pp 371–372), summarized the proce-dures for targeting using time intervals for variable heat demandchanging with time, developed by Klemeˇs et al [1] His exampleanalyzes variable heat hot utility during the day and during sum-mer/winter scenarios This has been followed by reschedulingpossibilities analysis
An idea is to extend the methodology to also account forthe variability of the renewable energy supplies as well Thereare some similarities, as there are different heat sources andheat demands inside specific time periods Also, a heat transferheat transfer engineering vol 31 no 9 2010
Trang 32P S VARBANOV AND J J KLEME ˇS 737
exploiting a short-term heat storage could save energy when in
one time period there is a surplus and in the following there is
a deficit This can be used to synchronize the heat sinks and
sources over the site at a specific time period
Potential steps that can be taken are:
• Optimal scheduling of heat exchanges between various
pro-cesses and buildings to maximize the heat recovery
• Energy storage in the form of heat fuel, chemicals, and other
energy carriers to enhance energy recovery between time
in-tervals
• Optimal scheduling of fossil energy supplies to cover the
remaining deficits
Among the very important issues related to optimal system
design is an option of heat storage If this is available at a given
time and required capacity with feasible cost, it can considerably
increase the system efficiency Energy storage is a complicated
and demanding issue, which is still waiting for a major
break-through The heat integration methodology could contribute to
this problem solution by providing targets that are supposed to
be achieved
SUGGESTED APPROACH
Based on the previous analysis, a suggestion of the following
steps could be made:
(i) Building a Heat Integration model for each unit (a plant,
building, farm) using assumed average supply and
de-mand figures
(ii) If there is a potential for Total Site Integration—creating
Total Site Profiles and Site Composite Curves
(iii) Analyze the heat supply (especially involving and
poten-tially maximizing renewables) and demand and specify
Time Intervals (Slices) when they are considerably
dif-ferent from base case model
(iv) Create Time Slice Total Site Profiles and Time Slice Site
Composite Curves Introducing Balanced Time Slice Site
Process C:
Hotel
Process D:
Residential area
Heat Storage System District
Heating
CHP Plant
Figure 3 Demonstration example: topology of the considered system.
Table 3 Streams for process A
(vi) Target for energy storage
(vii) Re-draw the Time Slice Total Site Profiles and the anced Time Slice Site Composite Curves
Bal-(viii) Complete the targeting inside Time Slices and overdrawthe Total Site
DEMONSTRATION CASE STUDY Description
The demonstration case study is based on a previously lished case [4], where four areas are integrated in a Total Site—two industrial plants, a hotel, and a residential area (Figure 3).Each of the areas is referred to as a process Each process fea-tures a number of streams—hot and/or cold
pub-The process streams with their main properties and periods
of activity are given in Tables 3 to 6 The heat flow is assumedpositive for cold streams and negative for hot streams The timeintervals are expressed for a 24 h cycle The residential area has
a number of solar thermal collector cells for generating domestichot water and space heating The utilities available at the totalsite are listed in Table 7
An assumed storage facility uses hot water This storage loses
a part of the heat and not all of the heat stored can be successfullyretrieved Generally, this also involves temperature decrease as
a result of cooling down the material In the current example,
Table 4 Streams for process B
Trang 33Table 5 Streams for process C
Number Stream Supply Target CP, kW/ ◦C Type Q, kW From To
1 Soapy water 85 40 0.444 Hot −20.0 6 17
it is assumed that the storage facility is operated continuously
Therefore, the temperature decrease is neglected and only a
duty loss is considered It amounts for 30% of the heat stored
For every 1 kWh stored heat, only 0.7 kWh can be retrieved
and used The operating temperature of the storage facility is
assumed to be 75◦C
Total Site Targeting With Time Slices
