integrated absorption refrigeration and thermoelectric
TRANSCRIPT
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The Pennsylvania State University
The Graduate School
Integrated Absorption Refrigeration and
Thermoelectric Based Cascaded Waste
Heat Recovery
A Dissertation in
Mechanical Engineering
by
Shahzaib B. Abbasi
© 2020 Shahzaib B. Abbasi
Submitted in Partial Fulfillment
of the Requirements
for the Degree of
Doctor of Philosophy
December 2020
ii
The dissertation of Shahzaib B. Abbasi was reviewed and approved by the following:
Alexander S. Rattner Assistant Professor of Mechanical Engineering
Dissertation Adviser
Chair of Committee
James D. Freihaut
Professor of Architectural Engineering
Bed Poudel
Associate Research Professor of Material Science and Engineering Special Committee Member
Stephen Lynch
Associate Professor of Mechanical Engineering
Matthew Rau
Assistant Professor of Mechanical Engineering
Daniel Haworth
Associate Head for Graduate Programs
iii
Abstract
Waste Heat Recovery (WHR) methods that use only a single process to recover heat may
be termed as single-pathway WHR methods. Common examples are technologies based on
Thermoelectric Generators (TEGs), Organic Ranking Cycles (ORC), or absorption-based
Thermally Activated Refrigeration (TAR). Most single-pathway WHR methods can only
effectively harness heat sources in certain temperature ranges. A system in which heat is recovered
as it cascades from a higher temperature to a lower temperature can be termed a cascaded WHR
system. Cascaded WHR methods are of a particular interest for applications like refrigerated
transport vehicles and industrial carburizing furnace operations where multiple outputs are needed,
such as electricity, refrigeration, and process heating.
This dissertation proposes a novel, integrated cascaded WHR system that uses a
temperature matched approach that can increase process efficiency by providing electrical power
and refrigeration through WHR. To investigate the potential of this approach, cycle models,
thermoeconomic studies, and an experimental investigation of a cascaded approach of WHR, is
performed.
Thermoelectric generators (TEGs) are solid state devices that perform at their optimum
when the temperature difference across their junctions is high. The efficiency of a TEG scales with
the hot junction temperature. However, the efficiency of TEGs is low (~5%), hence the TEG
pathway of WHR only harvests a small portion of a high-availability waste heat. Absorption
Refrigeration Systems (ARS), however, can operate efficiently with low temperature heat sources.
In this study, TEGs and Absorption Refrigeration (AR) subsystems are integrated using a coupling
fluid, such that high-grade waste heat cascades through the TEGs, and the low-grade waste heat
rejected by the TEG subsystem is used to operate an absorption refrigeration subsystem. Both
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systems have previously been investigated as single-pathway waste heat recovery methods but
owing their respective temperature ranges in which they operate efficiently (high for TEGs and
mid-to-low for Absorption Refrigeration), they are well-suited for integration.
A thermoeconomic study of the proposed cascaded WHR system is performed, and the
results from that study are compared to ORC based WHR systems. The study concluded that the
payback periods for TEGs and AR based cascaded WHR systems are comparable to those of ORC
based WHR systems.
In order to investigate the challenges of integrated, cascaded WHR and to provide a basis
for the thermoeconomic feasibility analysis, a 1/10th scale experimental facility was built based on
the results from an engineering model created for the refrigerated transport application. This facility
was experimentally tested at different inlet temperatures and flow rates for simulated exhaust and
coupling fluid, to simulate the operating conditions of the vehicle application.
An important part of the facility construction was the design and development of the Heat
Acquisition Unit which extracts waste heat from the exhaust stream and transfers it through the
TEGs to the coupling fluid, and the coupling fluid delivers heat to the absorption subsystem. Using
this experimental approach, different design tradeoffs like electrical power output vs. cooling
delivery and cost vs. WHR efficiency were explored. The experimental results were used for the
validation of subsystem models.
v
Table of Contents
List of Figures........................................................................................................................... vii
List of Tables ............................................................................................................................. ix
Acknowledgements ..................................................................................................................... x
Introduction ........................................................................................................... 1
1.1 Waste Heat Recovery (WHR) applications ................................................................... 2
1.1.1 Single-pathway WHR .......................................................................................... 3
1.1.2 Integration of TEG and ARS based pathways of WHR ......................................... 8
1.1.3 Integrated Cascaded Waste Heat Recovery Systems ........................................... 10
1.2 Heat Acquisition Unit (HAU) .................................................................................... 11
1.3 Summary of prior research into Waste Heat Recovery methods .................................. 13
1.3.1 Research Needs in Cascaded WHR .................................................................... 15
1.4 Goals of the Present research and Dissertation ........................................................... 15
Literature Review ................................................................................................ 20
2.1 Single-pathway, single-grade WHR studies ................................................................ 22
2.2 Cascaded WHR studies .............................................................................................. 25
2.3 Discussion ................................................................................................................. 26
Thermoeconomic Studies ..................................................................................... 27
3.1 Thermoeconomic Analysis of Vehicle and Furnace Application ................................. 28
3.1.1 System Modeling Results ................................................................................... 28
3.1.2 Thermoeconomic Analysis ................................................................................. 30
3.1.3 Conclusions ....................................................................................................... 37
3.2 ARS-TEG vs. ORC-VCC; Thermoeconomic Comparison .......................................... 38
3.2.1 ORC-VCC System Description .......................................................................... 39
3.2.2 Exhaust Temperature ......................................................................................... 40
3.2.3 Closure Parameters for Heat Exchangers ............................................................ 43
3.2.4 Isentropic Efficiency Assumptions and Working Fluids ...................................... 44
3.2.5 Cycle Model for ORC-VCC ............................................................................... 45
3.2.6 Results ............................................................................................................... 49
3.2.7 Economic Analysis ............................................................................................ 52
3.2.8 Discussion ......................................................................................................... 55
Heat Acquisition Unit Design, Development and Experimentation ....................... 57
vi
4.1 HAU Design Considerations ...................................................................................... 58
4.2 HAU Simple Engineering Model ............................................................................... 61
4.2.1 TEG Model ........................................................................................................ 61
4.2.2 Fin Array Resistance Model ............................................................................... 62
4.2.3 Spreading Thermal Resistance ........................................................................... 63
4.2.4 Model for Heat Lost to the Ambient ................................................................... 64
4.2.5 Stage-wise Total Resistance Calculations ........................................................... 65
4.2.6 Heat Transfer Calculations ................................................................................. 67
4.3 HAU Iterative Development and Experimental Procedure .......................................... 69
4.3.1 HAU Experimental Subsystem Description ........................................................ 71
4.3.2 Design improvements to HAU-2 based on HAU-1 performance ......................... 72
System Description and Cycle Modeling .............................................................. 80
5.1 System Description .................................................................................................... 81
5.2 Cycle Model .............................................................................................................. 84
5.2.1 Conservation Laws............................................................................................. 84
5.2.2 Closure Parameters ............................................................................................ 86
5.2.3 Results ............................................................................................................... 88
Experimental Results ........................................................................................... 90
6.1 Experimental Facility ................................................................................................. 91
6.2 Instrumentation .......................................................................................................... 92
6.3 Exergetic Efficiency .................................................................................................. 93
6.4 Thermoelectric Power ................................................................................................ 95
6.5 Comparison of HAU Model Predictions and Experimental Results............................. 96
6.6 Discussion ............................................................................................................... 101
Conclusions and Recommendations for Future Research .................................... 104
7.1 ARS-TEG thermoeconomic studies ......................................................................... 105
7.2 HAU design ............................................................................................................. 106
7.3 HAU Model prediction and experimental results ...................................................... 107
7.4 Recommendations for future research ...................................................................... 107
7.4.1 Advanced HAU design .................................................................................... 108
7.4.2 Experimental Investigations ............................................................................. 108
Appendix: Velocity Measurement Calibration ......................................................................... 109
Bibliography ........................................................................................................................... 111
vii
List of Figures
Figure 1.1 Waste heat recovery categorization according to the mode of WHR ............................ 3
Figure 1.2 Single-pathway WHR using Thermoelectric generators (TEGs) .................................. 4
Figure 1.3 Single-pathway WHR using Absorption Refrigeration System (ARS) ......................... 6
Figure 1.4 ARS and TEG performance at different WHR Source Temperature ............................ 8
Figure 1.5 (a) Individual, single-pathway WHR systems, (b) More compact, integrated cascaded
WHR system that eliminates the need for additional heat acquisition unit and air-coupled HX ... 11
Figure 1.6 An illustration of a typical Heat Acquisition Unit (HAU) used for integrated WHR .. 12
Figure 2.1 Venn-diagram showing different types of WHR Systems .......................................... 21
Figure 3.1 Cascaded WHR cycle diagram ................................................................................. 28
Figure 3.2 NPV versus time for the vehicle application ............................................................. 35
Figure 3.3 NPV versus time for the carburizing furnace application ........................................... 36
Figure 3.4 NPV versus time for the carburizing furnace application ........................................... 37
Figure 3.5 Schematic for ORC-VCC based WHR system .......................................................... 40
Figure 3.6 (a) Cumulative frequency of exhaust temperatures (b) Cumulative frequency of
available exhaust heat to operate a 5kW absorption refrigeration system for different driving
cycles [15] ................................................................................................................................ 41
Figure 3.7 Real-world velocity and SCR temperature data for a heavy-duty diesel vehicle [51] . 42
Figure 3.8 Frequenc distribution of exhasut temperatures for a heavy-duty refrigeated transport
truck [51] .................................................................................................................................. 42
Figure 4.1 Thermal Circuit for Heat Lost model ........................................................................ 65
Figure 4.2 Stage-1 Thermal Resistance Model ........................................................................... 66
Figure 4.3 Stage-2 Thermal Resistance Model ........................................................................... 67
Figure 4.4 (a) Cover-plate sub-assembly (b) Base-plate subassembly ........................................ 69
Figure 4.5 (a) TEGs mounted on the outside the HAU (b) Coupling plate mounted on top of the
TEGs (c) Complete assembly of the HAU-1 with oil-block cover plate ..................................... 70
Figure 4.6 Assembled HAU-1 ................................................................................................... 71
Figure 4.7 Experimental setup for HAU subsystem evaluation .................................................. 72
Figure 4.8 Equivalent Thermal Circuits:(a) No heat-spreader: Largest temperature drop occurs
between the air-side fins and TEG, (b) Copper heat-spreader: Spreading resistance drops and the
largest temperature drop occurs between the two sides of the TEG ............................................ 74
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Figure 4.9 (a) Thermal Oil side view of HAU-2 (b) Air side view of HAU-2 (c) Illustration of
HAU-2 thermal oil flow path .................................................................................................... 76
Figure 4.10 TEG Power Output Comparison between HAU-1 and HAU-2 ................................ 78
Figure 4.11 Oil Outlet Temperature Comparison between HAU-1 and HAU-2 .......................... 79
Figure 5.1 Cycle diagram for the proposed TEG and ARS based Integrated cascaded WHR
system ...................................................................................................................................... 81
Figure 6.1 Experimental Facility: (a) Experimental Facility highlighting experiment controls
(including DC load) and thermal system with the location of HAU highlighted, (b) Picture shows
the HAU-1 model and simulated exhaust (hot air blower) .......................................................... 91
Figure 6.2 Exergetic Efficiency of HAU-2 at different inlet conditions ...................................... 94
Figure 6.3 TEG Power produced by HAU-2 at different inlet conditions (uncertainty of ±0.2%) 95
Figure 6.4 Qoil Model vs Experiment Comparison ..................................................................... 98
Figure 6.5 Qair Model vs Experiment Comparison ..................................................................... 98
Figure 6.6 Toil Model vs Experiment Comparison ...................................................................... 99
Figure 6.7 Tair Model vs Experiment Comparison ...................................................................... 99
Figure 6.8 ΔPexp vs ΔPmodel comparison .................................................................................... 100
Figure A.1 Stage-2 Airflow, immediately at the exit of the HAU, is contracted due to the shape of
the HAU. This contraction leads to an increase in the velocity at the location of temperature and
velocity measurement ............................................................................................................. 109
Figure B.2 Stage-2 Vcorr (Velocity measurement downstream of original measurement) vs Vmsrd
(original measurement) ........................................................................................................... 110
ix
List of Tables
Table 1.1 Key parameters of applications considered for integrated WHR study ......................... 9
Table 1.2 Classification of WHR methods with respect to source type and grade ...................... 14
Table 3.1 The UA values for refrigerated vehicle and industrial carburizing furnace ................. 30
Table 3.2 Comparison of payback periods with previous studies ............................................... 31
Table 3.3 Comparison of payback periods with previous studies ............................................... 38
Table 3.4 Heat Exchanger Closure Parameters for ORC-VCC .................................................. 43
Table 3.5 Working Fluid pairs and isentropic efficiencies for different studies .......................... 44
Table 3.6 Isentropic Efficiencies of ORC-VCC components ..................................................... 45
Table 3.7 Properties at various state points for ORC-VCC ........................................................ 49
Table 3.8 Important properties of different components of ORC-VCC system........................... 50
Table 3.9 Properties at various state points for ARS-TEG ......................................................... 50
Table 3.10 Important properties of different components of ARS-TEG system.......................... 51
Table 3.11 U values for HXs....................................................................................................... 51
Table 3.12 Areas of the ORC-VCC HXs ................................................................................... 52
Table 3.13 Areas of the ARS-TEG HXs.................................................................................... 52
Table 3.14 Component cost relation matrix .............................................................................. 53
Table 3.15 Capital costs of ORC-VCC system .......................................................................... 54
Table 3.16 Capital cost of the ARS-TEG system ....................................................................... 54
Table 4.1 HAU-1 component details ......................................................................................... 70
Table 4.2 Results from HAU-1 experimental runs ..................................................................... 73
Table 4.3 Test Matrix for HAU-2 ............................................................................................. 77
Table 5.1 Heat Exchanger Closure Parameters for ARS-TEG ................................................... 86
Table 5.2 Summary of model results ......................................................................................... 88
Table 5.3 Point-wise results for ARS-TEG cycle ...................................................................... 89
Table 6.1 Control Matrix for Experimental Facility................................................................... 92
Table 6.2 Measurement types and uncertainties for ARS-TEG experimental facility ................. 93
x
Acknowledgements
Firstly, I would like to thank my advisor, Dr. Alexander S. Rattner, who has guided me
throughout my PhD research. He has been patient and extremely kind. I would also like to thank
all my past and present lab mates, especially Yue Cao and Chris Greer, for helping me with my
research.
I would also like to thank the U.S. Department of State, the Fulbright Program, the
International Institute of Education, the United States Educational Foundation in Pakistan and the
Higher Education Commission of Pakistan for their financial support through a Fulbright Grant.
I am really thankful to my committee members: Dr. James Freihaut, Dr. Bed Poudel, Dr.
Stephen Lynch and Dr. Matthew Rau.
Finally, I would like to thank my wonderful wife Britt, my parents and my sisters for their
ever-present, unwavering love and support.
xi
Disclaimer
Any opinions, findings, and conclusions or recommendations expressed in this publication
are those of the author and do not necessarily reflect the views of the U.S. Department of State, the
Fulbright Program, the International Institute of Education, the United States Educational
Foundation in Pakistan and the Higher Education Commission of Pakistan.
1
Introduction
2
1.1 Waste Heat Recovery (WHR) applications
Dependence on fossil fuels for energy and sustained population growth have resulted in
large amounts of greenhouse gas emissions [1]. This has led to a rise in global temperatures and
environmental degradation. Electricity production and transportation account for the major portion
of greenhouse gas emissions in the US, each contributing about 30% of the national total [2].
Industries like iron and steel, textiles, petroleum, and fertilizer production account for almost 40%
of global energy uses, and contribute to about 37% of global greenhouse emissions [3]. Two-thirds
of energy input for electricity generation in the US is lost as waste heat, and 12.5% of this generated
electricity is used for low availability tasks like water, space heating and low-temperature process
heating [4].
Waste Heat Recovery (WHR) technologies can help reduce emissions by increasing energy
efficiency while also reducing operating costs. Waste heat sources are often graded according to
their temperatures. For example, Jouhara et al. [5] surveyed industrial waste heat sources and
classified them as high-grade (>400°C), mid-grade (100°C – 400°C) or low-grade (<100°C) waste
heat .WHR technologies can be classified according to the mode of energy conversion, Figure 1.1
illustrates this classification and the typical technologies associated with these modes of energy
conversion.
According to a report from the US Department of Energy [6], the most frequently employed
use for WHR in US industries is process heating – specifically, the use of heat exchangers like
regenerators, recuperators and preheaters.
At lower temperatures in industrial exhaust waste heat streams, water vapor condensation
can lead to corrosion and, to achieve any significant heat transfer, large heat transfer surfaces are
3
needed. Combined with a source to end use temperature mismatch, low-to-mid grade waste heat is,
typically, not utilized.
WASTE
HEAT
HEATING COOLINGPOWER
GENERATION
• Space Heating
• Recuperators
• Regenerators
• Preheaters
• Heat pumps and
Transformers
• Thermally Activated
Refrigeration (TAR):
Absoption/Adsorption
• Desiccant
dehumidification
• Steam Rankine Cycles
(SRC)
• Organic Rankine
Cycles (ORC)
• Thermoelectric
Generators (TEGs)
Figure 1.1 Waste heat recovery categorization according to the mode of WHR
1.1.1 Single-pathway WHR
WHR systems that recover heat using one specific technology are called single-pathway
WHR systems. An example of such a system is a typical single-stage Thermoelectric Generator
(TEG)-based WHR system. TEGs are solid state devices that operate on the Seebeck effect in which
heat flow along dissimilar conductors (N-type and P-type) effects an electric current. Figure 1.2
shows a TEG based single-pathway waste heat recovery system.
4
TH
TC
Waste Heat in (Qin)
RL
Heat Rejected (Qout)
I
N P N P N P
Figure 1.2 Single-pathway WHR using Thermoelectric generators (TEGs)
The electrical power produced by a TEG based WHR system can be approximated using
Equation 1.1 [7]:
𝑷𝑻𝑬𝑮 = (𝑵𝑻𝑬𝑮𝑺∆𝑻𝑻𝑬𝑮)𝟐/(4𝑵𝑻𝑬𝑮𝑹𝑻𝑬𝑮) 1.1
Here, PTEG is the maximum power delivered by the TEGs to the load, in W, with resistance
RL = NTEG •RTEG in Ohms. NTEG is the number of TEGs modules in series, S is the Seebeck
Coefficient of a TEG module in V/K, ΔTTEG is the temperature difference across the TEGs in K,
5
and RTEG is the electrical resistances of one TEG module in Ohms. The model presented in Equation
1 assumes that S, RTEG and RL are constants and independent of temperature.
As can be seen from Equation 1.1, for a TEG based WHR system, the amount of waste
heat converted to electric power strongly depends on the temperature difference between the hot
and cold side of the TEGs. Therefore, TEG based, single-pathway, WHR systems can typically
only effectively harness high grade waste heat. Such systems are typically designed to reject heat
at the lowest feasible temperature (e.g., to the ambient) to maximize conversion efficiency.
