Performance and Reliability of WHRUs

7/25/2019 Performance and Reliability of WHRUs

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Performance and Reliability of Water Jacket

Exhaust Gas Waste Heat Recovery Units

Koh Chuan Heng Erik1, MEng(Hons), MSc

SYNOPSIS

Exhaust gas Waste Heat Recovery (WHR) is a proven technique where energy from exhaust gases producedby power plants that would have been lost, is recovered for useful purposes such as heating and powergeneration, thereby reducing operational costs and emission. Research and development efforts to improve

the performance and reliability of waste heat recovery units (WHRUs) are ongoing as they are the mostcrucial, but also one of the most vulnerable components in exhaust gas WHR systems. Exposure to harshoperating conditions such as adverse temperature gradients and the corrosive nature of exhaust gases wereidentified as common causes of failure. Although heat recovery performance increases with bigger

temperature difference between the exhaust gas and the working fluid, adverse temperature gradients inducesthermal stresses within and between components resulting in premature failures. This paper presents theeffects of exhaust tube length-to-diameter (L/d) ratio and placement of coolant inlet and outlet on temperatureprofile, heat recovery performance, and exhaust side pressure drop in a counter-flow water jacket WHRU. A

water jacket WHRU configuration minimises redesign requirements during the retrofit of existing powerplants with WHR. The configuration also allows modularity, and offers better access for inspection andmaintenance. Non-dimensional parametric studies of the WHRU were conducted using analytical andComputational Fluid Dynamic (CFD) models. Exhaust gas Reynolds numbers between 20,000 and 400,000,

representative of exhaust gas flow in the exhaust stacks of the MEP803A diesel generator and Rolls Royce501-K17 gas turbine generator were used in this study. Results indicate heat recovery improved withincreasing L/d. A key finding from the study revealed optimally positioned inlet and outlet of the water jacketimproved heat recovery by up to 19%, and was also very effective at mitigating adverse temperature profiles.

This in turn contributes improves the reliability of the exhaust gas WHRU.

INTRODUCTION

Fossil fuel based power plants such as diesel engines and gas turbines are commonly used on naval ships and shore

installations for propulsion and power generation. Despite technological advances, less than 40% of the energy from

the fuel consumed is actually used for its intended purpose. Nearly 40% of the fuel energy is lost via exhaust gases

[1], [2]. With depleting sources of fossil fuel, increasing concerns of harmful emissions and growing requirements

for energy efficiency, the impetus to research and develop technologies such as waste heat recovery is expected to

increase [3]. Exhaust gas WHR is a proven technology whereby a portion of energy from exhaust gases is recovered

for useful purposes such as heating and power generation, thereby reducing operational costs and emissions [4], [5],

[6]. With the widespread use of diesel engines and gas turbines in installations both ashore and onboard ships,

organisations recognise the potential of WHR and are investing significant efforts to develop and improve WHR

technology [3], [7]. A typical layout of an exhaust gas WHR system is shown in Figure 1.

One area of development is the quest to improve the performance and reliability of exhaust gas waste heat recoveryunits (WHRUs). WHRUs are essentially gas-to-gas/liquid heat exchangers where energy from hot exhaust gases is

transferred to a cooler working fluid through heat exchange surfaces such as tubes or plates. WHRUs are the most

crucial, but also one of the most vulnerable components in an exhaust gas WHR system. Although heat recovery

performance improves with larger temperature difference between the exhaust gases and working fluid, adverse

temperature gradients cause differential expansion within and between components. This induces thermal stresses

Erik currently heads the Mechanical section in RSN's Submarine Maintenance and Engineering Centre. He holds a Masters of

Science (Mechanical Engineering) from the Naval Postgraduate School and a Masters of Engineering (Marine Engineering) fromNewcastle University. His research interest includes waste heat recovery, heat and fluid flow as well as renewable energy.


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which results in component distortions, stress corrosion cracking, and fatigue making it one of the most common

causes of WHRUs failures [8].

Figure 1: Layout of a typical exhaust gas WHRS used for combined heat and power production [19]

In order to maximise heat recovery, exhaust gas WHRUs are usually positioned within exhaust stacks across thepath of the exhaust gas flow. This arrangement induces back pressure in exhaust system which in turn decreases the

performance of power plants upstream. Additional back pressure is induced when heat recovery enhancement

devices such as fins are added. For a typical gas turbine, every 245 Pascal increase in exhaust gas back pressure

reduces the power output and heat rate by 0.25% and 0.08% respectively [9]. Hield [10] also highlighted that

excessive exhaust back pressure in diesel engines can increases fuel consumption, induces instability, increases

wear, overheating, and thermal failures of engine components, severely affecting the reliability of WHRUs.