After analyzing the process streams data from Tables 3 to
6, three switching time points over the 24 h horizon have been
identified: 6 h, 17 h, and 20 h These define three time slices
visualized in Figure 4
The solar thermal collectors in the residential area (process
D) have total collection area of 196.5 m2 The corresponding
average heat flows captured by the collectors are assumed as:
• During Slice 1: 112.9 kW
• During Slice 2: 92.1 kW
• During Slice 3: no capture
The activity of the streams within the Site processes has been
analyzed and is summarized in Table 8
The energy targets have been evaluated for the Total Site
described earlier For each Time Slice, the following steps are
performed:
Table 6 Streams for process D
Noumber Stream Supply Target CP, kW/ ◦C Type Q, kW From To
1 Space heating 15 25 8.800 Cold 88.0 0 24
2 Hot water base 15 45 0.833 Cold 25.0 0 24
3 Hot water
daytime
15 45 2.167 Cold 65.0 6 20
Table 7 Site utility specifications
Solar hot water Hot 80 ◦C to 50◦C
District hot water Hot 75 ◦C to 50◦C
(i) Construction of the Total Site Profiles [8]
(ii) Construction of Total Site Composite Curves [8]
(iii) Placement of utilities In this step, the renewables are givenprecedence before fossil-fuel-based utilities (low-pressure[LP] or medium-pressure [MP] steam for the consideredexample)
Steps (i) and (ii) are well known in the literature and industrialpractice, and follow straightforward algorithms The result ofapplying them to Time Slice 1 is shown in Figure 5
Step (iii) needs a proper judgment to ensure feasibility of theutility placement If after site-wide heat recovery there is stillneed for utility heating, the respective amount of utility can bedirectly provided [4] It is possible that the recovered heat fromthe Site Source Profile and the solar heat generated, if directlyplaced, violate the temperature feasibility of the problem Thisoption is shown in Figure 6 There it is attempted to matchthe captured solar heat against a part of the Utility Use SiteComposite Curve If placed directly as a single utility stream, apart of the solar heat plot crosses the Utility Use Site CompositeCurve and lies below it This would be physically infeasible,violating the second law of thermodynamics
The described problem can be resolved by following thesesteps:
Figure 4 Time Slices for the example.
heat transfer engineering vol 31 no 9 2010
Trang 34P S VARBANOV AND J J KLEME ˇS 739
Table 8 Process streams activity during the Time Slices
Process Slice 1: 6–17 h Slice 2: 17–20 h Slice 3: 20–26 h
Process A A2, A1, A5-1,
— Process C Soapy water,
Process D Space heating, hot
water base, hot
water day
Space heating, hot water base, hot water day
Space heating, hot water base
Figure 5 Time Slice 1: Site Composites for interprocess heat recovery.
Figure 6 Time Slice 1: Site targets including the solar—initial placement.
Figure 7 Time Slice 1: Site targets for solar capture and partial storage.
(a) Split of the solar heating stream into two branches.(b) Place one of the branches to serve the process heating.(c) Collect the residual solar heat and any excess district hotwater recovered from the Utility Generation Site CompositeCurve and transfer them into the storage
The results are shown in Figure 7: 209.3 kW recovered trict hot water and 112.9 kW of solar hot water are utilized
dis-by the Site Heat Sinks The extra 66.4 kW of solar hot water(the second branch) is available for the heat storage For thewhole duration of Time Slice 1 this totals to 730.4 kWh of heatadmitted to the storage
For the second Time Slice, the active process streams aresupplemented with the heat available from the storage Afterthe assumed heat deterioration rate 0.7 (i.e., 30% loss), the heatretrieved from storage during Time Slice 2 is 511.3 kWh It isdistributed over the Slice duration and generates a 170.4 kW hotstream running from 75◦C to 50◦C, which is embedded in theSite Source Profile
The resulting targets for Time Slice 2 are illustrated inFigure 8 The total heat recovered from the Site Source Pro-file covers entirely the needs for process heating, represented by
Figure 8 Time Slice 2: Site targets for solar capture and storage.
heat transfer engineering vol 31 no 9 2010
Trang 35Figure 9 Time Slice 3: Site targets for solar capture and storage.