Absorption Refrigeration Systems (ARS) are a type of Thermally Activated Refrigeration
(TAR) system that uses a refrigerant and solvent working fluid pair for operation (e.g., NH3-H2O,
H2O-LiBr, NH3+LiNO3). Figure 1.3 shows a schematic representation of a single effect ARS cycle.
The operation of a typical single-effect absorption refrigeration system is described here
assuming the NH3-LiNO3 working fluids under investigation in the present study. Waste heat is
supplied to the desorber (points 1→2), heating the dilute NH3-LiNO3 salt solution and desorbing
refrigerant vapor (NH3) from the solution along the desorber. The exiting concentrated salt solution
and refrigerant are separated in a separator tank. The concentrated solution enters the recuperative
solution heat exchanger (SHX) and cools as it preheats the dilute solution (Points 3→4).
Downstream of the SHX, the concentrated solution expands to a lower pressure through a valve
(Points 4→5) before entering the absorber at Point 6.
6
Figure 1.3 Single-pathway WHR using Absorption Refrigeration System (ARS)
The refrigerant vapor exiting the separator tank liquefies as it rejects its heat to the ambient
in the condenser (Points 9→10). The liquefied refrigerant is recuperatively subcooled as it flows
through the refrigerant pre-cooler (RPC) between Points 11→12 and it is then expanded through a
valve (Points 12→13). The refrigerant evaporates in the evaporator delivering cooling to the
conditioned space (Points 13→14). The refrigerant continues through the RPC, precooling the
refrigerant liquid between Point 14→15.
The refrigerant mixes with the low-pressure concentrated solution at the inlet of the
absorber (Point 5). The vapor absorbs into the solution in the absorber while rejecting heat to the
ambient (Points 6→7). The dilute solution exits the absorber, is pumped to a higher pressure (Points
7
7→8), and continues through the SHX where it is recuperatively preheated (Points 8→1),
completing the cycle.
The performance of a refrigeration system is typically rated using the Coefficient of
Performance (COP):
𝑪𝑶𝑷 = 𝑸𝒆𝒗𝒂𝒑/𝑸𝒊𝒏 1.2
Where Qevap is the cooling load in W and Qin is the heat input to the system in W.
The COP of ARS is typically lower than vapor compression refrigeration systems (0.7 for
ARS, compared to 3.5 for VCC) [8]. However, compared to conventional vapor compression
refrigeration systems, ARSs do not require a large power consuming compressor. Instead, they use
a pump which consumes much less electrical power (~20W for a pump and ~1kW for a
compressor). This is due to the fact that the refrigerant in an ARS exists as a solution in a salt-
solution pair, whereas the refrigerant in a VCC system is a gas that requires compression.
Typically, >95% of the input energy to ARSs is in the form of desorber heating to thermally
separate the refrigerant form the salt solution. Therefore, ARS systems are frequently used where
a heat source is readily available and low cost and refrigeration is required, but electricity may be
limited or expensive (e.g. off-grid solar absorption refrigeration [9], in refrigerated trucks [10], for
industrial waste heat recovery [11]).
Absorption refrigeration-based single pathway WHR methods have received renewed
interest in recent years as promising technologies for harnessing low-grade waste heat (60°C –
140°C) [12].
8
1.1.2 Integration of TEG and ARS based pathways of WHR
Figure 1.4 shows a comparison of COP of a representative water/lithium bromide based
ARS [13] (~6°C cooling delivery temperature and 25°C heat rejection temperature), and the
efficiency of a representative TEG module [14] (~30°C cold side temperature) vs. waste heat source
temperature.
Figure 1.4 ARS and TEG performance at different WHR Source Temperature
From the figure it can be seen that the representative ARS only operates effectively with
low-to-medium grade waste heat (maximum COP = 0.74 at 75°C). System performance actually
degrades with higher temperature heat input. According to Herold et al. [13] with an increase in
desorber/source temperature for single-effect ARS, the capacity of the ARS increases with
increased source temperature. This increase in capacity results in an increase in heat exchanger
duty in all heat exchangers of the ARS, this increase in duty is accompanied by greater heat transfer
irreversibilities in all of the heat exchangers. Therefore, an increase in desorber/source temperature
for single effect ARS can result in degradation of the COP.
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
0
1
2
3
4
5
6
7
8
0 50 100 150 200 250 300
TA
R C
OP
TE
G E
ffic
ien
cy (
%)
Source Temperature (°C)
Efficiency COP
9
In contrast, the representative TEG system achieves increased efficiency with source
temperature. In the past, ARS and TEGs have been primarily studied for single pathway WHR
applications. In integrating the two systems, if the >90% of input heat rejected from the TEG system
is delivered to drive an absorption system, the TEG rejection temperature may only be increased
slightly. Both processes can operate efficiently through cascading the single heat source. The
present investigation seeks to explore this potential integration of these WHR pathways, using a
coupling fluid that transfer heat between the two sub-systems, which could result in a more
compact, economical design that may offer greater exergetic efficiencies than can be achieved with
either technology individually.
The two applications of interest for the current study are refrigerated transport vehicles and
carburizing furnaces. Table 1.1 lists the key representative of these application from previous WHR
studies:
Carburizing Furnace [14] Refrigerated Transport Vehicle [15]
Source Temperature 280°C 440-490°C
Rejection Temperature 30°C 20-30°C
Waste Heat Available 20kW 33kW
Single Pathway Type TEGs NH3-H2O ARS
Waste Heat Recovered 240W 5kW
Table 1.1 Key parameters of applications considered for integrated WHR study
For these applications an integrated, TEG and ARS based WHR system can deliver both,
cooling and electrical power. In the case of the carburizing furnace application the WHR system
can space cooling and electrical power for facility lighting. For the refrigerated transport truck
application, the WHR system can deliver refrigeration and supplement auxiliary power to the
truck’s battery. Therefore, the aim of current study is to explore the potential of this proposed
10
integrated waste heat recovery system to deliver the required cooling and electrical power for the
applications described above.
1.1.3 Integrated Cascaded Waste Heat Recovery Systems
An WHR system in which waste heat cascades through WHR pathways that harness, both,
high and low-grade waste heat is classified as an integrated cascaded WHR system. The benefits
of a cascaded approach to the integration of two single-pathway WHR methods can result in a more
compact, lightweight and economical overall system.
Figure 1.5 shows a comparison of individual, single-pathway ARS and TEG based WHR
systems and an integrated WHR system. Figure 1.5(A), shows the two WHR sub-systems
implemented as single-pathway recovery devices that could individually provide cooling or
electrical output. However, as can be seen in Figure 5(B), their integration could deliver both
cooling and electrical output. For the integrated system shown in Figure 5(B), heat cascades
through the TEG subsystem (high-grade WHR) to the ARS subsystem (low-mid-grade WHR), this
system can be classified as an integrated, cascaded WHR system. In an exhaust-gas-heated and air-
cooled WHR system, the heat acquisition (heat acquisition unit, HAU) and rejection heat
exchangers will be the largest components. As the proposed integrated configuration will have
comparable heat acquisition and rejection loads to either stand-alone unit, the overall system
volume need not increase significantly.
To integrate the two pathways, a subsystem is needed that extracts waste heat from the
exhaust stream and delivers heat to the TEG modules, and can deliver the heat rejected via the TEG
cold junctions to the ARS. Such a device can be called a Heat Acquisition Unit (HAU). The next
section describes this device.
11
Figure 1.5 (a) Individual, single-pathway WHR systems, (b) More compact, integrated
cascaded WHR system that eliminates the need for additional heat acquisition unit and air-
coupled HX
1.2 Heat Acquisition Unit (HAU)
Figure 1.6 shows an illustration of a HAU, for a TEG and ARS based WHR system. As
can be seen in the figure, an HAU for an integrated, cascaded WHR system, provides an enclosure
for exhaust gases to transfer heat using extended surfaces (finned surfaces), through a solid
interface (heat spreaders and TEGs) and into a coupling fluid using extended surfaces (water blocks
or finned surfaces). The heat spreader and TEG receive the highest grade of waste heat, whereas
12
the TEG cold side rejects heat to the coupling fluid at lower temperature as heat cascades through
the heat spreader and TEG.
Figure 1.6 An illustration of a typical Heat Acquisition Unit (HAU) used for integrated WHR
The design of a HAU greatly impacts the performance of the WHR system. Some of the
key challenges in designing a HAU for a WHR system are:
• Extended surface sizing: To transfer the required heat from an exhaust stream to
the WHR system. Therefore, an HUA must be designed with sufficient heat
exchange surface area on the exhaust stream and coupling fluid side.
• Pressure drop: In close relation to the extended surface sizing is the problem of
pressure drop in a HAU. The principal in this context is to not
13
• Spreading resistance: Thermal spreading resistance occurs at the interface of large
heat transfer surfaces and the TEGs. An effective HAU must minimize these
spreading resistances.
• Optimum number of TEGs: Adding more than an optimum number of TEGs can
result in lower HAU performance as an increase in TEGs results in an overall lower
ΔT.
• Single/Dual stage WHR: TEG are high thermal resistance devices that require large
heat exchange surfaces on either side to enable any significant heat transfer across
them. Dividing an HAU into one stage with TEGs and another without TEGs can
allow more waste heat recovery than with a single stage HAU. Therefore,
depending on the application requirements, an HAU design can either be single
stage TEG design or a dual stage TEG design.
These and other aspects of HAU design are discussed in detail in Chapter 5.
For the current study, the TEG subsystem is integrated into the HAU and as such, it can be
operated independently of the ARS subsystem by passing the coupling fluid through a heat
exchanger where it rejects heat to a recirculating cooler; this also allows for the control of the inlet
temperature of the coupling fluid to the HAU.
1.3 Summary of prior research into Waste Heat Recovery methods
Research into Waste Heat Recovery (WHR) has intensified dramatically in the last decade.
According to the publication data provided by Scopus[16], of the 7,100 journal articles published
on the topic from 1969 to 2019, over 50% of these articles were published from 2015-2019.
14
Waste heat recovery methods can be categorized with respect to the waste heat source
temperatures. US Department of Energy report [6] categorizes waste heat recovery sources
according to temperatures as under:
a. Low: Less than 230°C (e.g. Flue gas [17], Coolants[18])
b. Medium: 230° – 650°C (e.g. Process gases and vapors[19])
c. High: Greater than 650°C (e.g. Molten slag from steel industry[20], vehicle
exhaust[21])
Organic Rankine Cycles are well suited for low-to-medium range source temperature waste
heat recovery owing to the low boiling point of the organic working fluids [10,11]. Other low-to-
mid temperature waste heat recovery methods include thermally activated cooling (TAC) systems
like solid and liquid desiccant cooling, absorption and adsorption chillers [24]. For high
temperature WHR, methods like steam Rankine cycle (SRC), regenerative burners, waste heat
boilers (WHB), thermophotovoltaic (TPV) generators, thermoelectric generators (TEGs) and
thermionic generator can be used [5] (Table 1.2)
WASTE HEAT
GRADE
TEMPERATURE
RANGE SOURCES
SUITABLE WHR
METHODS
High 650°C and above Molten Steel Slag, Vehicles
Exhaust
SRC, TEGs, TPV,
Regenerators, RGB
Medium 230°C – 650°C Process gases and vapors ORC, TAC
Low Below 230°C Flue Gas, Coolants ORC, TAC
Table 1.2 Classification of WHR methods with respect to source type and grade
Of all the waste heat generated by energy consumption in industry, only 32% is recovered
[25]. Owing to its difficulty, low-grade waste heat accounts for nearly zero percent of this recovered
energy [26]. Therefore, low-to-mid grade waste heat recovery methods have received significant
attention in recent years [5], [19], [25], [27]–[29]. For a more detailed review of WHR state of the
art a review paper by Mahmoudi et al. [19] can be consulted.
15
1.3.1 Research Needs in Cascaded WHR
As discussed in the last section, single pathway WHR, has received significant research
attention. However, WHR systems that can harness both high-grade and low-to-medium grade
portions of waste heat sources have received limited attention. Prior studies into cascaded WHR
explored integrated ORC and SRC cycles [30], ORC and Vapor Compression Cycle (VCC) [31],
combined ARS and VCC systems [32], Cascaded Thermoelectric WHR [33] and other combined
power cycles [26,27,28]. These studies highlight benefits for a cascaded approach to low-to-mid
grade WHR, and motivate the current investigation into high-grade to low/medium grade waste
heat.
The current study proposes to experimentally investigate an integrated, cascaded WHR
approach. The waste heat sources, considered for this study, are the exhausts from refrigerated
transport vehicles and carburizing furnaces. In this approach, a Heat Acquisition Unit (HAU)
receives high-grade waste heat which cascades through TEG subsystem, which produces
electricity, and instead of rejecting the low-grade portion to the ambient, the ARS subsystem
recovers this cascaded low-grade portion of the waste heat and delivers cooling.
1.4 Goals of the Present research and Dissertation
In the proposed Ph.D. dissertation study, the following research questions will be
addressed:
1) What is the thermoeconomic feasibility of the proposed TEG and ARS
based integrated, cascaded WHR system?
16
Any proposed WHR system must be economically viable for significant adoption. This
study investigates the thermoeconomic feasibility of the proposed system by the following:
• Developing application-specific thermodynamic models and conducting a
payback period study
• Developing lab-scale model to provide practical insights into hardware cost,
sizing and thermoeconomic optimization
2) How does the proposed ARS-TEG WHR system compare with
conventional WHR systems?
This study also includes a comparative thermoeconomic study of the proposed ARS-TEG
WHR system with an integrated Organic Rankine Cycle (ORC) and Vapor Compression Cycle
(VCC). The comparative study explores this by means of:
• Developing an application specific thermodynamic model for each WHR
system based on real-world vehicle data and reasonable closure parameters
• Conducting a capital cost analysis
3) What are the design considerations for application specific ARS and TEG
based integrated cascaded WHR systems?
For a specific application of WHR, certain system tradeoffs must be considered during its
design phase:
• Technical tradeoffs in system performance: To recover sufficient heat from the
exhaust stream to operate the WHR system, finned surfaces are used.
Introduction of these finned surfaces results in an increase in system
backpressure. This increase in backpressure requires additional work from the
17
source to move the exhaust gases through the HAU. For each specific
application, the increase in exhaust backpressure must be such that it does not
decrease the waste heat source’s system efficiency. For example, Mikulic et
al. [37] state that fuel consumption of a Diesel engine increases by ~0.13% for
every kPa of backpressure. Using this ratio, the allowable backpressure for the
vehicle application can be calculated. Vehicle engines typically reject 33% of
fuel of energy as exhaust, if, for the cascaded system proposed in this study,
50% of this exhaust waste heat is recovered with TEG conversion efficiency
at 5% and COP of ARS is at 0.5, the overall increase in efficiency will be
around 8% i.e. for a diesel application the increase in backpressure for this
representative case should not exceed ~61.5kPa for the system to remain
efficient.
• Electrical Power vs. Cooling: In designing an integrated cascaded TEG and
ARS based WHR system, the competing nature of electrical power and cooling
requirements must be taken into account. As discussed, in in reference to
equation 1, thermoelectric power generation strongly depends on the
temperature difference across the two TEG junctions. For a specific average
exhaust temperature, the temperature difference across TEG junctions can be
increased by lowering the coupling fluid temperature or by increasing the
number of TEGs in the array. However, lowering the coupling fluid inlet
temperature to the HAU also lowers the desorber heat delivery temperature for
the ARS and if this temperature is lower than the temperature range of efficient
operation for the ARS subsystem, this will lead to an inefficient overall WHR.
The other option of increasing the number of TEG modules also has an upper
18
limit in terms of number of modules. This is discussed in detail in Chapter 5,
but, in short, increasing the number of TEGs beyond an upper limit brings the
temperatures of coupling fluid and exhaust streams closer together in HAU
and causes a “thermal short-circuit”, lowering TEG array performance.
4) Does the proposed WHR system have the capacity to meet cooling and axillary
power requirements for the proposed applications?
To ascertain the proposed WHR system’s viability for delivering useful and relevant
streams of electrical power and cooling, application specific thermodynamic models were created.
A lab scale facility has been built to assess these thermodynamic models and to investigate if the
proposed WHR system can recover from a heat source of approximately 2.0 kW at 400°C, electrical
energy of approximately 25W and deliver 0.8 kW of cooling at -15°C. These parameters represent
a 1/10th scale of a vehicle application’s waste heat requirements. This experimental portion of the
study will also provide insights into the practical challenges of system operation that are not be
captured in the modeling study portion. This will serve as a basis for hardware sizing and
thermoeconomic study.
5) What are the tradeoffs associated with the design of proposed system’s
Heat Acquisition Unit (HAU)?
Design of a HAU is of critical importance in assurance of exergetically efficient operation
of a cascaded WHR system. Two different HAU designs were developed for this study and
compared over a range of input waste heat stream flow-rate and temperatures. Their comparison
(Chapter 5) sheds light on the effects of spreading resistance, choice of flow direction, number of
TEGs and other design parameters on the overall exergetic efficiency of the WHR system.
19
6) How does the ARS sub-system’s cooling delivery performance vary at
different operating conditions?
This study also explores the cooling delivery performance of the absorption refrigeration
sub-system by varying different operating conditions like the heat input rate and temperature, and
the cooling delivery temperature.
The following chapters address these questions by first, presenting a review of recent
research in the area of WHR involving thermally activated refrigeration and thermoelectrics based
WHR methodologies, in Chapter 2. Thereafter, Chapter 3 presents details on the two
thermoeconomic studies conducted during the course of this research. Chapter 4 presents a detailed
discussion on design and development of the Heat Acquisition Units (HAUs) developed during this
research. This chapter also includes a performance comparison between the two HAUs developed
during this study (HAU-1 and HAU-2) in light of experimental results. Discussion regarding the
parameters of HAU design that have a controlling effect on its performance is also provided in
Chapter 4. Chapter 4 also includes a detailed engineering model of the HAU. In Chapter 5, cycle
model for the integrated ARS and TEG based WHR system is presented along with a discussion of
the closure parameters considered for the system and cycle model results for the two applications
considered for this study. Chapter 6 describes the experimental facility and presents the results from
the experiments performed on the lab-scale experimental facility. This chapter also presents a
comparison between the HAU model predictions and the experimental measurements. Chapter 7
provides a brief summary and key finding of this dissertation work alongside recommendations for
future research pathways.
20
Literature Review
21
Figure 2.1 shows a Venn diagram of different types of WHR systems. A WHR system can
be single pathway system (one WHR technology) that utilizes a single waste heat grade (Single-
pathway, single-source) or it can utilize multiple grades of waste heat (Single-pathway cascaded).
Examples of single-pathway, single-grade WHR systems are ORC and TAR based WHR systems.
Whereas, cascaded TEGs are an example of cascaded single-pathway WHR systems. Whereas, a
system that employs multiple WHR technologies and grades to harvest waste heat is called an
integrated WHR system. The present research is an ARS-TEG based integrated, cascaded WHR
study.
SINGLE PATHWAY
INTEGRATED
CASCADED
Figure 2.1 Venn-diagram showing different types of WHR Systems
In this chapter a review of prior relevant studies in WHR is presented. As described earlier,
prior WHR research has focused primarily single grade, single-pathway waste heat recovery.