During the retrofit of WHRS onto existing power plants, there is a need to adopt a total system approach. In tandem

to achieving the desired heat recovery performance, system reliability, ease of access for inspection and

maintenance, impact of power plant upstream and ability for dry operations should also be considered. This is in

response to lessons learnt from previous unsuccessful attempts to implement WHR onto existing power plants.

Despite being able to fulfill the desired WHR performance, attempts to retrofit WHRS onto United States Navy

(USN)'s DD-963 destroyers, and Canadian Navy's DDH 280 destroyers in the 1970s received limited success due to

low reliability and operational problems. The program was subsequently cancelled [11], [12], [13].

In order to mitigate the issues, a simple but novel counter-flow water jacket WHRU configuration (Figure 2) was

selected and studied using analytic and CFD models to understand the effect of different design options on heat

recovery and pressure drop performance, as well as temperature profiles; a crucial factor concerning reliability in the

WHRU.

BACKGROUND

A counter flow configuration was chosen over a parallel one due to higher heat transfer effectiveness and smaller

temperature differences between the hot and cold fluid domains throughout the WHRU. As shown in Figure 2, the

water jacket surrounds the exhaust stack of the power plant. Water is introduced into the jacket via a pipe situated on

top of the WHRU, flows through the jacket in a direction opposite to the exhaust gas, and exits the jacket through

another pipe situated at the bottom. Heat energy from the exhaust gas is transferred to the water through the exhauststack. This configuration allows the dimensions of existing exhaust stacks to remain unchanged.

The water jacket WHRU has the potential to be made modular; allowing it to be scaled accordingly to the exhaust

stacks of different power plants. This configuration also allows easier access for inspection and maintenance. With

no components positioned within the exhaust stack, blockage and back pressure is avoided. This allows existing

exhaust layouts and configurations of power plants undergoing WHR retrofits to remain unchanged. This in turn

minimise redesign and modification of the engine intake and exhaust system and reduces the cost and complexity of

WHR implementation onto existing power plants.


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Figure 2: Schematic view of the counter-flow water jacket WHRU configuration

In this paper, the results of different length-to-tube diameter ratio (L*

= L/d) and three different water inlet and outlet

placements on heat recovery performance, exhaust side pressure drop, and temperature profiles, based on analytical

and Computational Fluid Dynamic (CFD) models are presented. Jacket to exhaust stack diameter ratio (d*

= D/d)

was kept constant at 1.25. Exhaust gas parameters used were based on the MEP803A Diesel Engine, commonly

used by the United States Marine Corp (USMC) as field Diesel Generators (DG); and the Rolls Royce 501K Gas

Turbine, commonly onboard USN ships as Ship Service Gas Turbine Generators (SSGTG). Steady state Reynolds

numbers of the exhaust gas flow (Reexh) from the power plants ranged from 20,000 and 400,000. In order to cover

the range of exhaust parameters and allow results to be applied to non-dimensionally similar configurations,

parameters and models were non-dimensionalised. Non-dimensional analysis using the Buckingham Pi theorem was

also conducted to consolidate and identify the parameter's relations on the impact of heat recovery performance and

exhaust back pressure.

For analytic models, heat recovery performance (q) was obtained using the effectiveness-NTU (-NTU) method.

The Gnielinski correlation, Filonerko Fanning friction factor ( ) were used to determine the required Nusselt

number, the overall heat transfer coefficient and finally the heat recovery effectiveness () required [14], [15].

min , ,( )exh in water i nq C T T (1)

Non-dimensional heat recovery ( ) for both analytical and CFD models were obtained by expressing heat recovery

as a fraction of the maximum recoverable heat energy. This provides a good indication of heat recovery performance

based on the operating environment. The equation for is shown in equation (2). Non-dimensional analysis also

revealed that heat recovery is impacted by WHRU geometry; through WHRU Length to stack diameter ratio (L/d)

and water jacket to stack diameter ratio (D/d), flow regime; through exhaust gas Reynolds number (Reexh) and fluid

properties; through Prandlt number (Prexh).

, , ,

1* ( , , , Pr )

( ) Reex h

exh p exh exh in wa ter in exh

q L Dq

m C T T d d

(2)

Equation (3) is used to estimate the back pressure induced by the analytic WHRU models. The surfaces of the

WHRU were assumed to be smooth in this study. Additional back pressure caused by surface roughness must beconsidered in actual designs.

22

m

eff

fL UP

D

(3)

The non-dimensional pressure drop ( *P ) is obtained using the equation (4) and is characterised by the WHRU

geometry represented by L/d and D/d, and flow regime represented by exhaust gas Reynolds number (Reexh).*

P


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is also the reciprocal the Euler number which characterizes the ratio of local pressure drop to dynamic pressure due

to fluid friction in conduits [16]. An Euler number of 1 corresponds to a perfect frictionless flow.