the Site Sink Profile The excess 107.3 kW of district hot water
and the smaller amount of captured solar heat (92.1 kW) can be
sent to the storage
The overall heat supplied to the storage during Time Slice
2 is 501.3 kWh After the deterioration 350.9 kWh remains
available during Time Slice 3 For the 10 h duration 35.1 kW
average flow is retrievable from the storage and embedded into
the Site Source Profile for Time Slice 3
The Total Site Targets for Time Slice 3 are given in
Figure 9 Although no solar heat is captured during this time,
the industrial processes provide some excess heat, which can
potentially lead to storage build-up and exceed the storage
capacity
Analysis of the Targets
The targeting shows, that over a short-term horizon, there is
a trend of heat storage build-up In this particular situation, the
capacity of the heat storage cannot be directly targeted This can
be resolved in several possible ways:
• Trying to find an economically feasible potential use of the
waste heat inside or outside the Total Site
• If this is not possible, purging, i.e., wasting part of the hot
water recovered during Time Slice 3 This would produce a
target of 730.4 kWh for the storage capacity
• Performing a more thorough analysis over a longer time
horizon—e.g., a whole month, season, or year This would
re-veal the time-global needs for heat storage capacity, whereby
the demands may substantially increase after the good days
reflected by the example data This would result from potential
drops in ambient temperature or possibly from an increased
number of guests in the hotel In such situations the build-up
behavior identified in the current analysis may be significantly
reduced or disappear altogether
CONCLUSIONS
The inclusion of renewables with their changing availabilityrequires extensions of the traditional heat integration approach.The problem becomes more complicated and has several moredimensions Revisiting some previously developed Process In-tegration tools and their further development enables solvingthis extended problem The presented contribution has been astep in this direction summarizing the problem and suggestingsome options for its solution A demonstration case study illus-trates the heat saving potential of integrating various users andusing heat storage The advanced tools based on the suggestedmethodology have been under development
NOMENCLATURE
CHP combined heat and power generation
CP capacity flow rate, kW/◦C
MP medium pressure (steam)
LP low pressure (steam)
REFERENCES
[1] Klemeˇs, J., Linnhoff, B., Kotjabasakis, E., Zhelev, T K., mouti, I., Kaliventzeff, B., Heyen, G., Mar´echal, F., Lebon, M.,Puigjaner, L., Espuña, A., Graells, M., Santos, G., Prokopakis, G.J., Ashton, G J., Murphy, N., Paor, de A M., and Kemp, I C.,
Gre-Design and Operation of Energy Efficient Batch Processes, nal Report, Commission of the European Communities Brussels,
[6] Bance, P., Residential-Scale Fuel Cell CHP: A Better Match for
Domestic Loads, Cogeneration & On-Site Power Production,
vol 9, no 3, www.cospp.com/display article/330132/122/CRTIS/none/none/1/Residential-scale-fuel-cell-CHP-a-better-match-for-domestic-loads, retrieved April 7, 2008
heat transfer engineering vol 31 no 9 2010
Trang 36P S VARBANOV AND J J KLEME ˇS 741[7] Dhole, V., R., and Linnhoff, B., Total Site Targets for Fuel, Co-
Generation, Emissions, and Cooling, Computers and Chemical
Engineering, vol 17, supplement, pp S101–S109, 1993.
[8] Klemeˇs, J., Dhole, V., R., Raissi, K., Perry, S., J., and Puigjaner, L.,
Targeting and Design Methodology for Reduction of Fuel, Power
and CO2on Total Sites, Applied Thermal Engineering, vol 7, pp.
993–1003, 1997
[9] Linnhoff, B., and Hindmarsh, E., The Pinch Design Method for
Heat Exchanger Networks, Chemical Engineering Science, vol.
38, pp 745–763, 1988
[10] Smith, R., Chemical Process Design and Integration, John Wiley
and Sons Ltd., Chichester, 2005
[11] Kemp I., Pinch Analysis and Process Integration A User Guide
on Process Integration for Efficient Use of Energy, 2nd ed.,
Butterworth-Heinemann, Elsevier, IChemE, 2007
[12] Varbanov, P., Perry, S., Klemeˇs, J., and Smith, R., Synthesis of
Industrial Utility Systems: Cost-Effective De-Carbonization,
Ap-plied Thermal Engineering, vol 25, no 7, pp 985–1001, 2005.