22
However, some more recent theoretical studies have also considered cascaded WHR systems; both
single-pathway (one WHR technology) and integrated (multiple WHR technologies). This chapter
provides a summary of recent research into single-pathway and cascaded WHR studies.
2.1 Single-pathway, single-grade WHR studies
Li et al. [38] performed a theoretical investigation into using Absorption-Compression
Hybrid Refrigeration Cycle (ACHRC) for Coach Air Conditioning using exhaust gas WHR. Their
proposed system comprised of two subcycles; the absorption cycle would be thermally activated
using vehicle waste heat, whereas the vapor compression subcycle would use engine power.
According to the parameters considered in their investigation, at vehicle speeds exceeding 105
km·hr-1 the absorption subcycle could provide all of the required 30 kW of cooling load. At speeds
below 40 km·hr-1 all of the cooling could be provided by the compression subcycle. At speeds in
between 105 and 40 km·hr-1 the two subsystems would work in tandem. This study is an example
of a single-pathway ARS based WHR approach. The size of their system could be reduced by
eliminating the vapor compression cycle either by the use of PCMs or charging a refrigerant
reservoir at higher speeds which can then be utilized at lower speeds. Moreover, the high-grade
waste heat source (480°C) considered in their study may be utilized to produce electricity using
TEGs, which can then be cascaded to the ARS to produce the required cooling, thereby, increasing
exergetic efficiency compared to their proposed system.
Lu et al. [12] performed modeling and analysis of a large temperature span ARS based
WHR system aimed at recovering heat from flue gas during recovery process. In their system, they
employ temperature matching which they refer to as “continuous-temperature-changing generation
process”. This process takes place in the desorber, where the waste heat source and ammonia-water
23
mixture would be in counter-flow arrangement and the dilute mixture would be looped back into
the generator transferring heat to the incoming ammonia-water mixture and in turn, cooling down.
According to their study, this process could lead to a stable 5°C heat transfer temperature difference
between the waste heat source and solution, allowing for a waste heat temperature span of 67.5K.
For their proposed study, with a waste heat source of 10.16 kW at 150°C and 9 kW of cooling at -
15°C, a COP of 0.93 and an exergetic efficiency of 49.7% was predicted. This is an example of a
low-grade, single-pathway WHR method where the source of heat is flue gas with a large
temperature span. In their study, Lu et al. have used a relatively complex desorber, pre-mixer and
internal heat recovery system. For flue gases with higher temperatures (400°C), a TEG and ARS
based WHR approach will yield similar exergetic efficiencies and provide 1 kW of electrical energy
and a similar cooling capacity.
Sonthalia et al. [39] performed a theoretical study where waste heat from a single cylinder
Internal Combustion (IC) engine was recovered using vapor absorption refrigeration (VAR) and
thermoelectric converters. The electric power from the TEG modules was used to drive the VAR
pump. The maximum COP achieved was 0.28 with a 10% engine load at 2000 rpm. For their
proposed system, the TEG hot junction side would be in direct contact with the exhaust stream and
reject heat to a cooling chamber. The exhaust stream would continue to the desorber and supply
heat to the VAR subsystem. Their approach is an example of single-pathway WHR as the heat
rejected by the TEGs will would not be routed to the desorber but, instead, it will be rejected
through a cooling chamber. With the cascaded approach proposed in the present study, this rejected
heat can be harnessed to operate an ARS subsystem.
Meng et al. [40] performed a multiphysics modeling based investigation of TEG based
single-pathway, WHR from an automobile exhaust. Their key findings from their modeling results
are that the flow-arrangement between the hot stream and cold stream does not affect the overall
24
power output, however, using counterflow arrangements results in a more consistent temperature
difference across the number of TEG modules used. They also conclude that for a specific
geometrically constrained system, adding additional TEGs leads to a deterioration in power output.
Liu et al. [41] performed an experimental investigation where, using TEGs, 944 W of
electrical energy was recovered from an automotive engine’ exhaust waste heat of 51 kW at an
average temperature of 312°C and TEG average cold side temperature of 69°C. This cold side
temperature is sufficient to drive thermally activated refrigeration cycles. If such a system were
incorporated here in a cascaded approach with COP = 0.5, an additional ~12 kW of cooling could
be delivered.
Kaibe et al. [14] experimentally investigated WHR from a carburizing furnace where 240
W were recovered from an exhaust steam at of 20 kW at 240°C. The concluded that the TEG power
output could be increased with more effective heat collection as, of the 20 kW of waste heat
available, they were only able to recover 5 kW as TEG heat input. In their experimental study, the
TEGs were mounted on a heat sink board and flue gases from the carburizing unit directly impinged
on the board. On the cold side, a water-block was used to maintain an average temperature of 30°C.
As proposed above, their system could also be integrating with a bottoming thermally activated
refrigeration system that could meet other plant thermal needs. Such an application may result in
slightly reduced electrical conversion efficiency as the TEG cold side temperature would need to
be increased.
He et al. [42] performed a theoretical investigation on the influence of different cooling
methods for TEG based WHR systems. They considered co-flow and counter-flow arrangements
for air and water as coolants. The key findings from their study are that when using water as coolant,
the power output did not differ by much for counter and co-flow arrangements. For air flow,
25
counterflow arrangement resulted in greater power output for the same module areas. They also
observed that water cooling typically results in greater electrical power from the TEGs.
Rattner et al. [43] conducted a theoretical study to develop an analytical model for
optimally sizing TEG module arrays for WHR. Their study concludes that for a high-flux,
commercially available TEG, arranged in a spanwise array of co/counter-flow arrangement,
thermally connected to the hot and cold stream at isothermal temperatures via highly conductive
heat spreaders, there exits an optimum number of TEGs given by the “thermal-fluid figure of
merit”, HT.
2.2 Cascaded WHR studies
Many experimental studies in recent years have considered high to medium grade, single-
pathway WHR [25], [44]–[47]. Summaries of some of these studies are presented in this sub-
section.
Zhang et al. [25] performed an experimental study between a SRC and a cascaded WHR
system that employs SRC and ORC in tandem (S-ORC). They concluded that for exhaust
temperatures in the range of 150-210°C, ORC based WHR performs better than S-ORC, however,
within an exhaust temperature range of 210-350°C, S-ORC WHR systems would perform better
than ORC based WHR. This is due to better heat source temperature matching in the S-ORC based
WHR system. The cascaded WHR system built for their study had a peak thermal efficiency of
9.6%. The TEG and ARS based cascaded WHR system proposed the current study has the potential
to recover twice the amount of waste heat compared to their study.
More recently, theoretical studies of thermoelectric based WHR typically investigate the
effect of temperature matching of different TEG materials with their optimal temperatures of
26
operation. One such study was conducted by Shen et al. [7] where TEGs of three different material
composition were used to harness waste heat from methane burner. The high-temperature TEG
module receives heat directly from the methane burner’s exhaust, subsequently the high
temperature flue gases provide heat to the medium-temperature TEG module. For the low-
temperature TEGs heat is provided by the water coolant loop used to cool the high-temperature and
medium temperature TEG modules. Through this approach, they were able to achieve a WHR
system efficiency of 7% which is greater than single-stage TEG WHR (~5%).
Although cascaded TEG based single-pathway WHR system operate at a greater efficiency
than if only one type of TEG material were used, single-pathway thermoelectric based WHR reject
nearly 90% of the waste heat at a lower temperature, resulting in an overall low WHR efficiency.
2.3 Discussion
As can be seen form the discussion of past relevant studies provided above, single-pathway
ARS and TEG based WHR systems have been identified as promising pathways for WHR.
However, there is a need to demonstrated their potential as an integrated, cascaded WHR system.
Therefore, the current study proposes to explore the potential of an integrated cascaded, TEG and
ARS based WHR system.
27
Thermoeconomic Studies
28
3.1 Thermoeconomic Analysis of Vehicle and Furnace Application
A thermoeconomic study of cascaded WHR using TEGs and Absorption Refrigeration is
presented here for two power and cooling applications: (1) in a refrigerated transport truck and (2)
for an industrial carburizing furnace. For the vehicle application, TEGs augment the electrical
energy supply (battery) of the vehicle and the absorption refrigeration system provide the cooling
required at -15°C. For the carburizing furnace application, TEGs provide electrical power for the
plant and the absorption refrigeration system provides space cooling at 5°C.
Figure 3.1 Cascaded WHR cycle diagram
3.1.1 System Modeling Results
Figure 3.1 illustrates the integrated cascade refrigeration system. A steady-state cycle
model was developed that applies mass, species, and energy balances to each component assuming
representative closure parameters (e.g., heat exchanger approach temperatures, pump efficiency,
Commercial Off-the-Shelf (COTS) TEG performance parameters). The absorption system is
29
modelled assuming use of the NH3-LiNO3 working fluid pair, which avoids the need for a rectifier
and permits a low desorber temperature [10]. System operation is described below with
representative values for the refrigerated transport truck application.
Hot exhaust gas enters the heat acquisition heat exchanger (HX) (Point 1: 500°C and 0.007
kg s-1), where it rejects 25 kWth to the TEGs, reaching an exit temperature of T2 = 300°C. The HX
has TEGs mounted on the outside surface of the exhaust gas channel, which collect and convert a
portion of acquired heat into electricity. The array of TEG modules would produce 1.63 kW of
electric power, with an efficiency of approximately 6.5%. The hot side temperature of the TEGs is
425°C and the cold side temperature is 180°C. The remaining heat is rejected by the cold-side of
the TEGs directly to the absorption system desorber. Dilute ammonia-lithium nitrate solution
(NH3+LiNO3) solution enters the desorber (Pt. 8: 112°C, 1643 kPa, ψ = 62% LiNO3 mass fraction
in solution). NH3 is evaporated from the solution, bringing the concentrated outlet liquid stream to
T3 = 138°C and ψ3 = 69%. The concentrated LiNO3 solution enters the recuperative solution heat
exchanger (SHX) (Pt. 3) and cools to T4 = 55°C as it preheats the dilute solution. The concentrated
solution expands through a valve to reach (T5 = 55°C, P5 = 174 kPa).
The NH3 vapor is the refrigerant, and liquefies as it rejects heat to the ambient in the
condenser (Pt. 9: 140°C, 1643 kPa → Pt. 10: 40°C, 1643 kPa). The liquid refrigerant is subcooled
in the refrigerant pre-cooler (RPC, Pt. 11: 17°C, 1643 kPa), and is expanded through a valve to the
evaporator (Pt. 12: -22°C, 174 kPa). The NH3 evaporates, delivering cooling to the conditioned
space (Pt. 13: -20°C, 174 kPa). The refrigerant vapor continues through the RPC, precooling the
refrigerant liquid (exiting at Pt. 14: 30°C, 174 kPa).
The NH3 vapor (Pt. 14) mixes with the low pressure concentrated LiNO3 solution (Pt. 5) in
the absorber. The vapor absorbs into the solution rejecting heat to the ambient, and exits at Pt. 6
(40°C, 174 kPa, ψ6 = 62%). The liquid mixture is pumped to high pressure (Pt. 7: 40.5°C, 1643
30
kPa) and continues to the SHX where it is recuperatively preheated to Pt. 8 (112°C, 238 psi),
completing the cycle.
For the refrigerated transport truck application, 25 kWth of high temperature input heat is
acquired from the engine exhaust (T= 500 °C). The TEGs convert 1.63 kW to electricity (ηe =
6.5%), and deliver the remaining 23.4 kWth to the absorption system at 180°C. The absorption
system delivers 12.5 kWth of cooling to a refrigerated trailer at -15°C (COP = 0.53). This would
be sufficient to satisfy the cooling loads of typical large refrigerated transport trucks. This proposed
cascaded system can also provide a significant portion of the auxiliary electrical load of refrigerated
trucks (1.35 kWe net, accounting for pump motor power).
A similar analysis was performed to assess the performance of the proposed cascaded
WHR system using high temperature exhaust from an industrial carburizing furnace. In that
application, 50 kWth would be acquired at 500°C. The TEG stage would produce 3.3 kW of electric
power (ηe = 6.5%), which would result in a net power output of 2.9 kWe accounting for the
absorption pump requirement (377 W). The absorption system would deliver 26.6 kWth of cooling
at 5°C (COP = 0.6).
3.1.2 Thermoeconomic Analysis
3.1.2.1 Heat Exchanger Costs
For the two applications, the following table lists the calculated, required UA values:
Desorber Condenser RPC Evaporator Absorber SHX
UA for Vehicle (W K-1) 1248 1528 126 2585 2290 1615
UA for Furnace (W K-1) 9000 3053 172 5320 4145 1635
Table 3.1 The UA values for refrigerated vehicle and industrial carburizing furnace
31
Heat exchanger costs vary widely, depending on the working fluids, construction materials,
operating pressures, production volume, and heat exchanger design. Here, the heat exchanger
costing guides of Hewitt et al. [48] and Caputo et al. [49] were consulted to arrive at an estimated
air-coupled heat exchanger cost (absorber and condenser) of $3.50 W-1 K. The more complex heat
acquisition HX, which includes the TEGs and desorber channels, is estimated to be $4.50 W-1 K
(based on the air-side UA). TEGs are assumed to cost $0.27 W th-1 [50]. The liquid-liquid/working
fluid heat exchangers (SHX, RPC) are assumed to be compact and low cost compared with the
air/gas coupled heat exchangers, and are costed as a group with other system hardware.
3.1.2.2 Vehicle application
The thermoeconomic performance of the proposed cascaded WHR system is determined
by comparison with TEG-only and absorption refrigeration-only WHR systems. For the TEG-only
system, the heat acquisition UA is calculated to be 390 W K-1 ($1,365), accounting for both heat
acquisition and rejection stages. The TEG module cost is $6,750 and electric output is 1.88 kWe.
To calculate the payback period of TEG-only WHR system, the assumption used, along
with their references from published literature are presented in Table 3.2:
Parameter Qty Units Ref.
Duty hours 8 hr [51]
Duty days 330 hr -
Duty hours per year 2640 hr/yr -
Fuel cost 0.8 $/L [52]
Fuel consumption rate 0.4 L/kWh [53]
Table 3.2 Comparison of payback periods with previous studies
Based on the assumption listed in Table 3.2, yearly fuel consumption (1056 L/kW-1 yr-1)
and yearly cost of fuel ($844.8 $/kW-1 yr-1) can be calculated. Since the TEGs produce an average
total of 1.88 kWe, the savings can be calculated by multiplying the yearly cost of fuel to the power
generated by the TEGs ($1588.2 yr-1). The total capital investment (TEG+HX = $8115) can then
32
be divided by the yearly savings to calculate the payback period, which, for the TEG-only case is
5.1 years.
In the case of using only a WHR absorption refrigeration system (12.5 kWth cooling
capacity), the UA for heat acquisition system is 343 W K-1 and the condenser and absorber require
another 3818 W K-1. Assuming that additional $5,000 in costs will be required (compact HXs,
pump, framing), the total capital cost is $22,478. This can be compared with a conventional diesel-
powered refrigeration unit. Consumer costs for such units are ~$45,000, but assuming that the
manufacturing cost is 1/3rd of this, the conventional system capital cost would be $15,000[54]. The
COPs for the refrigeration systems of trucks typically range from 0.5–1.5 [55]. Assuming the COP
of a diesel-powered refrigeration system to be 1.5, the total power required to deliver a 12.5 kWth
cooling load can be calculated (8.3kW). By using the assumption made in Table 3.2, the total cost
of operation per year for the diesel engine powering the refrigeration system can be calculated
($7,011.8). This gives a simple payback period for the absorption WHR system of just 1.1 years.
For the proposed cascaded WHR system, the efficiency of the TEGs would drop from 7.5%
(standalone) to 6.5% because heat rejection to the desorber at 112-140°C would decrease the
temperature difference across the TEGs. The main HX costs would be $18,382 and additional
hardware costs would be $6,000. The TEG module cost is $6,750. The total capital cost is $31,132.
The TEGs produce 1.63 kW of power; therefore, they contribute $1,377.02 yr-1 in fuel savings. The
equivalent conventional system would have a capital cost of $15,000 and a yearly recurring cost of
$7,011.8. The payback period for this WHR system is calculated as 1.9 years.
The payback for the cascaded system is higher than absorption standalone system but it is
about 2.7 times lower than TEGs alone. Thus, for an application where electricity and refrigeration
are required, the cascaded system offers a significant advantage. Taking into account the factors of
33
“learning curve” [56] and “scale effects” [57] it can be surmised that future TEGs will cost
significantly less, thus, improving the attractiveness of this option.
3.1.2.3 Carburizing Furnace Application
The thermoeconomic performance of the proposed cascaded WHR is assessed by
comparison with TEG-only and absorption chiller-only WHR systems. For an input heat of 50 kW,
the TEG-only heat acquisition HX UA would be 1750 W K-1 ($6,125). The TEG module capital
cost would be $13,500. The total system capital cost would be $19,625. At ηe = 7.5%, the TEGs
would produce 3.75 kW. If this electricity were to be purchased from a utility at $0.11 kWh-1 [58],
the simple payback period would be 5.43 years. The reduced payback period compared with the
truck TEG WHR can be achieved because the industrial furnace is assumed to operate continuously
(vs. 2640 hrs yr-1 for the truck).
The absorption chiller-only WHR would require a UA of 300 W K-1 for the heat acquisition
HX and a UA of 8000 W K-1 for the condenser and absorber (combined). The total cost of these
HXs would be $33,741. Other system costs are estimated to be $15,000 (total cost: $47,940). The
chiller capacity would be 28.5 kW (233 MWh cooling per year). Producing the same cooling with
a conventional electric powered chiller (including capital costs), would cost $4,275 yr-1. The
payback period for this WHR system is 6.86 years. The projected payback period is sensitive to
assumed specific conventional chiller costs ($ kWth-1), which scale dramatically with overall
capacity.
For the integrated cascaded WHR system, the efficiency of the TEGs will drop to 6.5%, as
in the vehicle application case. Total HX costs (condenser, absorber and heat acquisition) would
be $38,516 and additional costs are estimated to be $16,000. The total system capital cost is $
68,016. The equivalent total conventional power and chiller system capital cost is $3,997. The
simple payback or payback for this integrated system is 6.7 years. The integrated system approach
34
has a lower payback period compared to the absorption system but it has a higher payback in
comparison to the TEGs alone.
3.1.2.4 Net Present Value Analysis
In addition to the simple payback period, a Net Present Value (NPV) analysis can lend
more insights into the economic competitiveness of the proposed WHR systems (integrated and
standalone) with each other. NPV is calculated using the formula presented in equation 3.1:
𝑁𝑃𝑉 = −𝑪0 + ∑𝐹𝑛
(1 + 𝑘)𝑁
𝑁
𝑖=1
3.1
In equation 3.1, C0 is the total capital investment of the WHR system and k is the interest
rate assumed to be 5% and Fn are the total savings. A non-negative NPV value indicates a profitable
state of operation, therefore, the earlier a WHR technology approaches a non-negative NPV value,
the more economical it is.
NPV analysis can be sensitive to the capital cost assumptions, therefore, a capital cost
uncertainty of 30% was added to the present analysis. It was also assumed that the WHR systems
have a lifespan of 15 years.
From Figure 3.2, it can be seen that the ‘TEG only’ approach takes the longest to achieve
a non-negative NPV for the vehicle application case. Whereas, the ‘ARS only’ approach is the
fastest to reach a non-negative NPV. The integrated approach, although, takes longer than the ‘ARS
only’ approach, however, in the long-term, within the margin of uncertainty in capital cost, it is
possible for the integrated approach to WHR to yield a greater NPV value.
In this analysis, the TEG power output corresponds to the efficiency of COTS TEG.
However, the analysis could yield different results if the efficiency of the TEG were to improve in
the future. To simulate the effects of such an improvement, for the ‘TEG only case’, a 50% increase
35
in TEG efficiency was assumed. Note that a 50% increase in TEG efficiency would result in a 25%
increase in total power output as half of the TEG power is consumed in overcoming TEG internal
resistance.
Figure 3.2 NPV versus time for the vehicle application
As can be seen in the graph shown in Figure 3.3, a 50% increase in the efficiency of the
TEG modules does not significantly impact the time at which the NPV value reaches a non-negative
number. Moreover, the overall NPV performance, over a period of 15 years, for the increased
efficiency ‘TEG only’ system is only slightly better than the present COTS TEG based system.
Therefore,
36
Figure 3.3 NPV versus time for the carburizing furnace application
Figure 3.4 shows the NPV values over time for the carburizing furnace application. Due
to the large capital cost of the ARS system, the ‘ARS only’ and integrated approaches of WHR take
longer to reach a non-negative NPV value. However, over the entire lifespan of the of the WHR
systems, the integrated approach yields a greater NPV value and is, therefore, the most economical
approach to WHR in a carburizing furnace.
37
Figure 3.4 NPV versus time for the carburizing furnace application
3.1.3 Conclusions
Integrated cascaded WHR systems allow use of both high- and low-grade portions of waste
heat. Such technologies could enable greater efficiencies than single pathway WHR. In this study,
cascaded TEG and absorption refrigeration WHR was modeled for two potential applications:
WHR from refrigerated transport truck exhaust and an industrial carburizing furnace exhaust.
Findings suggests that this cascaded WHR approach could offer a payback period of 1.9 years for
the refrigerated truck application, and 6.7 years for the carburizing furnace. These payback periods
are competitive with those of other waste heat recovery technologies reported in the literature
(Table 3.3)
38
Author System Type Net Input Energy (kW) Heat Source Payback
(Years)
Patel et al. [59] Trigeneration ORC 41.35 Waste Heat 6.2
Shu et al. [60],[61] Transcritical ORC 338 Waste Heat 7.8
Fontalvo et al. [62] Dual Pressure ORC 244 Waste Heat 8
Table 3.3 Comparison of payback periods with previous studies
3.2 ARS-TEG vs. ORC-VCC; Thermoeconomic Comparison
ORC based WHR systems have been a topic of great interest in recent years. According
the publication data from Scopus[16], of the nearly 1700 journal and conference articles discussing
WHR using ORC published in between the period of 1973 to 2020, over 90% were published during
the period between 2010 to 2020. A number of recent publications [63]–[66] have studied the
integration potential of ORCs with Vapor Compression Cycles (VCC) as an integrated WHR mode.
In this section, a comparative thermoeconomic analysis for the application case of
refrigerated transport truck [15] between ORC-VCC and the novel ARS-TEG system, will be
presented. The purpose of this section is to investigate if the proposed ARS-TEG system can
provide the required cooling and how its economic outcomes differ from the more conventional
ORC-VCC based WHR system.
First, a description of the ORC-VCC model is presented. Next, a discussion of the exhaust
temperature and flowrate is presented. Then the closure parameters, device isentropic assumptions,
and working fluid selection criterion is discussed, following which the cycle model for the ORC-
VCC system is explained. This discussion is followed by the presentation of the results from the
cycles of the respective WHR systems. Finally, an economic analysis of the cost of the respective
39
systems will be presented along with discussion of the costing guides in relevant published
literature.
3.2.1 ORC-VCC System Description
ORC-VCC based systems can have a number of different configurations. Several recent
studies have employed a varied number of configurations for the ORC-VCC systems, such as,
multiple regenerators [67] in the ORC subsystem or employed a single working fluid with
combined condenser [68]–[70], shaft connected expander and compressor [66], [71], have been
studied.
For the purposes of the present study, a basic ORC cycle is combined with a VCC by means
of an electricity generator. Figure 3.5 shows the schematic for the ORC-VCC based WHR system.
The pressure of the ORC working fluid in a subcooled state at 1 is increased by a pump
located between 1→2. Between 2→3, the fluid’s temperature rises at it receives heat form the
recuperator. The boiler receives heat from the vehicle exhaust which is transferred to the ORC fluid
between 3→4. The fluid leaves the boiler as a superheated vapor at point 4. Between 4→5 the ORC
fluid expands to a lower pressure as the Expander/Turbine produces mechanical power which is
converted to electricity by an electricity generator. Between 5→6 the ORC fluid is cooled down in
the Recuperator. A condenser between 6→1 cools down the ORC fluid to a subcooled state.
40
1
2
3 4
5 G
6
7
8
9
10
Tcool
Ec
ORCVCC
Boiler
Expander
Recuperator
Condenser
Pump
Evaporator
CompressorThrottling
Valve
Condenser
Figure 3.5 Schematic for ORC-VCC based WHR system
All of the portion of the electricity generated by the ORC subsystem is supplied to the
pump of the ORC subsystem and the remainder is delivered to the compressor of the VCC
subsystem.
In the VCC subsystem, a high pressure, subcooled, refrigerant expands through an
isenthalpic throttling valve between 7→8, to lower temperature and pressure. Between 8→9 the
refrigerant receives heat from a cooled space in the Evaporator as it gets superheated at 9. The
refrigerant is compressed form 9→10 to the subsystem high pressure. The refrigerant is subcooled
in a condenser from 10→7 as heat is rejected to the ambient, completing the cycle.
3.2.2 Exhaust Temperature
The vehicle exhaust temperature is a function of the fuel requirements during different
cycles of operation [15] (plain, city traffic or mountains). Within each cycle, the initial flowrate
and temperature are typically (in case of cold-start) not sufficient to operate the WHR system.
41
Moreover, in the case of city traffic operation cycle, the vehicle is mostly in an idling mode,
therefore, the exhaust temperature is not high enough for WHR operation.
(a) (b)
Figure 3.6 (a) Cumulative frequency of exhaust temperatures (b) Cumulative frequency of available exhaust heat to operate a 5kW absorption refrigeration system for different driving
cycles [15]
Figure 3.6 shows two graphs from the work of Koehler et al. [15] on WHR in refrigerated
transport vehicle using absorption refrigeration. Figure 3.6(a) shows a qualitative distribution of
cumulative frequency of exhaust temperature ranges. From this graph it can be seen that the higher
temperature frequency of plain cycle is the highest. The horizontal axis of Figure 3.6(b) shows
normalized generator/desorber heat inputs, where a value of unity represents a generator/desorber
heat input sufficient, to provide cooling of 5 kW. Negative values occur when the temperature is
lower than the desorption temperature of the refrigerant in the absorption system.
Figure 3.7 shows real-world velocity of a heavy-duty diesel vehicle and the selective
catalytic reduction (SCR) device inlet temperature data , with respect to time, from a cold-start
stage [51]. From the graph, clear fluctuations in exhaust temperature can be seen. To model the
ORC-VCC and ARS subsystems, an average temperature may be selected to conduct a comparison
42
Figure 3.7 Real-world velocity and SCR temperature data for a heavy-duty diesel vehicle [51]
Additionally, Figure 3.8 shows a frequency distribution graph for the exhaust temperature
of a heavy-duty refrigerated transport truck (~450hp) [51]. From the data presented in the graph,
an average exhaust temperature of ~250°C is chosen as the heat input temperature for ORC-VCC
and ARS-TEG systems for the purposes of thermoeconomic comparison.
Figure 3.8 Frequenc distribution of exhasut temperatures for a heavy-duty refrigeated transport
truck [51]
43
3.2.3 Closure Parameters for Heat Exchangers
Heat exchanger sizing is significant in establishing basis for thermoeconomic analysis and
comparison of any thermal cycles. For refrigeration systems, like the ARS subsystem discussed in
this study, AHRI [72], [73] guidelines and standards can be used to set key parameters like degree
of subcooling (condensers) and superheating (evaporators or boilers) and closest approach
temperatures (recuperators).
With respect to ORC, closure parameters were chosen by reviewing previous studies of
similar/proportionate source temperatures, sink temperatures, and input waste heat rate and power
generation. Majority of literature reviewed [66], [68], [74] refer to ‘Pinch Point Temperature’
(ΔTpp) for the ORC subsystem (typically 5-10 K). Other published works assumed no superheat
(evaporators/boilers) or subcooling (condensers) [69]. According to Park et al [75], ORC systems
with a power output of 3-50 kW are typically superheated types. To maintain consistency of
comparison between the two WHR systems, and based upon the preceding discussion, AHRI
guidelines were adopted for both ORC-VCC and ARS-TEG systems.
Table 3.4 lists the closure parameters for heat exchangers in the ORC-VCC system:
Subsystem Component Parameter
ORC
Boiler CAT 10 K
DoSH 2 K
Recuperator CAT 10 K
εHX 0.75
Condenser CAT 10 K
DoSC 2 K
VCC
Evaporator CAT 5 K
DoSH 2 K
Condenser CAT 10 K
DoSC 2 K
Table 3.4 Heat Exchanger Closure Parameters for ORC-VCC
44
In the above table CAT, DoSH and DoSC stand for ‘Closest Approach Temperature’,
‘Degree of Superheat’ and ‘Degree of Subcooling”, respectively. Moreover, the heat source
temperature, cooling delivery and the ambient temperature (250°C, -15°C and 30°C respectively)
are assumed to be the same for both the ORC-VCC and ARS-TEG systems.
3.2.4 Isentropic Efficiency Assumptions and Working Fluids
To estimate the performance of an ORC-VCC WHR, representative isentropic efficiencies
are assumed for the compressors, expanders, and pumps. Actual values of isentropic efficiencies of
these components vary with the size of the component and cycle pressure ratios, among other
parameters.
For example, according Park et al’s review [75] of experimental studies of ORC systems,
ORC based WHR systems that output 1-5kW of power, typically have an expander efficiency in
the range of 0.5-0.75. Table 3.5 lists a summary of isentropic efficiencies of different components
along with the waste heat input, net electric power out and cooling capacities for several ORC-
VCC theoretical studies.
Ref Qin
(kW)
Enet
(kW)
Qcool
(kW)
Working Fluid VCC ORC
ORC VCC Compressor Expander Pump
Type ηS Type ηS Type ηS
[76] 50 1.5 5 R245fa R245fa Piston 0.75 Scroll 0.6 - 0.8
[64] 40 0.6 5 R134a R134a - 0.8 - 0.8 - 0.8
[71] 1600 90 200 R245fa Butane - 0.75 - 0.8 - 0.7
[77] 21 0 4.2 R134a Isobutane - 0.8 - 0.8 - 0.8
[78] 10 0 5.3 R245fa R245fa - 0.8 Scroll 0.75 Piston -
Table 3.5 Working Fluid pairs and isentropic efficiencies for different studies
45
For selection of ORC working fluid, a number of published works were reviewed [70],
[77], [79]. Hærvig et al. [79], state that, to ensure sub-critical operation, the critical temperature of
the ORC working fluid should be, at most, less than 30°C of that of the heat source temperature.
For the refrigerated transport application (Tin = ~250°C), the ORC working fluid used for this
analysis is Cyclopentane which has a relatively high critical temperature (~240°C). For the vapor
compression cycle, according to Gopalnarayanan [80], for commercial, medium temperature
refrigeration applications (-10°C – -25°C), R134a is well suited.
Based on the literature reviewed, Table 3.6 lists the isentropic efficiencies and working
fluids used for the ORC-VCC model developed for the current study.
Subsystem Component Isentropic Efficiency Working Fluid
ORC Expander 0.6
Cyclopentane Pump 0.7
VCC Compressor 0.7 R134a
Table 3.6 Isentropic Efficiencies of ORC-VCC components
Additionally, the mechanical to electrical conversion efficiency of electricity generator
coupled with the expander is assumed to be 0.9.
3.2.5 Cycle Model for ORC-VCC
With the representative closure parameters selected in the preceding sections, a steady state
model was developed for the ORV-VCC based WHR system with the aim of modeling system
performance through parameters like thermal efficiency, Net Electric Power generated, COPVCC
and the overall COP. Additionally, the following assumptions were made:
• Energy and Mass are conserved and the system operate at a steady state.
46
• Negligible pressure losses occur in the heat exchangers.
• Friction, heat losses and kinetic energy changes are negligible.
• The turbine, pump, throttling valve and compressor operate adiabatically with respect to
the ambient.
• Boiler heat input from the exhaust is assumed to be 25 kW with exhaust inlet temperature
at 250°C and the cooling load is assumed to be 5kW.
Given the assumptions mentioned above, the mass flowrate of the ORC subsystem can be
calculated by the following equation:
𝑸𝒃 = ��𝑶𝑹𝑪(𝒉4 − 𝒉3) 3.2
Where Qb is the boiler heat input to the boiler form the exhaust and h4 and h3 are the
enthalpies of ORC working fluid at the boiler outlet and inlet, respectively and ṁORC is the mass
flowrate of the ORC subsystem.
Similarly, the VCC mass flow rate can be found by the following equation:
𝑸𝒄𝒐𝒐𝒍 = ��𝑽𝑪𝑪(𝒉9 − 𝒉8) 3.3
Where Qcool is the evaporator heat input from the refrigerated space form and h9 and h8 are
the enthalpies of the VCC working fluid at the boiler outlet and inlet, respectively and ṁVCC is the
mass flowrate of the VCC subsystem.
Turbine work is calculated using the following equations:
𝜼𝒔,𝒕 = (𝒉4 − 𝒉5)/(𝒉4 − 𝒉𝟓𝒔) 3.4
��𝒕 = ��𝑶𝑹𝑪(𝒉4 − 𝒉5) 3.5
47
Where ηs,t is the isentropic efficiency of the Turbine and Ẇt is the Turbine work rate, and
h4 and h5 are the enthalpies at the turbine inlet and outlet, respectively, and h5s is the isentropic
enthalpy of turbine outlet.
To find pump work, the following equations were used:
𝜼𝒔,𝒑 = (𝒉2 − 𝒉𝟏)/(𝒉𝟐 − 𝒉𝟏𝒔) 3.6
��𝒑 = ��𝑶𝑹𝑪(𝒉2 − 𝒉𝟏) 3.7
Where ηs,p is the isentropic efficiency of the pump and Ẇp is the pump work, and h1 and h2
are the enthalpies at the pump inlet and outlet, respectively, and h1s is the isentropic enthalpy of
pump inlet.
The following equation is used to calculate the heat transfer in the ORC subsystem
condenser.
𝑸𝒄𝒐𝒏,𝒑 = ��𝑶𝑹𝑪(𝒉𝟔 − 𝒉𝟏) 3.8
Where Qcon,p is the condenser heat rejection to the ambient and h6 and h1 are the enthalpies
of ORC working fluid at the condenser inlet and outlet, respectively, and ṁORC is the mass flowrate
of the ORC subsystem.
To find the rate of electric power generated by the electric generator, the following equation
is used:
��𝒈 = 𝜼𝒈 ∙ ��𝒕 3.9
Where ηg is the generator efficiency and Ėg is the electric power generated, and Ẇt is the
Turbine work.
In the VCC subsystem, compressor work rate can be calculated using Equations 3.10 –
3.111:
48
𝜼𝒔,𝒄 = (𝒉10 − 𝒉9)/(𝒉10𝒔 − 𝒉9) 3.10
��𝒄 = ��𝑽𝑪𝑪(𝒉10 − 𝒉9) 3.11
Where ηs,c is the isentropic efficiency of the compressor and Ẇc is the compressor work
rate, and h10 and h9 are the enthalpies at the compressor outlet and inlet, respectively, and h10s is
the isentropic enthalpy of compressor outlet.
To calculate the rate of heat rejected by the VCC condenser to the outlet Equation 3.12 is
used:
𝑸𝒄𝒐𝒏,𝒓 = ��𝑽𝑪𝑪(𝒉𝟏𝟎 − 𝒉𝟕) 3.12
Where Qcon,r is the condenser heat rejection to the ambient and h10 and h7 are the enthalpies
of the VCC working fluid at the condenser inlet and outlet, respectively, and ṁVCC is the mass
flowrate of the VCC subsystem.
The thermal efficiency of ORC subsystem can be calculated by using Equation 3.13:
𝜼𝑻 = ��𝒕/𝑸𝒃 3.13
The VCC COP is given by Equation 3.14:
𝑪𝑶𝑷 = 𝑸𝒄𝒐𝒐𝒍/��𝒄 3.14
To calculate the overall COP of the combined system, Equation 3.15 is used:
𝑪𝑶𝑷𝒐𝒗 = 𝑸𝒄𝒐𝒐𝒍/(��𝒄 + 𝑸𝒃) 3.15
Finally, with all the heat rates, and inlet and outlet temperatures calculated for the heat
exchangers of the combined VCC system, UA, for each heat exchanger can be calculated using
Equation 3.16:
49
𝑼𝑨 = 𝑸/𝑳𝑴𝑻𝑫 3.16
Where Q is the rate of heat transfer in a heat exchanger and LMTD is the log-mean
temperature difference of the heat exchanger in question.
3.2.6 Results
From the cycle model described above a code was developed in EES [81]. The results from
the code are presented in this section. The mass flowrate for the ORC and VCC systems were
calculated as 0.036 kg s-1 and 0.043 kg s-1, respectively.
Table 3.7 lists properties of ORC-VCC system at various state points corresponding to
Figure 3.2:
State Point T(°C) h(kJ/kg) P(kPa)
ORC
1 50 1.828 73.96
2 52.4 8.663 3531
3 89.57 84.61 3531
4 220 582.1 3531
5 114.1 486.8 73.96
6 64.21 410.8 73.96
VCC
7 50 123.5 1319
8 -22 123.5 121.7
9 -20 238.8 121.7
10 61.71 310.8 1319
Table 3.7 Properties at various state points for ORC-VCC
Table 3.8 lists the UA (product of overall heat transfer coefficient and Area of a Heat
Exchanger), the rate of work (Ẇ) and heat rate (Q) of different components of the ORC-VCC
system:
Component ṁ [kg s-1
] UA [kW/K] Ẇ [kW] Q [kW]
ORC Boiler
0.036 1.81 - 18.11
Condenser 1.49 - 14.89
50
Turbine - 3.47 -
Pump - 0.17 -
Recuperator 0.14 - 2.76
VCC
Condenser
0.043
0.81 - 8.123
Evaporator 1 - 5
Compressor - 3.12 -
Table 3.8 Important properties of different components of ORC-VCC system
The COP of the VCC subsystem is 1.6, whereas the overall COP of the ORC-VCC system
is 0.27.