[13] Kotjabasakis, E., and Linnhoff, B., Sensitivity Tables for the
De-sign of Flexible Processes (1)—How Much Contingency in Heat
Exchanger Networks Is Cost-Effective?, Chemical Engineering
Research and Design, vol 64, pp 197–211, 1986.
[14] Wang, Y P., and Smith, R., Time Pinch Analysis, Chemical
Engi-neering Research and Design—Transactions of IChemE, vol 73,
no A8, pp 905–914, 1995
Petar Sabev Varbanov is an Associate Professor and
Senior Researcher at the Centre for Process tion and Intensification (CPI 2 ), Research Institute of Chemical Technology and Process Engineering, Fac- ulty of Information Technology, University of Pan- nonia, Veszpr´em, Hungary He graduated from the University of Chemical Technology and Metallurgy
Integra-in Sofia, Bulgaria, with an M.Sc Integra-in chemical gineering His professional interests include process modeling and optimization of chemical processes and energy systems He worked several years in the field of energy efficiency, spe-
en-cializing in integration, at the IChE–Bulgarian Academy of Sciences He ceived his Ph.D in optimization and synthesis of process utility systems from University of Manchester Institute of Science and Technology, Manchester, UK For performing research on minimizing and mitigating climate change he was awarded a scholarship from the UK Tyndall Centre Later he was awarded a Marie Curie EIF Fellowship and successfully performed research on optimizing the start-up of distillation columns at The Technische Universit¨at Berlin This was followed by a Marie Curie ERG Fellowship for assisting his integration into the University of Pannonia–Hungary Presently he is a member of the team
re-of the Marie Curie Chair (EXC) “INEMAGLOW.”
Jiˇr´ı Jarom´ır Klemeˇs is a P´olya Professor and EC
Marie Curie Chair Holder (EXC), Head of the tre for Process Integration and Intensification (CPI 2 )
Cen-at The University of Pannonia, Veszpr´em, in gary Previously he worked for nearly 20 years in The Department of Process Integration and the Cen- tre for Process Integration at UMIST and after the merge at The University of Manchester, UK, as a Senior Project Officer and Honorary Reader He has
Hun-an M.Sc in mechHun-anical engineering Hun-and a Ph.D in chemical engineering from Brno Technical University–VUT Brno, Czechoslo- vakia, and an H D.Sc from National Polytechnic University Kharkov, Ukraine.
He has many years of research and industrial experience, including research
in process integration, sustainable technologies, and renewable energy, which has resulted in many successful industrial case studies and applications He has extensive experience managing major European and UK know-how projects and has consulted widely on energy saving and pollution reduction Previously
he ran research in mathematical modeling and neural network applications at the Chemical Engineering Department, University of Edinburgh, Scotland He
is an editor-in-chief of Chemical Engineering Transactions, subject editor of Journal of Cleaner Production, deputy regional editor of Applied Thermal En- gineering, associate editor for Heat Transfer Engineering, and a member of the editorial board for ENERGY—The International Journal; Cleaner Technologies and Environmental Policies; Resources, Conservation and Recycling; and Inte- grated Technologies and Energy Saving In 1998 he founded and has been since
the president of the international conference “Process Integration, cal Modeling and Optimization for Energy Saving and Pollution Reduction— PRES.”
Mathemati-heat transfer engineering vol 31 no 9 2010
Trang 37Copyright Taylor and Francis Group, LLC
ISSN: 0145-7632 print / 1521-0537 online
DOI: 10.1080/01457630903500874
Alternative Design Approach for
Plate and Frame Heat Exchangers
Using Parameter Plots
MART´IN PIC ´ ON-N ´ U ˜ NEZ, GRAHAM THOMAS POLLEY,
and DIONICIO JANTES-JARAMILLO
Department of Chemical Engineering, University of Guanajuato, Guanajuato, Mexico
The simultaneous design and specification of heat exchangers of the plate-and-frame type is analyzed A pictorial
represen-tation of the design space is used to guide the designer toward the selection of the geometry that best meets the heat duty
within the limitations of pressure drop The design space is represented by a bar plot where the number of thermal plates is
plotted for three conditions: (1) for fully meeting the required heat load, (2) for fully absorbing the allowable pressure drop
in the cold stream, and (3) for fully absorbing the allowable pressure drop in the hot stream This type of plot is suitable
for representing the design space, given the discrete nature of the plate geometrical characteristics, such as effective plate
length and plate width Applications of the use of bypasses as a design strategy are also presented.
INTRODUCTION
The specification of requirements for a plate-and-frame heat
exchanger is usually the responsibility of the process engineer
The design of the actual exchanger is the responsibility of a
specialist engineer The specification of the unit involves the
statement of flow rates of hot and cold fluids, their inlet
tem-peratures, and required outlet temperatures It also involves the
specification of the limits on pressure drops encountered by each
of the streams
Unfortunately, the specification of allowable pressure drop is
often undertaken in a subjective manner (sometimes pure guess),
and poor specification leads to over-design Working from the
specification produced by the process engineer, the designer
seeks a geometry (generally consisting of plate type, count and
surface configuration) that provides the required thermal duty
while observing the two pressure drop constraints
Software used by the designer will typically examine
allow-able geometry and produce a list of exchanger configurations
that provide at least the required amount of heat transfer within
the constraints of pressure drop It is common to find that the
specified configurations satisfy a limiting constraint In some
Address correspondence to Dr Mart´ın Pic´on-N´u˜nez, Department of
Chem-ical Engineering, University of Guanajuato, Noria Alta S/N, Guanajuato, Gto.,
Mexico, C.P 36050 E-mail: picon@quijote.ugto.mx
cases all of the pressure drop specified for a stream will be usedand the unit will transfer more than the required quantity ofheat Thus, in terms of heat transfer the recommended unit isover-designed For the designer this is acceptable since the spec-ification is met However, for the user it can result in lifetimecosts that are orders of magnitude greater than the purchase cost
of the exchanger
Thermal over-design is undesirable It results in exchangeroutlet temperatures that are different from those used in the de-sign of the system When such heat exchangers are incorporatedinto systems they cause under-performance of other units in thesystem when the other unit is correctly sized for its specifiedduty When the heat exchanger is a cooler, over-design of theunit places unnecessary load on cooling systems, which leads
to additional operating costs and reduction in cooling systemcapacity (Picon et al [1])
One means of correcting over-performance on a process plant
is the installation and manipulation of a partial bypass aroundthe oversized unit What does not seem to be widely appreci-ated is that such action can also be used to reduce the pressuredrop across a unit Such action can be considered during design
as a means of circumventing the effects of the pressure dropconstraint on exchanger size
In this article, the authors look at how the specification anddesign of plate-and-frame heat exchangers can be undertakenconcurrently
742
Trang 38M PIC ´ON-N ´U ˜NEZ ET AL 743
Figure 1 Geometrical features of a chevron plate.
EXCHANGER GEOMETRY
The sizing of plate-and-frame heat exchangers has been the
subject of a number of publications (Shah and Focke [2], Focke
[3], Kandlikar and Shah [4] Pic´on et al [5], Buonopane et al
[6], Marriott [7], Kumar [8]) In operation, plate and frame
exchangers exhibit a series of phenomena such as flow
maldis-tribution due to pressure drop effects (Bassiouny and Martin [9,
10], Sunden et al [11], Muley and Manglik [12]) and thermal
distortions due to end channels and middle effects (Polley and
Abu-khader [13]) None of these aspects are considered in the
work presented here
Plate-and-frame heat exchanger geometry is characterized
by the following terms (Figure 1): number of plates (N ), plate
length (LP ), plate width (W ), chevron angle (β), plate spacing
(b), and port diameter (dport)
Most of these factors are linked Tooling costs form an
impor-tant factor in the overall cost of manufacturing plate-and-frame
heat exchangers Consequently, plate geometries are generally
restricted to a rather small fixed range Each specific plate length
has associated width, gap and port diameter Each plate size has
a range of available surface type The available range is a
func-tion of plate size In the case of chevron-type plates, the surface
is characterized by the chevron angle
The free flow area in a plate heat exchanger (Af) can be
defined as:
The wavy shape of the plates increases its surface area and it
depends on the depth of its channels The enlargement factor
(ϕ) is defined as the ratio between the actual surface area (Ad)
and the projected area (Ap):
ϕ= A d
where the projected area (Ap) is defined as:
Since the sectional area for the flow of fluid is irregular, the
equivalent diameter (De) can be expressed as:
D e=4 (Free flow area)
The Reynolds number is defined as a function of the mass
velocity rate (Gc) and the equivalent diameter (De), is expressed
as:
Re= Gc D e
The mass velocity rate as a function of the mass flow rate
(m) and the number of channels per stream (Nc) can be defined
as:
The number of channels per stream (Nc) and per pass (Np)
as a function of the number of thermal plates is:
h H + 1
h C + τ
k w + Rf C + Rf H (13)
In Eq (13), the individual heat transfer coefficients (hH and
h C) are calculated using the appropriate correlations
The effective heat transfer area (Ae) can be calculated by multiplying the surface area (Ad) by the number of thermal
plates (NT):
where the total number of plates can be calculated from:
The log mean temperature difference (TML) is computed
from the following expression:
T LM= (Tin − tout) − (Tout − tin)
Trang 39• Pressure drop in ports and fluid collectors.