Using the same inputs and closure parameters for ARS-TEG, the properties with respect to
the state point corresponding to Figure 3.1 are presented in Table 3.9:
State Point ṁ [kg s-1
] T [°C] P [kPa] X [-] H
[kJ/kg]
3 0.04123 137.7 1643 0.6879 23.8486
4 0.04123 55 1643 0.6879 -0.3678
5 0.04123 55.5 173.8 0.6879 -0.3678
6 0.04581 40 173.8 0.6191 -7.3976
7 0.04581 40.49 1643 0.6191 -7.1341
8 0.04581 112.4 1643 0.6191 14.6607
9 0.004581 139.7 1643 0 156.6
10 0.004581 40 1643 0 19.079
11 0.004581 16.39 1643 0 7.7246
12 0.004581 -22 173.8 0 7.7246
13 0.004581 -20 173.8 0 124
14 0.004581 30 173.8 0 135.3
Table 3.9 Properties at various state points for ARS-TEG
Table 3.10 lists the UA (product of overall heat transfer coefficient and Area of a Heat
Exchanger), the rate of work (Ẇ) and heat rate (Q) of different components of the ARS-TEG
system:
Component UA [kW/K] Ẇ [kW] Q [kW]
Desorber 0.51 - 10.29
Condenser 0.63 - 6.30
51
RPC 0.052 - 0.52
Evaporator 1.065 - 5.33
Absorber 0.94 - 9.44
SHX 0.67 - 9.98
Pump - 0.12 -
Table 3.10 Important properties of different components of ARS-TEG system
Meanwhile, 212 TEGs generate 0.81 kW of electrical power from 11.11 kW of heat input
from the exhaust, at a conversion efficiency of ~7%. Whereas the overall COP of the ARS-TEG
system is 0.51.
In order to calculate the area of each heat exchanger in both, ARS-TEG and ORC-VCC,
systems, an estimate of the overall heat transfer coefficient (U) is required. Based on the type of
heat exchanger (Gas-to-Liquid, Gas-to-Gas, plate, shell-tube etc.) exhaustive calculations can be
performed to obtain these estimates. However, for the purposes of current high-level comparison
between ARS-TEG and ORC-VCC systems, estimates from previous studies [82], [83] are listed
in Table 3.11.
Component U [Wm-2
K-1
]
Absorber 940
Boiler 150
Condenser 940
Desorber 150
Evaporator 940
Recuperator 580
SHX 580
RPC 580
Table 3.11 U values for HXs
Using the estimated values from Table 7.10, the areas of respective heat exchangers can
be calculated for the ARS-TEG and ORC-VCC systems. The area values for ORC-VCC, and ARS-
TEG heat exchangers are listed in Table 3.12. and Table 3.13, respectively.
52
Component UA [kW/K] A [m2]
ORC
Boiler 1.81 12.93
Condenser 1.49 1.58
Recuperator 0.14 0.25
VCC Condenser 0.81 0.86
Evaporator 1 1.06
Table 3.12 Areas of the ORC-VCC HXs
Component UA [kW/K] A [m2]
Desorber 0.51 3.43
Condenser 0.63 0.67
RPC 0.052 0.09
Evaporator 1.065 1.13
Absorber 0.94 1
SHX 0.67 1.14
Table 3.13 Areas of the ARS-TEG HXs
Using the values calculated in this section, costs for different components and devices of
the ARS-TEG and ORC-VCC systems can be calculated for comparison.
3.2.7 Economic Analysis
To object of the current economic analysis for the ARS-TEG and ORC-VCC systems is to
compare the capital investment required to achieve a cooling of 5kW for the refrigerated transport
truck application using waste heat.
The total cost of the systems is given by the Equation 3.17:
𝒁𝑻𝒐𝒕𝒂𝒍 = 𝒁𝑪 + 𝒁𝑳 3.17
53
Where ZTotal is the total cost in dollars and ZC and ZL are the component and labor costs in
dollars, respectively.
To estimate the costs, correlations from previously published studies were referenced [83]–
[85]. These studies rely upon data from CEPCI [86], with the latest data available from the year
2011. The relevant costing correlations are listed in Table 3.14.
Component Governing Variable Cost Year Cost Index
Compressor V [m3/s] (Volumetric
flowrate) 258.75 + 195.5�� 2011 585.7
Heat Exchangers AHX [m2] (Area of HX) 130(𝐴𝐻𝑋 0.093⁄ )0.78 2000 394.1
Pump Ẇ (Electric Power
input) 1035(�� 300⁄ )
0.25 2001 402.0
Turbine V [m3/s] (Volumetric
flowrate) 1.73(225 + 170��) 2001 394.3
Labor TCC (Total Capital
Cost) 0.3TCC - -
Table 3.14 Component cost relation matrix
To correct the cost for 2019 (latest available yearly average CEPCI cost index), Equation
3.18 is used:
𝒁𝟐𝟎𝟏𝟗 = 𝒁𝑩𝒂𝒔𝒆𝑪𝑰𝟐𝟎𝟏𝟗/𝑪𝑰𝑩𝒂𝒔𝒆 3.18
Where, Z2019 and ZBase are the costs for 2019 and year of reference, respectively, and CI2019
and CIBase are the cost indexes for the year 2019 and year of reference. The value of CI2019 is 607.5
[87].
For the present analysis, heat exchanger costs are calculated using the simplified relation
presented in Table 3.13. However, it should be noted that heat exchanger costs can vary from this
base cost based on the type of heat exchanger (plate, shell-tube, Finned-tube, etc.), material used
and the operational pressure.
54
Based on these data and correlations presented, total cost for each system can be calculated.
Table 3.15 lists the component costs for the ORC-VCC system:
Component V [m3/s] A [m
2] Ẇ [kW] Z2019 [$]
ORC
Boiler -
12.93 - 9,408.3
Condenser 1.58 - 1,825.66
Turbine 0.022 - 3.47 610
Pump - 0.17 241.3
Recuperator 0.25 - 433.4
VCC
Condenser -
0.86 - 1,136
Evaporator 1.06 - 1,337.2
Compressor 0.00071 - 3.12 398.8
Total: $ 15,390.7
Table 3.15 Capital costs of ORC-VCC system
The total labor costs can be calculated by using the relation provided in Table 3.13. The
total labor cost for the ORC-VCC system is calculated to be $4,617.2. The total capital cost of the
ORC-VCC is, therefore, 20,007.9.
Table 3.16 lists the component cost of the ARS-TEG based WHR system:
Component A [m2] Ẇ [kW] Z2019 [$]
Desorber 3.43 - 5012.85
Condenser 0.67 - 935
RPC 0.09 - 195.3
Evaporator 1.13 - 1,405.6
Absorber 1 - 1,277.8
SHX 1.14 - 1,415.3
Pump - 0.12 221.2
Total: $ 10,463.05
Table 3.16 Capital cost of the ARS-TEG system
Based on the component cost, the total labor cost can also be calculated as $3,138.9. The
TEG cost must also be included in the total calculated, according to the latest available TEG cost
55
rate [88], the cost of TEGs is calculated as $10,269.Therefore, the total capital cost of ARS-TEG
WHR system is $23,870.1.
3.2.8 Discussion
From the economic analysis conducted, it can be deduced that the cost of ARS-TEG WHR
system is only ~19% higher than the cost of ORC-VCC based WHR system. Moreover, the ARS-
TEG based WHR system delivers 0.81 kW of electrical power in addition to delivering 5 kW of
cooling.
The cost of ARS-TEG system can be significantly lowered by reducing the number of
TEGs. For example, by lowering the number of TEGs from 212 to 106, the total cost of the ARS-
TEG based WHR system can be reduced by $5,134.5 to be 18,736.5, with the TEG power being
reduced from 0.81 kW to 0.41 kW. This arrangement of ARS-TEG WHR will require a desorber
redesign, where half of the desorber heat from the exhaust stream will be delivered to the working
fluid through the TEG cold side, while the remaining heat would be recovered downstream of the
TEG stage.
In practice both WHR systems must be modified to meet the transient temperature
challenges that were alluded to in section 3.2.2 of this chapter. Specifically, the challenges of
delivering the required cooling when exhaust temperatures are either higher or lower than the
design temperature.
For the case of lower than required exhaust temperatures, electric heaters can be employed
to deliver heat to the desorber for a steady operation. In the case of higher than desired temperatures
of the exhaust gases, a bypass may be employed to divert the exhaust gases form the desorber.
56
Considering the requirement of an electric heater for steady operation of a refrigerated
transport vehicle’s WHR system, ARS-TEG based WHR system is preferable over ORC-VCC
based WHR system, as it can off-set the power required by the electric heater.
57
Heat Acquisition Unit Design, Development and
Experimentation
58
The HAU for the proposed integrated, cascaded TEG-ARS WHR system, provides an
enclosure for exhaust gases to transfer heat through the TEGs into the heat transfer fluid. Extended
heat transfer surfaces (fins) and heat spreaders are needed at the different interfaces for effective
operation. Design considerations for such an HAU involve flow arrangement of hot and cold stream
fluids (e.g., co- vs. counter-flow) and location and alignment of TEG modules with respect to the
exhaust stream flow direction. Extended surface sizing and the associated pressure drops, high
thermal resistance of TEGs, and spreading resistances are some of the technical challenges in
designing a HAU
This chapter will discuss these considerations and also provide details about the iterative
design and development of the HAU. That discussion will be accompanied by a comparative
analysis, based on experimental results, for the two HAU designs developed during this study.
4.1 HAU Design Considerations
Several different coupling-fluid flow arrangements have been proposed in previous studies
for optimal performance of fluid-heated TEG systems. Bejan et al. [89], using a first law analysis,
derived a power density comparison for co-flow and counterflow heat transfer to argue for a co-
flow arrangement as optimum for TEG systems. However, Meng et al. [40] and Bell et al. [90]
predicted better overall performance with counter-flow arrangements in their studies. He et al. [42]
found that flow direction did not significantly impact TEG system performance in the case of gas
heated and liquid cooled TEG systems. However, for gas cooled TEG systems, they found that
counterflow arrangements produced greater power for the same module areas.
The power generation of a thermoelectric module array can be estimated as[91]:
59
𝑷𝑻𝑬𝑮 = (𝑵𝑻𝑬𝑮𝑺∆𝑻𝑻𝑬𝑮)𝟐/(4𝑵𝑻𝑬𝑮𝑹𝑻𝑬𝑮) (4.1)
Equation 4.1 indicates that the power produced by the TEGs is proportional to the square
of the temperature difference between its hot and cold sides. Considering this, for maximum power
generation, the TEG array in an HAU should be positioned where the hot and cold stream have the
greatest temperature difference. The quadratic dependence favors co-flow arrangements which
have a greater temperature difference near the inlet over counter-flow arrangements with more
uniform, but lower peak temperature differences. The local temperature difference can be
maximized at the inlet of a co-flow arrangement of heating and cooling fluids.
Another important consideration for an efficient HAU design is in sizing of the extended
surfaces used in transferring heat from the hot gas stream into the cold liquid stream. Considering
the WHR application of these systems, increase in pressure drops associated with increased size of
extended surfaces, especially on the exhaust/gas side, must be taken into account. For a pin-fin
array type extended surfaces, as illustrated in Figure 1.6 in Chapter 1, the pressure drop can be
predicted with the following correlation [92]:
∆𝒑 = 𝑵𝑳 ∙ 𝝌 ∙ (𝝆 ∙ 𝑽𝒎𝒂𝒙𝟐 𝟐⁄ ) ∙ 𝒇 4.2
Where Δp is the pressure drop in Pa, NL is the number of lateral pin fins, χ is an empirical
correction factor, ρ is density in kg·m-1
, Vmax is the maximum velocity in m·s-1 and f is the friction
factor. For the air-side fins used in this study, f and χ have values of 0.5 and 1 respectively.
Quoilin et al. [84], using a modified form of Equation 4.2, conclude that increasing the
width of a heat exchanger drops 𝑉𝑚𝑎𝑥, consequently lowering the pressure drop, but it also increases
the required area for heat transfer. In light of this observation combined with the observation from
equation 1, a span-wise arrangement of TEGs, close to the inlet of the HAU would result in a more
efficient WHR.
60
On the cold stream/thermal oil side, the correct sizing of the extended surface/oil-blocks is
important for efficient TEG performance, since the cold side of the TEGs needs to be maintained
at lower temperature to ensure a large temperature difference across the TEGs.
Over-sizing the TEG module array will result in reduced average hot-to-cold side
temperature difference for a given heat transfer rate due to increased overall thermal conductance.
However, under-sizing the array will limit total heat transfer, and reduce available thermal power
for conversion. Rattner [43], using an analytical model, proposed a thermal-fluid figure of merit,
HT, which can be used to identify the optimal number of TEGs for a WHR system that maximizes
delivered electric power by balancing these factors:
𝑯𝑻 = [𝑲 ∙ (𝟏 𝜺𝑯 ∙ ��𝑯⁄ + 𝟏 𝜺𝑪 ∙ ��𝑪⁄ )]−𝟏
4.3
HT is the thermal-fluid figure of merit, K is the TEG module conductance in WK-1, εH and
εC are the thermal effectivenesses of hot and cold heat exchangers, respectively, and ĊH and ĊC are
hot and cold stream thermal capacities in WK-1, respectively.
According to Rattner [43], for a span-wise, single row TEG array, with hot and cold sides
at isothermal temperatures, and a specified overall UA, the optimal number of discrete TEG
modules is equal to HT.
The HAU design must also account for the thermal spreading resistance owing to the
relatively smaller contact surface area of the TEGs and thermal oil side fin blocks compared with
the large heat exchange surfaces of the gas side. Lee et al. [93] provide simple spreading resistance
correlations for uniform heat-flux source, according to which the spreading resistance, for relatively
thick plates, depends entirely on the contact size of the heat source.
To operate at their highest efficiency, TEGs require a high temperature across their hot and
cold sides. If the HAU were divided into two stages, in the first stage high availability waste heat
61
could be harnessed by the TEGs. While, in the second stage, the absence of high thermal resistance
of TEGs allows for low availability waste heat to be harnessed by the thermal oil stream. This
concept will be further discussed in the next section.
4.2 HAU Simple Engineering Model
This section will discuss the development of a simple engineering model for the HAU-2.
The model solves for the intermediate temperatures of the gas and liquid streams by stage-wise
modeling of the thermal resistances of different components of the HAU. The computation of the
engineering model was conducted using Engineering Equation Solver (EES)[81].
4.2.1 TEG Model
Thermoelectric power delivery can be estimated using the following equation:
𝑷𝑻𝑬𝑮 = (𝑵𝑻𝑬𝑮𝑺∆𝑻𝑻𝑬𝑮)𝟐/(4𝑵𝑻𝑬𝑮𝑹𝑻𝑬𝑮) 4.4
In the case of experiments conducted in this study, ΔTTEG is not known, however, a direct
measurement of the power provided to an external D.C load is recorded for each experimental run,
using which the total power generated by the TEG module array can be calculated. With PTEG
known, ΔTTEG can be inferred from Equation 4.4.
Once the temperature difference across the TEGs is known, the heat across the TEG
module array, QTEG, can be found using Equation 4.5:
𝑸𝑻𝑬𝑮 = ∆𝑻𝑻𝑬𝑮/(𝑵𝑻𝑬𝑮 ∙ 𝑹𝑻𝑬𝑮,𝒄𝒐𝒏𝒅) 4.5
Where, RTEG is the TEG conductive Resistance in K·W-1.
62
4.2.2 Fin Array Resistance Model
A model was developed to predict the thermal resistances of a fin block arrays, with respect
to geometry, on both air and thermal oil sides. Inlet flowrates and inlet temperatures can be set to
experimental values. The thermophysical properties of the fluid streams were calculated for their
average temperatures in stage 1, using built-in functions in EES.
The Nusselt number correlation used for this analysis is provided in Incropera et al. [92]:
𝑵𝒖 = 𝑪𝟏 ∙ 𝑹𝒆𝑫,𝒎𝒂𝒙𝒎 ∗ 𝑷𝒓𝟎.𝟑𝟔 ∙ (𝑷𝒓 𝑷𝒓𝒔)⁄ 𝟏/𝟒
4.6
Where Nu is the Nusselt number, ReD,max is the maximum Reynolds number occurring
within the fin blocks, Pr and Prs are the Prandtl numbers evaluated at mean temperature of flow
and surface temperature, respectively, C1 and m are a constants dependent on ReD,max. For the fin
array used in this investigation the C1 values for air and oil side were 0.8 and 0.9 respectively,
whereas the m value for both was 0.4.
Once the Nusselt number is calculated the heat transfer coefficient for the fin block array
can be calculated using the following equation:
�� = 𝑵𝒖 ∙ 𝒌𝒇 𝑫⁄ 4.7
Where h is the heat transfer coefficient in W·m-2·K-1, 𝑵𝒖 is the Nusselt number, kf thermal
conductivity of the fluid in W·m-1·K-1, and 𝐷 is the diameter of the pin fin in m.
The efficiency of the pins, 𝜼𝒑, fins was calculated by performing a 1-D heat transfer
analysis and assuming an adiabatic tip condition and the following equations:
𝜼𝒑 =𝐭𝐚𝐧𝐡 𝒎𝑳
𝒎𝑳 4.8
63
𝒎 =��𝑷
𝒌𝑨𝒄 4.9
Where P is the perimeter of the fin tip in m and Ac is the cross-sectional area of the fin in
the direction of heat flow in m2 and L is the length of the fin.
For the Overall efficiency of the Fin blocks, the following relation was used:
𝜼𝒐 = 𝟏 − [(𝑨𝒇 𝑨𝑻)⁄ ∙ (𝟏 − 𝜼𝒑)] 4.10
Where 𝜼𝒐 is the overall fin array efficiency, 𝑨𝒇, 𝑨𝑻 are the total areas of pin fins and total
area (Pin Fins + Base) in m2, respectively, and 𝜼𝒑 is the fin efficiency.
The overall thermal resistance of fin block array was found using this relation:
𝑹 = 𝟏 (𝑨𝑻 ∙⁄ �� ∙ 𝜼𝒐) 4.11
Where R is the overall fin array thermal resistance W·K-1
, AT is the total area (Pin Fins +
Base) in m2
, ℎ is the heat transfer coefficient in W·m-2
·K-1
, and 𝜼𝒐 is the overall fin array efficiency.
4.2.3 Spreading Thermal Resistance
Due to the difference between the contact surface areas of the heat exchange surfaces, the
engineering model must account for the thermal spreading resistances. The largest spreading
resistances occur at the interface of the copper spreader plate and the TEG module array in stage 1
for HAU-2.
Lee et al. [93] provide a simple spreading resistance model which was applied to the current
model:
64
𝑹𝒔𝒑 = 𝝓𝒎𝒂𝒙 𝒌𝒔𝒑 ∙ 𝑹𝟏 ∙ √𝝅⁄ 4.12
Where 𝑹𝒔𝒑 is the spreading resistance in W·K-1
, 𝑹𝟏 is the square root of the ratio of area
of TEGs to the constant 𝝅, 𝒌𝒔𝒑 thermal conductivity of the heat spreader in W·m-1
·K-1
and 𝝓𝒎𝒂𝒙
is a constant depending on the geometry of the heat spreader is given by the following equation:
𝝓𝒎𝒂𝒙 = 𝟏
√𝝅(𝟏 − 𝝐)𝚽𝒄 4.13
Where, ε is the dimensionless contact radius given by the ratio of ratio of the hydraulic
diameter of the TEG to the heat spreader and Φc is given by:
𝚽𝒄 = 𝒕𝒂𝒏𝒉 (𝝀𝒄𝝉) +
𝝀𝒄
𝑩𝒊
𝟏 + 𝝀𝒄𝑩𝒊 𝒕𝒂𝒏𝒉 (𝝀𝒄𝝉)
4.14
Where Bi is the Biot number of the heat spreader and τ is the dimensionless spreader
thickness calculated by dividing the thickness of the spreader by its hydraulic diameter. λc is an
empirical parameter given by the following equation:
𝛌𝒄 = 𝛑 + 𝟏
√𝝅𝝐 4.15
4.2.4 Model for Heat Lost to the Ambient
The mathematical model for the current study also accounts for the heat lost to the ambient.
The heat lost was modelled as being dependent on the difference between the temperatures of the
ambient and the average stage temperature on the air side. The following relation was used to
estimate the heat loss for each stage:
𝑸𝒍𝒐𝒔𝒕 = (𝟏 𝑹𝒍𝒐𝒔𝒕⁄ ) ∙ (𝑻𝒂𝒗𝒈 − 𝑻𝒔𝒖𝒓𝒓) 4.16
65
Where 𝑸𝒍𝒐𝒔𝒕 is the heat lost in W, 𝑅𝑙𝑜𝑠𝑡 is sum of all the thermal resistances of hot air, steel
wall, thermal insulation and the ambient air in W·K-1, 𝑻𝒂𝒗𝒈 is the average temperature for air for
the stage, and 𝑇𝑠𝑢𝑟𝑟 is the ambient air temperature.
As illustrated in Figure 4.9, the heat travels from the hot air stream and the thin steel plate,
in through the fiberglass insulation and to the ambient air. The total resistance of the resulting
thermal circuit is a sum of the convective resistance of the hot stream and ambient air, and the
conductive resistance of the steel wall and fiberglass insulation. The convective resistances were
calculated using the built-in EES functions.
Tst,i
Tst,o = Tfg,i
Tfg,o
Tair,av
Tsurr
Rair,av
Rst
Rfg
Rsurr
Tst,i
Tst,o = Tfg,i
Tfg,o
Tair,av
Tsurr
Qlost
Figure 4.1 Thermal Circuit for Heat Lost model
4.2.5 Stage-wise Total Resistance Calculations
With the total resistances for both fluid streams, the heat transfer and temperatures for each
stage can be calculated. For Stage 1 (Figure 4.10) the resistances are calculated as two parts of the
thermal circuit between the node TTEG,H:
𝑹𝒕𝒐𝒕𝒂𝒍,𝑯 = 𝑹𝒇𝒂,𝟏 + 𝑹𝒇𝒃𝒂,𝟏 + 𝑹𝑪𝒔𝒑 4.17
66
𝑹𝒕𝒐𝒕𝒂𝒍,𝑪 = 𝑹𝑻𝑬𝑮 + 𝑹𝑨𝒔𝒑,𝟏 + 𝑹𝒇𝒃𝒐,𝟏 + 𝑹𝒇𝒐,𝟏 4.18
Where 𝑹𝒕𝒐𝒕𝒂𝒍,𝑯 and 𝑹𝒕𝒐𝒕𝒂𝒍,𝑯 are the thermal resistances for the heat delivered to and from
the hot junction and cold of the TEG (QH & Qoil,1) respectively. 𝑹𝒇𝒂,𝟏, 𝑹𝒇𝒃𝒂,𝟏 are the Stage 1 thermal
resistances for air side fins and fin base, respectively. 𝑹𝑪𝒔𝒑, 𝑹𝑨𝒔𝒑,𝟏 are the Stage 1 spreading
resistances for aluminum and copper plates, respectively. 𝑹𝑻𝑬𝑮 is the TEG thermal resistance.
𝑹𝒇𝒃𝒐,𝟏, 𝑹𝒇𝒐,𝟏 are the Stage 1 thermal resistances for oil side fin base and fins, respectively. All
resistances have units of W·K-1.
Rfa,1 Rfba,1 RCsp RTEG RAsp,1 Rfbo,1 Rfo,1
PTEG
QH Qoil,1
Toil,in Toil,int
Tair,in Tair,intQH
Qoil,1
PTEG
TTEG,H TTEG,C
Qlost,1
STAGE 1
Figure 4.2 Stage-1 Thermal Resistance Model
For Stage 2 (Figure 4.11), total thermal resistance is calculated using the following
equation:
𝑹𝒕𝒐𝒕𝒂𝒍,𝟐 = 𝑹𝒇𝒂,𝟐 + 𝑹𝒇𝒃𝒂,𝟐 + 𝑹𝑨𝒔𝒑,𝟐 + 𝑹𝒇𝒃𝒐,𝟐 + 𝑹𝒇𝒐,𝟐 4.19
Here, 𝑹𝒕𝒐𝒕𝒂𝒍,𝟐 is stage 2 total thermal resistance, 𝑹𝒇𝒂,𝟐, 𝑹𝒇𝒃𝒂,𝟐 are the Stage 2 thermal
resistances for air side fins and fin base, respectively. 𝑹𝑨𝒔𝒑,𝟏 is the Stage 2 spreading resistances
for Aluminum plate, 𝑹𝒇𝒃𝒐,𝟐, 𝑹𝒇𝒐,𝟐 are the Stage 2 thermal resistances for oil side fin base and fins,
respectively.
67
Rfa,2 Rfba,2 RAsp,2 Rfbo,2 Rfo,2
Qoil,2
Toil,int Toil,out
Tair,int Tair,out
Qoil,2
Qlost,2
STAGE 2
Figure 4.3 Stage-2 Thermal Resistance Model
4.2.6 Heat Transfer Calculations
For Stage 1, the model solves the following equations simultaneously to find the stage exit
temperatures and heat transfers:
𝑸𝒂𝒊𝒓,𝟏 = 𝑸𝒐𝒊𝒍,𝟏 + 𝑸𝒍𝒐𝒔𝒕,𝟏 + 𝑷𝑻𝑬𝑮 4.20
𝑻𝒐𝒊𝒍,𝒂𝒗𝒈,𝟏 = (𝑻𝒐𝒊𝒍,𝒊𝒏 + 𝑻𝒐𝒊𝒍,𝒊𝒏𝒕)/𝟐 4.21
𝑻𝒂𝒊𝒓,𝒂𝒗𝒈,𝟏 = (𝑻𝒂𝒊𝒓,𝒊𝒏 + 𝑻𝒂𝒊𝒓,𝒊𝒏𝒕)/𝟐 4.22
𝑸𝒂𝒊𝒓,𝟏 = (𝑻𝒂𝒊𝒓,𝒂𝒗𝒈,𝟏 − 𝑻𝑻𝑬𝑮,𝑯)/𝑹𝒕𝒐𝒕𝒂𝒍,𝑯 4.23
𝑸𝒐𝒊𝒍,𝟏 = (𝑻𝑻𝑬𝑮,𝑪 − 𝑻𝒐𝒊𝒍,𝒂𝒗,𝟏)/𝑹𝒕𝒐𝒕𝒂𝒍,𝑪 4.24
𝑸𝒂𝒊𝒓,𝟏 = ��𝒂𝒊𝒓 ∙ [𝒉𝒂𝒊𝒓(𝑻𝒂𝒊𝒓,𝒊𝒏) − 𝒉𝒂𝒊𝒓(𝑻𝒂𝒊𝒓,𝒊𝒏𝒕)] 4.25
𝑸𝒐𝒊𝒍,𝟏 = ��𝒐𝒊𝒍 ∙ [𝒉𝒐𝒊𝒍(𝑻𝒐𝒊𝒍,𝒊𝒏𝒕) − 𝒉𝒐𝒊𝒍(𝑻𝒐𝒊𝒍,𝒊𝒏)] 4.26
𝑸𝒂𝒊𝒓,𝟏, 𝑸𝒐𝒊𝒍,𝟏, 𝑸𝒍𝒐𝒔𝒕,𝟏 are the Stage 1 heat Transfer rates for air, oil and heat loss,
respectively, in W. 𝒉𝒂𝒊𝒓(𝑻𝒂𝒊𝒓,𝒊𝒏𝒕), 𝒉𝒂𝒊𝒓(𝑻𝒂𝒊𝒓,𝒊𝒏) are the enthalpies for air exiting stage 1 and air
inlet, respectively, in J·kg-1. 𝒉𝒐𝒊𝒍(𝑻𝒐𝒊𝒍,𝒊𝒏𝒕), 𝒉𝒐𝒊𝒍(𝑻𝒐𝒊𝒍,𝒊𝒏) are the enthalpies for oil exiting stage 1 and
68
oil inlet, respectively, in J·kg-1, while ��𝒂𝒊𝒓 , ��𝒐𝒊𝒍 are the mass rates for air and oil, respectively, in
kg·s-1. While 𝑻𝒐𝒊𝒍,𝒂𝒗𝒈,𝟏 and 𝑻𝒂𝒊𝒓,𝒂𝒗𝒈,𝟏 are the average temperatures of oil and air side in K, within
the 1st stage.
A similar algorithm is used for stage 2 to find the temperatures of the liquid and gas streams
at the exit and the heat rates in stage 2.
69
4.3 HAU Iterative Development and Experimental Procedure
For the proposed study, two different HAU designs were developed, experimentally tested
and the results were compared to evaluate different HAU design tradeoffs. The first HAU (HAU-
1) was based on a simpler, single stage, compact design.
The HAU enclosure consists of a hexagonal base-plate sub-assembly with welded
sidewalls (Figure 4.4(b)) and a hexagonal cover-plate (Figure 4.4(a)) joined to it with standoffs.
Two staggered, pin-fin arrays, composed of three individual fin blocks, were screwed into each of
the two plates with a thermal contact compound applied at the mating surfaces.
10 cm
Figure 4.4 (a) Cover-plate sub-assembly (b) Base-plate subassembly
For this design, water-blocks were mounted using Aluminum coupling plates on either side
of the HAU, through which the cooling thermal oil flows in an oil-block. Manifolds were used for
thermal-oil flow distribution at both the inlet and outlet. The TEG arrays, comprising of three
individual commercially available TEG modules [94], on either side of this HAU were sandwiched
between the outer surface of the plates and the coupling plate (6 TEG modules in total).
70
Table 4.1 shows the details of the main components used in HAU-1:
Component Dimension (L x W x H) [mm] Material Qty
TEG [94] 56 x 56 x 3 Beryllium Telluride 6
Enclosure 400 x 250 x 100 Carbon Steel 1
Heatsink [95] 80 x 80 x 40 Aluminum 6
Coupling Plate 235 x 108 x 22 Aluminum 2
Oil-block 162 x 40 x 12 Aluminum 2
Table 4.1 HAU-1 component details
(a) (b) (c)
Figure 4.5 (a) TEGs mounted on the outside the HAU (b) Coupling plate mounted on top of the
TEGs (c) Complete assembly of the HAU-1 with oil-block cover plate
Figure 4.5 shows this assembly stack-up of HAU-1, with Figure 4.5-(a) showing the
location of three of the six TEGs and Figure 4.2-(b) shows the coupling plate that houses the water-
block, while, Figure 4.5-(c) shows the completed HAU-1 assembly.
The internal water-block structure and the complete HAU assembly is shown in Figure
4.6.
71
17 cm
4 cm
Coupling Plate
Water-block
internal
structure
Figure 4.6 Assembled HAU-1
4.3.1 HAU Experimental Subsystem Description
Figure 4.7 illustrates the schematic for the HAU evaluation experimental system.
Simulated hot exhaust gas (hot air) from a hot air blower enters the HAU at Point 1, where it
transfers heat to the TEG arrays and exits the system at a lower temperature at Point 2. The TEGs,
mounted on the outside surface of the HAU, convert a portion (~5%) of the acquired heat into
electricity. Much like any other D.C electric source, a TEG array’s maximum power output occurs
when it is connected to a matched DC load [96]. Therefore, the TEG array power is delivered to a
DC load that matches the internal electrical resistance of the TEG array. The cold junction sides of
the TEG modules are cooled by a coupling fluid (thermal oil) through aluminum heat sinks.
Between Points 3→4, the coupling fluid receives heat rejected by TEG array and its temperature
increases. By coupling the thermal oil with a laboratory chiller between 4→5 its temperature is
72
lowered to required inlet temperature to the HAU. A pump, located between 5→1, then recirculates
the coupling fluid through the system.
V
ΔP
P
T
T
T
Matched
DC Load
Pitot Tube Sensor
1
2
45
3
T
HAU Experimental Evaluation System
Lab Chiller
External Chiller
HAU & TEG
Hot Air
Blower
5.6975 V 39.998A
227.887 W 0.199Ω
ON
Figure 4.7 Experimental setup for HAU subsystem evaluation
This description of HAU refers to the first version of HAU (HAU-1) developed during this
study. In the second HAU version (HAU-2) the HAU is divided into two WHR stages. TEGs are
only present in the first stage and, in the second stage, heat is transferred directly between the
exhaust stream and coupling fluid through aluminum heat spreaders.
4.3.2 Design improvements to HAU-2 based on HAU-1 performance
HAU-1 was insulated using fiber glass insulation and its performance was evaluated by
performing experiments over a range of inlet simulated exhaust (air) temperatures and flow rates.
The simulated exhaust gas was provided by a hot air blower, which has controls built-in to regulate
73
the heat rate and flowrate of the air. The average ambient temperature was recorded as 20°C. The
oil mass flowrate, inlet temperature and air mass flowrate were held relatively constant (0.02 kg·s-
1, ~50°C and ~0.01 kg·s-1, respectively) for these experimental runs while the air inlet temperature
was varied. The experiment was performed under a ventilation hood and for each case, the HAU
was operated for 15-30 minutes at steady state and measurements were recorded every 2 seconds
and averaged at the end of the run. A summary of the results is presented in Table 4.1.
Max Blower Heat
& Flowrate
50% of Max Flowrate
& Max Blower Heat
50% of Max Blower
Heat & Max Flowrate
Tair,in (°C) 286±1 382±1 210±1
mair (kg s-1) 0.01±0.005 0.007±0.003 0.008±0.004
Toil,in (°C) 52.6±1 58.2±1 41.8±1
Toil,out (°C) 68.9±1 76.8±1 53.6±1
Qoil,in (kW) 0.67±0.06 0.81±0.06 1.034±0.05
moil (kg·s-1) 0.02±0.0002 0.02±0.0002 0.02±0.0002
ΔPHAU (Pa) 34.3±0.5 17±0.5 17.3±0.5
PTEG (W) 3.505±0.02 4.081±0.02 2.011±0.02
Table 4.2 Results from HAU-1 experimental runs
From the data presented in Table 4.2, it can be seen that the power generation by the TEGs
is lower than the design requirements (described in Section 1.4 → ~25W). Based on the results of
the experiments, the following observations were made:
• Spreading Resistance: As shown in Figure 4.8-(a), the mismatch between the
contact area of the TEG array relative to the larger area heatsink causes high
thermal spreading resistance (Rsp = 0.023 K·W-1). This results in lower hot
junction temperatures for the TEG, the efficiency of a TEG is higher at higher
74
hot junction temperatures for a given cold side temperature. In order to lower
this spreading resistance a thick copper heat spreader can be used (Figure 4.8-
(b)).
• Oil-block sizing: Another factor contributing towards a low TEG power
production is the high oil-block thermal resistance as shown in Figure 4.8 (Rob
= 0.38 K·W-1). This high thermal resistance results in higher than desired
temperatures on the TEG cold junction as the thermal oil cannot recover the
desired amount of heat. This results in a smaller temperature gradient across
the TEGs than desired, leading to a lower efficiency operation of the TEG
array.
Fin
blo
ck
TE
G
Wall
0.005
K/W
0.009
K/W0.0015
K/W
Oil
Block
ṁoil= 0.01 kg/s
ṁair= 0.001 kg/s
RfinRbRTEG
(b)0.38
K/W
Rob H
eat
Sp
read
erRsp
0.013
K/W
Fin
blo
ck
TE
G
Wall
0.005
K/W
0.023
K/W
0.0015
K/W
0.38
K/W
Oil
Block
ṁoil= 0.01 kg/s
ṁair= 0.001 kg/s
RfinRspRTEG Rob
(a)
Figure 4.8 Equivalent Thermal Circuits:(a) No heat-spreader: Largest temperature drop occurs
between the air-side fins and TEG, (b) Copper heat-spreader: Spreading resistance drops and the
largest temperature drop occurs between the two sides of the TEG
• TEG array size: Another contributing factor to the low temperature gradient
across the TEG array was identified as an oversized TEG array. For a higher
75
oil side thermal resistance and lower air side thermal resistance, increasing the
number of TEG results in more surface area for heat extraction, however, the
limiting heat transfer potential of the oil-side heat exchange surfaces results in
“thermal-short circuit” where the TEG’s hot and cold junction temperatures
come closer and little heat transfer can be achieved.
Based on the experimental results and observations from HAU-1, an improved design HAU
(HAU-2) was developed. The following major changes were made:
• The HAU was divided into two stages and the gas side fin-block arrays were
installed only on one side of the HAU.
• Cut-outs were made onto the baseplate sub-assembly allowing for gas side fin-
block bases to be directly mounted onto the coupling plates, thereby, reducing
thermal resistance.
• A high conductivity heat spreader plate, made of copper, was used in stage 1
to minimize spreading thermal resistance due to the difference between contact
surface area of the TEGs and heat exchange surfaces on the gas side.
• The sizing of the oil blocks, especially for stage 1, was increased to allow for
a lower temperature on the TEG cold side.
• The optimal TEG array size was determined based on the thermal-fluid figure
of merit parameter proposed by Rattner [43]. This led to a reduction in the
number of TEGs and a single array composed of two individual TEG modules
was used.
76
Figure 4.9 (a) Thermal Oil side view of HAU-2 (b) Air side view of HAU-2 (c) Illustration of
HAU-2 thermal oil flow path
The pictures in Figure 4.9(a) and Figure 4.9(b) show the dual stage arrangement on the
thermal oil side and simulated exhaust (air) side, respectively. Figure 4.9(c) shows an illustration
of the HAU, with the thermal oil flow path highlighted. The first stage, with the TEG array, has a
co-flow arrangement, whereas the second stage is in a counterflow arrangement.
All thermal contact surfaces were polished and cleaned to ensure good thermal contact.
Thermal paste and thermal adhesives were also applied where required.
60 cm
33 cm
77
For HAU-2, in the first stage, high temperature air passes through a set of fin blocks,
transferring heat to a copper coupling plate that servers as the heat spreader. The two TEG modules
yielded more electricity output than the six identical modules used in HAU-1. This electricity is
supplied to a variable DC load adjusted to match the internal resistance of the TEG array and
thereby operate at maximum power point. In the second stage (no TEGs) an Aluminum coupling
plate is used as spreading resistance is not as great a concern as in stage 1.