• Pressure drop due to friction in the channels of the exchanger
• Pressure drop due to changes in height
Empirically, the pressure drop in ports and collectors has
been calculated as a function of the velocity (v) and the number
The pressure drop in channels is the result of friction and the
contraction-expansion of the fluid due to temperature changes
Therefore, this component is expressed as:
P channel=4Np f L p G2C
2De
1ρ
+
1
ρo − 1ρi
N p G2C (19)
The pressure drop due to changes in height is given by:
P elevation = ±ρgLp (20)
In Eq (20), the positive sign is for ascending flow (pressure
drop due to increase in height) and the negative sign is for
descending flow; g is the gravitational constant and Lp is the
plate length
Considering that the height change in a plate exchanger is
relatively small and that for fluid, the momentum change is
usually negligible, the total pressure drop can be simplified as:
For a given configuration and for Reynolds numbers above
100, the viscous effects upon the total pressure drop PT are
no longer important and the effect of the mass flow becomes the
most significant one
PARAMETER PLOT FOR PLATE-AND-FRAME
EXCHANGERS
Pictorial representations can allow design information in a
much more informative manner than lists of numbers In the case
of shell-and-tube heat exchangers, design procedures have been
improved by the introduction of a “parameter plot” developed
by Poddar and Polley [14] With these exchangers, once a set of
preliminary decisions has been made (baffle type, ratio of shell
diameter to baffle spacing, tube size, number of tube passes, tube
pitch, and bundle layout) the design can be characterized using
two principle dimensions: the tube count and the tube length
The design objectives can then be displayed as three separate
curves, one giving the tube length required for a given duty, and
two giving tube lengths associated with the full absorption of
Tubeside pressure drop
Shellside pressure drop
Duty
Maximum tube length
Maximum tubes for 1 Parallel shell
0.21 0.22 0.24 0.25 0.27 0.29 0.32 0.35 0.39 0.43 0.49 0.57 0.68 0.83 1.09 1.47
Tubeside pressure drop
Shellside pressure drop
Duty
Maximum tube length
Maximum tubes for 1 Parallel shell
0.21 0.22 0.24 0.25 0.27 0.29 0.32 0.35 0.39 0.43 0.49 0.57 0.68 0.83 1.09 1.47
Figure 2 Parameter plot for shell and tube exchangers Plot for a 25% baffle cut and two tube passes.