To performance of HAU-2 was evaluated over a range of simulated exhaust and thermal
oil input temperatures and flow rates. The tests were performed according to a primary test matrix
which is presented in Table 4.3.
TEST MATRIX
Quantities Varied
Tair,in mair Toil,in moil
250, 312, 400°C 0.005, 0.01 kg s-1 50, 75, 100°C 0.02, 0.01 kg s-1
Quantities Measure
Air and Oil Inlet Temperatures and Flow Rates, Air Side Pressure Drop, DC Load Power, Ambient Temperature
Quantities Calculated
Heat Input to TEG array, Heat Input to Thermal Oil, Heat Lost to Surroundings, Total TEG Power
Table 4.3 Test Matrix for HAU-2
Similar to the test runs for HAU-1, the hot air blower was used to moderate the input
flowrate and heat rate for air. The input temperature for the thermal oil was regulated using an
external recirculating chiller. The experiment was run under a ventilation hood and the average
ambient temperature for the thirty test runs was approximately 20°C. As in the HAU-1 tests, the
system was operated at a steady condition for 15-30 minutes and the measurements were then
recorded every 2 seconds and averaged at the end of each run.
78
A graph illustrating the electrical power delivered by HAU-1 and HAU-2 for different air
inlet temperatures (Toil,in = ~50°C ��𝒐𝒊𝒍 = 0.02 kg · s−1, ��𝒂𝒊𝒓 = 0.01 kg · s−1), is presented in
Figure 4.10:
200 220 240 260 280 300 320 340 360 380 400
2
4
6
8
10
12
14
16
18
20
22 PTEG,tot,HAU-1
PTEG,tot,HAU-2
PT
EG [
W]
Tair,in [°C]
TEG Power Output
Figure 4.10 TEG Power Output Comparison between HAU-1 and HAU-2
From the graph it can be seen that for similar air inlet temperatures and operating
conditions, HAU-2 design generates substantially greater power compared to HAU-1.
Figure 4.11 show the oil outlet temperatures for HAU-2, for the same inlet conditions and
flowrates as described above:
79
200 220 240 260 280 300 320 340 360 380 400
50
60
70
80
90
Toil,out,HAU-1
Toil,out,HAU-2
To
il,o
ut [
°C]
Tair,in [°C]
Outlet Oil Temperature
Figure 4.11 Oil Outlet Temperature Comparison between HAU-1 and HAU-2
The higher thermal oil outlet temperatures for HAU-2 indicate a greater amount of heat
recovered by HAU-2, compared to HAU-1. These comparisons show that the improvements made
to the HAU design have resulted in a better WHR performance efficiency for HAU-2.
80
System Description and Cycle Modeling
81
5.1 System Description
Figure 5.1 illustrates the proposed experimental WHR system cycle. A steady-state cycle
model was developed that applies mass, species, and energy balances to each component assuming
representative closure parameters (e.g., heat exchanger approach temperatures, pump efficiency,
COTS TEG performance parameters). Details on the modeling approach are discussed in Chapter
4.
Figure 5.1 Cycle diagram for the proposed TEG and ARS based Integrated cascaded WHR
system
The absorption system is modeled based on the NH3-LiNO3 working fluid pair, which
avoids the need for a rectifier and permits a low desorber temperature. According to a study by Sun
[97], at lower desorber temperatures (55°C-75°C), NH3-LiNO3 based ARS can achieve higher
COPs better than ammonia-water based ARS.
82
ARSs like ammonia-water based absorption refrigeration require an analyzer and a rectifier
to remove water vapor from the refrigerant mixture leaving the desorber. Since, with the NH3-
LiNO3 working fluid pair, the absorbent is non-volatile, the analyzer and rectifier are not required.
The TEG subsystem shown in Figure 5.1, can be operated separately from the ARS
subsystem by connecting the coupling fluid stream to a heat exchanger coupled to a laboratory
chiller. In this manner the TEG subsystem’s performance can be evaluated independently of the
ARS subsystem.
When the ARS and TEG subsystems are thermally connected by the coupling fluid,
simulated hot exhaust gas (hot air) from a hot air blower enters the HAU at Point 1, in which it
transfers heat to the TEG arrays and exits the system at a lower temperature at Point 2. The TEGs
are mounted on the outside surface of the HAU. The TEGs convert a portion of the acquired heat
into electricity. Much like any other D.C source, a TEG array’s maximum power output occurs
when it is connected to a matched DC load [96]. Therefore, the TEG array power is delivered to a
DC load that matches the internal electrical resistance of the TEG array. The cold junction sides of
the TEG modules are cooled by a coupling fluid (thermal oil) through aluminum heat sinks.
Between Points 3→4, the coupling fluid receives heat rejected by TEG array and its temperature
increases. This description of HAU refers to the first version of HAU (HAU-1) developed during
this study. In the second HAU version (HAU-2), heat transfer to the thermal oil is divided into two
stages. TEGs are only present in the first stage and in the second stage, heat is transferred directly
between the exhaust stream and coupling fluid through aluminum heat spreaders. A more detailed
discussion of HAU design is provided in Chapter 5.
The coupling fluid transfers heat to the pressurized salt solution (NH3+LiNO3) as it flows
through the desorber (Points 4→5). The cooled oil is pumped back into the HAU (Point 3).
83
Dilute, high-pressure, ammonia-lithium nitrate solution (NH3+LiNO3) enters the desorber
at Point 6 and heat transfer from the oil stream desorbs ammonia vapor (NH3). At the outlet of the
desorber, the ammonia vapor and concentrated solution are separated in a small separator tank. The
concentrated NH3+LiNO3 solution enters the recuperative solution heat exchanger (SHX, Point
7→8) and cools as it preheats the dilute solution (Points 6→12). Downstream of the SHX, the
concentrated solution expands through a valve to a lower pressure before entering the absorber at
Point 10.
The NH3 vapor exiting the separator tank liquefies as it rejects heat to the ambient in the
condenser (Points 13→14). The liquid refrigerant is recuperatively subcooled as it flows through
the refrigerant pre-cooler (RPC, Points 14→15). It is then expanded through a valve between Point
16→17. The NH3 refrigerant evaporates in the evaporator, delivering cooling to a thermal load
(Points 17→18). The refrigerant vapor flows through the RPC, precooling the refrigerant liquid
and it exits the RPC at Point 19.
In the experimental facility, the evaporator thermal load is simulated by a stream of 50/50
ethylene glycol-water mixture (between Pts. 20→21). The inlet temperature of the ethylene glycol
mixture is controlled with an electric heater driven by an automated process controller.
The NH3 vapor exiting the evaporator mixes with the low pressure concentrated
NH3+LiNO3 solution (Point 9) at the inlet to the absorber (Point 10). The vapor absorbs into the
solution, rejecting heat to the ambient, and exits at Point 11. The liquid mixture is pumped to high
pressure (Points 11→12), and continues to the SHX where it is recuperatively preheated to Point
6, completing the cycle.
84
The absorber and condenser reject heat to the ambient air (nominally at room temperature)
with the aid of fans mounted on these heat exchangers. The fan speeds can be continuously varied
in the experimental facility.
The next section of this chapter discusses the instrumentation, measurement uncertainties
and data collection method.
5.2 Cycle Model
For the proposed study, steady state thermodynamic models for refrigerated transport
vehicle, carburizing furnace application and the scaled experimental facility were developed using
the Engineering Equation Solver (EES) [81] software. In case of vehicle application model and
experimental model, the cooling delivery temperature is -15°C and the ambient temperature is
30°C, and for the carburizing furnace model the cooling delivery temperature is 5°C. A constant
exhaust inlet temperature of 500°C was also assumed. The cooling delivery temperature of -15°C
for the refrigerated transport truck is the temperature of the refrigerated enclosure This chapter
discusses the approach used in developing the three models.
5.2.1 Conservation Laws
The conservation laws applied in these models are described below:
• Mass conservation:
All three models apply mass conservation law to the overall system and at the component
level (heat exchanger, pump, valve, etc.). When applied to a component of the ARS, the total mass
85
entering each component equals the mass flow out. This is represented mathematically in Equation
5.1:
∑ ��𝒊𝒏 = ∑ ��𝒐𝒖𝒕 5.1
• Species conservation:
In addition to mass flowrate, the models also apply species conservation at the component
level. Salt mass fraction, X, is described as the ratio between the mass of salt (NH3-LiNO3) and
total mass (ammonia + NH3+LiNO3). So, for instance, in the condenser where the only specie is
ammonia, X = 0. The specie conservation is expressed mathematically in Equation 5.2:
∑ ��𝒊𝒏 ∙ 𝑿𝒊𝒏 = ∑ ��𝒐𝒖𝒕 ∙ 𝑿𝒐𝒖𝒕 5.2
• Energy Balance:
The models apply energy balance at both component level and overall system level. At the
system level, the energy inputs to the system are in the form of waste heat recovered from the
exhaust stream, heat received by the evaporator and the pump work. Energy output from the system
is in the form of TEG electrical power generation, absorber and condenser heat rejection. A similar
approach is used at the component level where energy balance law is applied at the inlet and outlet
of each component.
86
5.2.2 Closure Parameters
In developing the cycle models, several closure parameters were used for the in this study.
The closure parameters for cooling delivery temperature and ambient temperature for the ARS
subsystem were based on the guidelines provided by AHRI [72].
A 10% desorber outlet vapor quality was assumed for all three models and reasonable
closest approach temperature (CAT) for heat exchangers, superheat for evaporator and subcooling
for absorber and condenser were also assumed. A reasonable pump efficiency of 50% was also
assumed as a closure parameter. Table 5.1 lists these values:
Component Parameter
Desorber CAT 10 K
DoSH 2 K
SHX CAT 15 K
RPC CAT 10 K
Condenser CAT 10 K
DoSC 2 K
Evaporator CAT 5 K
DoSH 2 K
Condenser CAT 10 K
DoSC 2 K
Absorber CAT 10 K
DoSC 2 K
Table 5.1 Heat Exchanger Closure Parameters for ARS-TEG
In the above table CAT, DoSH and DoSC stand for ‘Closest Approach Temperature’,
‘Degree of Superheat’ and ‘Degree of Subcooling”, respectively. Additionally, the CAT between
the liquid and vapor phase at the desorber outlet was assumed to be 2 K. Moreover, the CAT from
the exhaust stream to the TEG hot side is assumed to be 75 K and CAT form the TEGs to the
desorber is assumed to be 40 K. To model the commercial off-the-shelf (COTS) TEG module [94]
87
considered for this study, the closure parameters adopted include the assumption that the modules’
electrical resistance, conductivity and Seebeck coefficient, provided by the manufacturer, are
constant with respect to temperature. Reasonable CATs were assumed between TEG array and the
exhaust stream, and between TEG array and desorber.
Based on the conservation laws and closure parameters discussed above, desorber balances
are calculated using the following equations:
��𝟔 = ��𝟏𝟑 + ��𝟕 5.3
��𝟔𝑿𝟔 = ��𝟏𝟑𝑿𝟏𝟑 + ��𝟕𝑿𝟕 5.4
��𝟔𝒉𝟔 + 𝑸𝑫𝑬𝑺 = ��𝟏𝟑𝒉𝟏𝟑 + ��𝟕𝒉𝟕 5.5
Where ṁ6 is the mass flowrate of the concentrated solution (ṁcs) at the inlet of the desorber
and ṁ13 and ṁ7 are the mass flowrates of the refrigerant (ṁref) and dilute solution (ṁds) exiting the
desorber, respectively. The symbol h represents the enthalpy with the respective state point as its
subscript.
For the condenser the heat rejection rate is calculated using the following equation:
𝑸𝒄𝒐𝒏𝒅 = ��𝑟𝑒𝑓(𝒉𝟏𝟑 − 𝒉𝟏𝟒) 5.6
The cooling delivered by the evaporator is calculated using equation 5.7:
𝑸𝒆𝒗𝒂𝒑 = ��𝑟𝑒𝑓(𝒉𝟏𝟖 − 𝒉𝟏𝟕) 5.7
The absorber’s heat rejection rate is calculated using equation 5.8:
𝑸𝒆𝒗𝒂𝒑 + ��𝑐𝑠𝒉𝟏 = ��𝑟𝑒𝑓𝒉𝟏𝟗 + ��𝑑𝑠𝒉𝟗 5.8
The work done by the pump is calculated using equation 5.9:
��𝒑 = ��𝒄𝒔(𝒉12 − 𝒉11) 5.9
To calculate the overall COP of the combined system, Equation 5.10 is used:
88
𝑪𝑶𝑷 = 𝑸𝑒𝑣𝑎𝑝/𝑸𝑫𝑬𝑺 5.10
Finally, with all the heat rates, and inlet and outlet temperatures calculated for the heat
exchangers, UA, for each heat exchanger can be calculated using Equation 5.11:
𝑼𝑨 = 𝑸/𝑳𝑴𝑻𝑫 5.11
Where Q is the rate of heat transfer in a heat exchanger and LMTD is the log-mean
temperature difference of the heat exchanger in question.
5.2.3 Results
With the closure parameters as input and by using the conservation laws described above,
the model calculates, among other quantities, component wise input and output temperatures and
pressures, quality of vapor, composition of species and, HX overall heat transfer coefficients and
heat rates. A summary of results from the model for the two applications and experimental facility
considered in this study are provided in Table 5.2.
Application Texhaust(°C) QTEG,in(kWth) Qelec (kWe) Qcooling(kWth) ηe(%) COP
Carburizing Furnace 500 50 5.04 26 6.5 0.6
Vehicle Exhaust 500 18 0.564 9.23 6.5 0.53
Experimental Model 500 1.8 56 0.923 3 0.5
Table 5.2 Summary of model results
Table 5.3 shows state point-wise (corresponding to Figure 6.1) results for the carburizing furnace
application:
State Point m [kg s-1] h [J kg-1] P [Pa] T [C] X
6 0.15 70002.00 1643000.00 89.47 0.53
7 0.13 159021.00 1643000.00 115.90 0.62
8 0.13 -30021.00 1643000.00 55.00 0.62
9 0.13 -30021.00 398341.00 55.41 0.62
89
11 0.15 -93125.00 398341.00 40.00 0.53
12 0.15 -90684.00 1643000.00 40.41 0.53
13 0.02 1509000.00 1643000.00 117.90 0.00
15 0.02 190790.00 1643000.00 40.00 0.00
16 0.02 116352.00 1643000.00 24.64 0.00
17 0.02 116352.00 398341.00 -2.00 0.00
18 0.02 1265000.00 398341.00 0.00 0.00
19 0.02 1340000.00 398341.00 30.00 0.00
Table 5.3 Point-wise results for ARS-TEG cycle
90
Experimental Results
91
6.1 Experimental Facility
Figure 6.1 shows the experimental lab facility developed for the current study. The facility
was designed to deliver up to 50 W of electricity and 1 kW of cooling with a 3.7 kW electric hot
air blower simulating a waste heat source. The facility is designed to approximate a 1/10th of the
scale of the proposed refrigerated transport vehicle WHR system, which could provide 500 W of
electrical power and 10 kW of cooling. This experimental facility is operated under a ventilation
hood with curtains that can be closed when the system is charged with ammonia.
THERMAL SYSTEMEXPERIMENT CONTROLS
(a) (b)
HAU
Figure 6.1 Experimental Facility: (a) Experimental Facility highlighting experiment controls
(including DC load) and thermal system with the location of HAU highlighted, (b) Picture shows
the HAU-1 model and simulated exhaust (hot air blower)
The scaled experimental system shown above, was tested according to application specific
test matrix to simulate different operating conditions of the vehicle. The lower exhaust temperatures
and flowrates simulate vehicle’s low RPM operation whereas, the higher exhaust temperatures and
flowrates represent vehicle operating at a higher RPM. This matrix is presented in Chapter 4 –
Table 4.3.
92
Table 6.1 shows a control matrix that details the different key parameters that will be
controlled, measured or calculated.
CONTROL MEASURE CALCULATE
HAU
Simulated Exhaust Gas mair,in and Tair,in Tair,out Qlost
Coupling Fluid moil,in and Toil,in Toil,out Qoil,in
TEG Modules PTEG ηTEG
Refrigeration
Circuit
Pump mpump ppump,out, Tpump,out Ppump
Refrigerant pevap,in and Tevap,in mref ,Tref,out
Simulated Cooling Load mCL, TCL,in TCL,out Qload, COP
Table 6.1 Control Matrix for Experimental Facility
6.2 Instrumentation
Table 6.2 lists the different types of instruments used to measure temperatures, flowrates
and pressures for the four fluid loops in the WHR system. Also included in the table are the
measurement uncertainties associated with these measurements. Almost all of the temperature
measurement devices were calibrated against a high accuracy reference thermometer, yielding
estimated uncertainties of ±0.25°C. The high temperature thermocouples for the simulated exhaust
gas stream were not calibrated, and therefore are assumed to have ±1°C uncertainty.
For data collection, a data acquisition unit (DAQ) is interfaced with a PC running LabView
[98], and a LabView code was setup to save the collected measurements to a spreadsheet file.
The TEG power output is recorded manually. The readings from the DC resistive load
monitor are recorded twice; once at the start of the experimental run when the system achieves a
relative steady state, and again just before the experimental run is complete. The final reading ,
which is the average of the two readings, has an uncertainty of ±0.2% [99].
93
Temp (°C) Flowrate Pressure (Pa) Type Unc Type Unc Type Unc
Exhaust K ±1 Diff Pressure ±0.15 L m-1 Diff Pressure ±0.31
Coupling Fluid T ±0.25 Oval Gear ±0.04 L m-1 Transducer ±6 kPa
NH3-LiNO3 T ±0.25 Ultrasonic ±0.02 L m-1 Transducer ±6 kPa
NH3 T ±0.25 Ultrasonic ±0.006 L m-1 Transducer ±6 kPa
Ethelene Glycol T ±0.25 Oval Gear ±0.04 L m-1 - -
Table 6.2 Measurement types and uncertainties for ARS-TEG experimental facility
6.3 Exergetic Efficiency
The theoretical maximum heat available for recovery from an exhaust stream is the amount
of energy the stream must reject to achieve thermal equilibrium with the ambient. The ratio between
this theoretical maximum and the actual waste heat recovered by a WHR system is known as the
exergetic efficiency of a WHR system. For HAU-2, the exergetic efficiency is defined as:
𝜼𝒆𝒙 = (∆𝑬��𝒐𝒊𝒍 + 𝑷𝑻𝑬𝑮) 𝑬��𝒂𝒊𝒓,𝒊𝒏⁄ 6.1
Where 𝜼𝒆𝒙 is the Exergetic Efficiency of the HAU, ∆𝑬��𝒐𝒊𝒍 is the rate of change of
exergy of the oil as it passes through the HAU, in W. PTEG is the total electric power generated by
the TEG array in W and 𝑬��𝒂𝒊𝒓,𝒊𝒏 is the exergy rate of the air at inlet.
Exergetic efficiency, defined this way for a WHR application, indicates the ratio of useful
heat and electrical power recovered from an input hot stream of specific exergy content. In Figure
6.2, the exergetic efficiencies are graphically presented for different input temperatures and mass
flowrates of the hot and cold streams. It can be seen form the graphs that the exergetic efficiencies
are higher for high thermal oil inlet temperatures. This is expected as the smaller the difference
between delivery and source temperature for a WHR system, the higher is its exergetic efficiency.
94
0.005 0.01 0.005 0.01
0.01 0.02 ṁoil
ṁair
0.0
0.1
0.2
0.3
0.4
0.5
ηex [
-]
Toil,in = 50°C
[kg·s-1]
[kg·s-1]
(a)
0.005 0.01 0.005 0.01
0.01 0.02
0.0
0.1
0.2
0.3
0.4
0.5
ηex [
-]
250[°C]
313[°C]
400[°C]
Toil,in = 75°C
(b)
Exergetic Efficiency
0.005 0.01 0.005 0.01
0.01 0.02 ṁoil
ṁair
0.0
0.1
0.2
0.3
0.4
0.5Toil,in = 100°C
ηex [
-]
[kg·s-1]
[kg·s-1]
(c)
Tair,in
Figure 6.2 Exergetic Efficiency of HAU-2 at different inlet conditions
Another aspect to consider is the mass flowrates of the thermal oil stream and simulated
exhaust (air) stream. From the graphs, it can be observed that, within the margin of error, the
exergetic efficiency is strongly sensitive to the air mass flowrate, with greater exergetic efficiencies
recorded at higher air flowrate when other parameters remain unchanged. Since, air has a much
lower thermal capacity compared to the thermal oil, a higher flowrate of air results in a better
source-use match in terms of heat capacity rates, resulting in higher exergetic efficiencies.
Thus, a thermal oil inlet temperature closer to the hot exhaust stream inlet temperature, and
a higher air mass flow rate result in higher exergetic efficiencies.
95
6.4 Thermoelectric Power
Owing to the relatively small amount of waste heat recovered in the form of TEG electrical
power generation, exergetic efficiency, as defined previously, does not inform greatly as regards to
TEG performance. For this, the total TEG power generated for each experimental run is analyzed
in this section.
0.005 0.01 0.005 0.01
0.01 0.02 ṁoil
ṁair
0
5
10
15
20
25
PT
EG
,to
t [W
]
[kg·s-1]
[kg·s-1]
Toil,in = 50°C
(a)
0.005 0.01 0.005 0.01
0.01 0.02
0
5
10
15
20
25
PT
EG
,to
t [W
]
250[°C]
313[°C]
400[°C]
Toil,in = 75°C
(b)
Total TEG Power Tair,in
0.005 0.01 0.005 0.01
0.01 0.02 ṁoil
ṁair
0
5
10
15
20
25
PT
EG
,to
t [W
]
[kg·s-1]
[kg·s-1]
Toil,in = 100°C
(c)
Figure 6.3 TEG Power produced by HAU-2 at different inlet conditions (uncertainty of ±0.2%)
.
96
Graphs in Figure 6.3, for different inlet temperature and mass flowrate conditions, show
the TEG power produced for all 30 experimental runs. It can be observed that the thermoelectric
power output is greatly sensitive to the temperature difference between the hot and cold stream at
the inlet of the HAU-2.
For given inlet temperatures and air mass flowrate, the oil mass flowrate has no significant
impact on the total TEG power generated. But for the cases where the air mass flowrate is reduced,
there is a significant drop in TEG power output. This sensitivity to air mass flowrate is caused by
the much higher thermal capacity of the thermal oil stream relative to the thermal capacity of the
air stream.
Thus, TEG power is greater when the temperature difference between the fluid streams and
the flowrate of air are relatively high. Additionally, for the same hot and cold side temperatures,
the increase in the thermal fluid side flowrate only minimally affects the TEG power production,
instead, TEG power is more sensitive to the air flowrate.
6.5 Comparison of HAU Model Predictions and Experimental Results
The model developed for HAU-2 (Section 4.3) takes as inputs: the time averaged flowrates
of thermal oil stream and simulated exhaust (air) stream, the inlet temperatures of those streams,
and the average TEG power produced for each experimental run. The model calculates the outlet
temperatures, air side pressure drop, heat loss rate, heat transfer rates in each stage, and overall heat
transfers for each stream.
Figure 6.4 and Figure 6.5 present the comparison between the model’s predicted and the
experimentally measured Qoil and Qair, respectively, for all the experimental cases.
97
The average relative error, and the average absolute relative error in the predicted values
of Qoil is ~0.3, respectively. The average relative error, and the average absolute relative error in
the predicted values of Qair is ~0.2.
The model output of Qoil depends on the Tair,out , as shown by equations 6.2 and 6.3.
However, there is an overestimate error associated with the Tair,out measurement (discussed in the
Appendix).
𝑸𝒐𝒊𝒍 = 𝑸𝑎𝑖𝑟 − 𝑸𝒍𝒐𝒔𝒕 − 𝑷𝑻𝑬𝑮 6.2
𝑸𝒂𝒊𝒓 = ��𝒂𝒊𝒓𝒄𝒑,𝒂𝒊𝒓(𝑻𝒂𝒊𝒓,𝒊𝒏 − 𝑻𝒂𝒊𝒓,𝒐𝒖𝒕) 6.3
Meanwhile, the experimental measurement of Qoil is dependent on the Toil measurements
which have low uncertainties associated to them. In general, the model prediction for both Qoil and
Qair are consistent with the trend of experimental observation of greater heat transfer rates for larger
temperature differences between the hot and cold stream.
Moreover, due to the lower heat capacity rate of air compared to the thermal oil, the heat
transfer rate increase with an increase in air flowrate when other parameters are kept constant.
98
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02ṁoil
ṁair
Tair,in
0
200
400
600
800
1000
1200
1400
1600
1800Q
oil
,mo
d [
W],
Qo
il,e
xp [
W]
[kg·s-1]
[kg·s-1]
[°C]
Toil,in = 50°C
(a)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
200
400
600
800
1000
1200
1400
1600
1800
Qoil,mod Qoil,exp
Toil,in = 75°C
(b)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
200
400
600
800
1000
1200
1400
1600
1800Toil,in = 100°C
(c)Qoil Comparison
Figure 6.4 Qoil Model vs Experiment Comparison
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02ṁoil
ṁair
Tair,in
0
200
400
600
800
1000
1200
1400
1600
1800
Qai
r,m
od [
W],
Qai
r,ex
p [
W]
[°C]
[kg·s-1]
[kg·s-1]
Toil,in = 50°C
(a)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
200
400
600
800
1000
1200
1400
1600
1800
Qair,mod Qair,exp
Toil,in = 75°C
(b)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
200
400
600
800
1000
1200
1400
1600
1800Toil,in = 100°C
(c)Qair Comparison
Figure 6.5 Qair Model vs Experiment Comparison
99
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02ṁoil
ṁair
Tair,in
0
20
40
60
80
100
120
140T
oil
,ou
t,m
od [
°C],
To
il,o
ut,
exp [
°C]
[kg·s-1]
[kg·s-1]
[°C]
Toil,in = 50°C(a)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
20
40
60
80
100
120
140
Toil,out,mod Toil,out,exp
Toil,in = 75°C
(b) Toil,out Comparison
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
20
40
60
80
100
120
140 Toil,in = 100°C(c)
Figure 6.6 Toil Model vs Experiment Comparison
0.02
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01ṁoil
ṁair
Tair,in
0
20
40
60
80
100
120
140
160
180
200
Tai
r,o
ut,
mod [
°C],
Tai
r,out,
exp [
°C] Toil,in = 50°C
[°C]
[kg·s-1]
[kg·s-1]
(a)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
20
40
60
80
100
120
140
160
180
200
Tair,out,mod Tair,out,exp
Toil,in = 75°C
(b)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
20
40
60
80
100
120
140
160
180
200Toil,in = 100°C
(c)Tair,out Comparison
Figure 6.7 Tair Model vs Experiment Comparison
100
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02 ṁoil
ṁair
Tair,in
0
20
40
60
ΔP
mo
d [
Pa]
,ΔP
exp [
Pa]
[°C]
[kg·s-1]
[kg·s-1]
Toil,in = 50°C
(a)
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02
0
20
40
60
ΔP
mo
d [
Pa]
,ΔP
exp [
Pa]
ΔPmod
ΔPexp
Toil,in = 75°C
(b)
ΔP Comparison
0.005 0.01 0.005 0.01
250 313 400 250 313 250 313 400 250 313
0.01 0.02 ṁoil
ṁair
Tair,in
0
20
40
60
ΔP
mo
d [
Pa]
,ΔP
exp [
Pa]
[°C]
[kg·s-1]
[kg·s-1]
Toil,in = 100°C(c)
Figure 6.8 ΔPexp vs ΔPmodel comparison
In Figures 6.6 and 6.7, a comparison between the engineering model’s predicted Toil,out and
Tair,out with their corresponding experimental measurement is presented.
The absolute relative and relative errors in the predicted values of Toil,out are ~0.06 and -
0.06, respectively. Meanwhile, the absolute relative and relative errors in the predicted values of
Tair,out are both ~0.25 in value.
101
It can be seen that, for all cases, the predicted Toil,out has a much greater degree of qualitative
similarity (within the margin of error) with the measured Toil,out, while the error in the model
predicted Tair,out is greater than for Toil,out.
Another comparison to consider is the pressure drop in the air stream as it goes through the
HAU. From the graphs presented in Figure 6.8, it can be seen that the pressure drop is sensitive to
the air mass flowrate. The model’s prediction of pressure drop in the HAU is consistent with the
experimentally measured value (within the margin of error). The absolute relative and relative
errors in the predicted values of ΔP are ~0.14 and ~-0.13, respectively
6.6 Discussion
Two different designs for an integrated Thermoelectric Generator and Absorption
Refrigeration based WHR heat acquisition units (HAU) were developed. Their performance was
evaluated at different temperatures and flowrates. Through the analysis of the experimental results
from the first design of the HAU (HAU-1), important design factors were identified:
• Spreading Resistance is an increased thermal resistance when there exists an
area mismatch between the TEGs and the air-side heat exchange surface. This
leads a lower TEG hot junction temperature and less efficient TEG operation.
• An undersized oil-block results in higher TEG cold junction temperatures and
a reduction in TEG efficiency. An undersized oil-block also results in lower
thermal oil outlet temperatures.
102
• Optimal TEG array sizing is important in designing an HAU as an increase in
the number of TEGs results in a larger heat exchange area between the hot and
cold stream, thereby reducing the overall ΔT across the TEG junctions.
Based on these parameters, an improved HAU (HAU-2) was developed and experimentally
tested. The following changes were made to the TEG design:
• To increase thermal energy transfer to the thermal oil, HAU-2 was divided into
two-stages, with the TEGs present in the first stage. The first stage allows the
TEGs to operate at higher ΔT while the second stage recovers more heat
downstream of the TEG which increases the thermal oil outlet temperature
compared to the one-stage design of HAU-1.
• To minimized the effect of spreading resistance, a copper heat spreader was
utilized in the first stage.
• Based on Rattner’s [43] work, the optimum number of TEGs (2) were used in
HAU-2 as opposed to 6 TEGs in HAU-1.
These changes resulted in a HAU system with 4 times more TEG power output than HAU-
1. Moreover, the exergetic efficiencies for the HAU-2 (~40%) were found to be comparable to
ORC based WHR systems [100], [101]. The experimental results also indicate a great degree of
sensitivity of key system variables to the exhaust/air-side flowrate.
An engineering model for HAU-2 was developed and system performance parameters were
compared to the experimental observations. A general congruence of trend between the model’s
predictions and experimental measurements was observed after air-side flowrate corrective
calibration (see Appendix), though some error still exists due air-side temperature outlet
measurements.
103
Therefore, current study establishes that a cascaded approach to WHR based on TEG and
ARS is a viable pathway for WHR especially for application where cooling and electric power are
the desired modes of WHR.
104
Conclusions and Recommendations for Future
Research
105
This chapter provides a discussion of the major findings of this Ph.D. dissertation with
regards to ARS-TEG based integrated cascaded waste heat recovery systems. Thereafter, some
recommended pathways for future research on this topic are also suggested.
7.1 ARS-TEG thermoeconomic studies
Chapter 3 discusses the thermoeconomic studies conducted for the vehicle and carburizing
furnace applications of the proposed ARS-TEG based WHR system. Cycle models were developed
based on AHRI guidelines for closure parameters of heat exchangers and reasonable component
efficiencies were assumed. Another complementary study compared the performance of ORC-
VCC and ARS-TEG based WHR systems for vehicle application and compared the capital
investment cost of the two systems. The key findings for the thermoeconomic studies are:
1. The payback periods for ARS-TEG based cascaded WHR systems are competitive
with conventional ORC based WHR systems.
2. The temperature drop across the TEGs is a significant design parameter, as it affects
the efficiency of both pathways (ARS and TEGs). If the ΔT across the TEG junctions
is too large, the desorber heat delivery temperature can drop below the required
temperature.
3. In light of the real-world data [51] for a refrigerated transport vehicle’s exhaust
temperaure fluctuations, strategies must be employed to ensure a steady operation of
the WHR system. It was proposed in this disseration that electric heaters may be used
to deliver heat to the desorber when the exhaust tempeatures are too low. At higher
than design temperatures, the exhaust can be re-routed to bypass the desorber.
106
4. By comparing the captial cost of the ARS-TEG system to an ORC-VCC system, it was
surmized that the cost of the ARS-TEG sysytem ($10,464.05), without the TEGs was
significanly lower than the cost of the ORC-VCC system ($15,390.7). But the TEGs
cost another $10,269 for the ARS-TEG system, making it slightly more expensive.
Since, the electrical power requirement of a refrigerated transport vehicle is not the
primary requirement, a reduction in the number of TEGs can result in a more
competitive ARS-TEG based WHR system.
7.2 HAU design
In Chapter 4 of this Ph.D. dissertation a detailed study of the iterative design and
development of HAU for the ARS-TEG based WHR system was presented. Two HAUs were
constructed during the course of this study. Findings from the performance of the first HAU
informed the design of HAU-2. Some of the key findings form this section of study are:
1. Spreading resistance due to the mismatch in contact area of the TEGs and fins on the
air-side results in a large spreading resistance (Rsp = 0.023 K W-1). This results in a
lower TEG hot side temperature and lower TEG efficiency. Using a copper heat
spreader can minimize the spreading resistance (Rsp = 0.013 K W-1).
2. The number of TEGs in a WHR system must be optimized as increasing the number
of TEGs beyond an optimal number can lead to poorer WHR performance as the
average ΔT across the TEGs drops. For HAU-1, using 6 TEGs the maximum power
produced for a specific set of inlet conditions was ~4 W, whereas, for the same inlet
conditions, HAU-2 generated a maximum of ~21 W.
107
3. Heat exchange surface sizing on both the hot and cold sides of an HAU requires
optimization. Backpressure, especially on the hot exhaust side of the HAU, is a
significant design parameter as an increase in backpressure required additional work.
If the heat exchange surfaces on the thermal oil side are undersized, the TEG cold side
temperature and hot side temperature ‘pinch’ and the efficiency of the TEGs drops.
The thermal fluid’s outlet temperature is also lower than required if the oil side heat
exchange surfaces are undersized.
7.3 HAU Model prediction and experimental results
In Chapter 4 and Chapter 5 a detailed description of the HAU and cycle models is
presented. In Chapter 6 of this dissertation a detailed comparison of the model predication and
experimentally observed values is presented. Some of the key findings are:
1. Due to the smaller heat capacity (cp) of the exhaust side flow, parameters like Qoil,
PTEG and Toil are more sensitive to the inlet conditions on the air side.
2. The exergetic efficiency of the HAU is comparable to that of a typical ORC based
WHR systems (~40%).
7.4 Recommendations for future research
More studies are needed to improve upon the ARS-TEG based WHR system proposed in
this Ph.D. dissertation. My recommendations for future works include:
108
7.4.1 Advanced HAU design
The HAU developed during this study can be improved upon by eliminated the thermal
fluid coupling fluid between the HAU and desorber. This can be achieved by designing and
developing an integrated desorber where the heat from the exhaust cascades through the TEGs and
is delivered to the ARS working fluid solution directly. This has the potential to significantly
increase the performance of the ARS-TEG waste heat recovery system by:
• Delivering higher-grade waste to the ARS sub-system
• Eliminating parasitic heat losses due to the thermal oil flow circuit
• Reduction in cost by eliminated additional heat exchange surfaces and thermal oil
as working fluid
7.4.2 Experimental Investigations
In addition to the HAU specific experimental work reported in this Ph.D. dissertation
(Chapter 5 and Chapter 6), additional full system (ARS+TEG) experimental data is needed to gain
further insights into the operation of the WHR system at different conditions.
109
Appendex A: Velocity Measurement
Calibration
Preliminary analysis of the experimental data indicated the presence of an over-estimate
error in measurements related to the air-side flow. It was suspected that the temperature and
flowrate measurement locations at the exit of the HAU were possibly within the Vena Contracta
effect as shown in Figure A.1.
Vena Contracta Effect
Pitot TubeThermocouple
Figure A.1 Stage-2 Airflow, immediately at the exit of the HAU, is contracted due to the shape
of the HAU. This contraction leads to an increase in the velocity at the location of temperature
and velocity measurement
To minimize the effect of this error, a post-experiment, flowrate calibration was performed
as a correction for the measure data by taking flow measurements downstream using a hotwire
anemometer (Figure A.2). However, some error due to Tair,out measurements still persists.
110
Figure B.2 Stage-2 Vcorr (Velocity measurement downstream of original measurement) vs Vmsrd
(original measurement)
0
0.5
1
1.5
2
2.5
3
3.5
0 0.5 1 1.5 2 2.5 3 3.5 4
Vco
rr[m
/s]
Vmsrd [m/s]
Velocity Correction – Vmsrd vs. Vcorr
111
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Vita
Shahzaib B. Abbasi
Education
Ph.D., Mechanical Engineering, The Pennsylvania State University December 2020
M.S., Pakistan Institute of Engineering and Applied Sciences November 2013 B.E., N.E.D University of Engineering and Technology April 2010
Work Experience
Graduate Researcher – University of Twente, Enschede, The Netherlands Feb. 2020 – Present Doctoral Researcher – The Pennsylvania State University, State College, PA Jan. 2017 – Dec. 2020
Graduate Researcher – Mississippi State University, Starkville, MS Aug. 2015 – Dec. 2016
Junior Engineer – National Center for Non-destructive Testing, Islamabad, Pakistan Nov. 2013 – Jun. 2015 Fellow – P.I.E.A.S, Islamabad, Pakistan Nov. 2011 – Nov. 2013
Product Executive (Renewable Energy) – Makkays, Islamabad, Pakistan Aug. 2010 – Feb. 2011
Selected Publications
1. Abbasi, Shahzaib B., and Alexander S. Rattner. "Cascaded Thermoelectric Generation and Absorption
Refrigeration Waste Heat Recovery." ASTFE Digital Library. Begel House Inc., 2018.