allowable pressure drop This representation is used in ESDU’sEXPRESST[15] computer program, a typical output from which
is shown in Figure 2 In this plot the duty line crosses the standardlength line (6 m) at a point well above the two pressure droplines This means that the designer has a significant amount
of available pressure drop and can therefore modify both thebaffle design (originally 25%) and the number of tube passes(originally 2) specified The effects of reducing exchanger bafflecut to 20% and increasing the number of tube passes to 4 areshown in Figure 3
The number of tubes required to satisfy the required dutyhas fallen from 355 to 300 The shell-side pressure dropnow controls the design The tube count to meet the con-straint is seen to be 315 The designer now has the choice
of accepting a 5% over-design (by specifying 315 rather than
300 tubes) or making a small relaxation (2–3%) on able pressure drop Alternatively, the designer could choose700
Tubeside pressure drop
Shellside pressure drop
Duty
Maximum tube length
Maximum tubes for 1 Parallel shell
0.37 0.39 0.42 0.44 0.47 0.5 0.54 0.7 0.77 0.87 0.96 1.14 1.35 1.87 2.18 2.54
Tubeside pressure drop
Shellside pressure drop
Duty
Maximum tubes for 1 Parallel shell
0.37 0.39 0.42 0.44 0.47 0.5 0.54 0.7 0.77 0.87 0.96 1.14 1.35 1.87 2.18 2.54
Figure 3 Effect of design changes: four tube passes and 20% baffle cut.heat transfer engineering vol 31 no 9 2010
Trang 40M PIC ´ON-N ´U ˜NEZ ET AL 745
Figure 4 Parameter plot formulation for plate-and-frame heat exchangers.
to examine other options, such as opening up tube pitch or
changing baffle type The user interface incorporated into the
EXPRESS program allows the user to rapidly explore design
options
A similar pictorial representation of the design problem can
be developed for plate-and-frame exchangers The principal
de-sign variables are plate size, surface configuration, and plate
count However, as already noted, the available options take the
form of discrete rather than continuous functions In particular,
plate counter is an integer variable
Thus, rather than have a parameter plot in which all variables
are treated as being continuous a “bar diagram” is favored A
typical example is presented in Figure 4 Three columns are
shown for each available plate size (specified length and width)
One column indicates the number of plates required to
sat-isfy the thermal duty A second column indicates the minimum
number of plates required in order to meet the pressure drop
constraint specified for the cold stream The final column shows
the minimum number of plates required to meet the pressure
drop constraint for the hot stream
DERIVATION OF PARAMETER PLOT
The number of channels of given size and geometry required
to accommodate a specified pressure drop can be determined as
Figure 5 Flow diagram for the determination of hydraulic plates.
THERMAL PLATES
Specifications (m,T,Cp,ρ,µ, k) and
MASS VELOCITY: Gc=m/(Nc*b*W)
NUMBER OF TOTAL PLATES
NUMBER OF CHANNELS:
MASS VELOCITY: Gc=m/(Nc*b*W)
NUMBER OF TOTAL PLATES
NUMBER OF CHANNELS:
ERROR
END
Figure 6 Flow diagram for the determination of thermal plates.
1 The pressure drop that is allowed for passage through theexchanger channels is obtained by subtracting the losses as-sociated with flow through the exchanger ports from thespecified pressure drop limit
2 The friction factor relationship for the given surface is tified
iden-3 The allowable mass flux is then determined from allowablepressure drop and friction factor
4 The flow area is determined from the allowable mass flux
5 The number of channels providing this flow area is mined
deter-The number of plates required to accommodate the allowablepressure drop is twice this number of channels The number ofplates required to satisfy the specified thermal duty is obtainedthrough the simultaneous solution of the following equations(Figure 6):
1 Determination of plate count using the heat exchanger sign equation that relates surface area to mean temperaturedifference, heat duty, and overall heat transfer coefficient
de-2 Calculation of mass flux and film heat transfer coefficientfor number of channels (equal to half the number of platesdetermined earlier) for each stream
3 Calculation of overall heat transfer coefficient from ual stream heat transfer coefficients
individ-EXAMINATION OF BYPASS OPTIONS
Where passing all of a stream through an exchanger results
in a unit that needs to be oversized thermally in order to meet
a pressure drop constraint, the use of a bypass can result in
a reduction of exchanger size (thereby reducing capital cost)and in removal of thermal over-design (avoiding operationalproblems and additional operating cost)
The objective would be to determine the bypass fraction thatprovides full use of allowable pressure drop while achieving thespecified thermal duty The thermal duty of the unit is fixed.The required stream temperature is achieved once the streamheat transfer engineering vol 31 no 9 2010