*

2

1( , , )

Reexh exh

P D LP

U d d

(4)

The temperature (T) generated by CFD models were non-dimensioned as a ratio to the temperature difference

between the inlet exhaust gas (Texh,in) and water (Twater,in) as shown in equation (5).

,*

, ,

wa te r i n

exh in water i n

T TT

T T

(5)

Analytical models were able to generate heat recovery and pressure drop results over a wide range of L/d, D/d ratios,

and exhaust gas Reynolds numbers. These results provided an understanding of how WHRU performance (heat

recovery and exhaust side pressure drop) are affected by changes in WHRU geometry and exhaust Reynolds

number. However, the analytical models were unable to provide resolution of the temperature profiles within the

WHRU. The analytical model was also unable to account for the effects of more complex design options such as

different water inlet and outlet placements. These limitations were resolved through the use of CFD modeling.

Figure 3: Isometric view showing locations of temperature measurements

The ANSYS 15 CFX CFD package was used to model a water jacket WHRU with a D/d of 1.25 and non-

dimensional tube thickness (t* = t/d) of 0.0625 to investigate the effects of L/d and other WHRU features on heat

recovery, pressure drop performance, and temperature profiles within the WHRU. The exhaust Reynolds numberranged from 20,000 to 400,000 while the Reynolds number of the water is held constant at 8,300. Inlet exhaust

temperature (Texh,in) and inlet water temperature (Twater,in) were kept constant at 773K and 300K respectively. This

corresponds to non-dimensional temperatures of T* = 1 and T*= 0 respectively and represents the maximum and

minimum temperatures in this study. Temperatures were measured along the length of the WHRU tube in the CFD

models at 3, 6, 9, and 12 oclockpositions as shown in Figure 3.

The standard K-epsilon (K-) turbulence model was used to evaluate flow within the CFD models. This model is

based on the Reynolds Averaged Navier-Stokes (RANS) and is widely accepted and implemented for turbulence

modeling [17]. The K- model is known to be stable, accurate, and numerically robust, and is considered to be the


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standard model in the CFD industry. The K- model in CFX also uses the scalable wall-function approach which

allows solutions on arbitrarily fine near-wall grids to be made. This increased robustness and accuracy of the

solutions over standard wall functions. Heat transfer is modeled between the exhaust, water, and the WHRU tube

domain using the ANSYS-CFXs Thermal Energy model, which models the transport of enthalpy through the

fluid domains using conduction and convection [18].

CFD MODEL VALIDATION AND SENSITIVITY ANALYSIS

CFD models were validated by ensuring the overall mass and energy entering and leaving the controlled surfaces

(inlets and outlets) are balanced. In all the cases, energy balances better than 0.019% was achieved. The worst case

had an energy imbalance of less than 0.045%. Mass averaged values of inlet and outlet temperatures, specific heat

capacity, and mass flow rates were used to verify conservation of mass within the controlled volume. A sensitivity

analysis was conducted to find out the effect of different meshing size and solver target residual error value on the

result accuracy of energy balance and the resolution. Runs using three meshing levels (250,000, 500,000, and

1,500,000 nodes) and three levels of CFX solver target residual error values (1E-4, 1E-5, and 1E-6) were conducted

using a model with L/d =10, D/d= 1.25, t/d=0.0625 subjected to the same fluid flow conditions. No significant

improvement in accuracy for output parameters and energy balance were found when the higher residual error

values of 1E-6 were used over 1E-5. Models using a meshing size of 1,500,000 nodes provided slightly better

resolution in temperatures contours and profiles than models with 500,000 nodes. The model with 250,000 nodes

provided the least resolution in terms of contours but took the least amount of time to achieve solution convergence.The models with increasing mesh sizes took increasing amounts of time run. CFD models with a meshing size of

500,000 nodes and solver residual target error value of 1E-6 offered optimum balance of accuracy, resolution, and

processing time.

In addition, results from CFD models were compared with corresponding results from analytical models. Results

plotted in Figure 4 show heat recovery from CFD and analytical models exhibited similar trends with same orders of

magnitude. Heat recovery improves with higher L/d and lower exhaust Reynolds number. At higher Reexh of

400,000, q* from the CFD models is 6% to 26% lower than corresponding results analytical models. At lower

Reynolds numbers of 20,000, q* from the CFD models are 6% to 24% higher than in analytical models. A

difference up to 26% exists between the results from the CFD and analytical model. The difference increased with

higher L/d ratio and Reynolds numbers. However, it is not unreasonable to expect differences between the CFD and

analytical models, considering the simplifications and assumptions associated with the analytical model. The

difference between the water inlet and outlet placements of the CFD and analytical models would also contributed

toward this difference. In addition, uncertainty of approximately +10% also exists in the analytical correlations used[14] [15].

Fig 4: Non-dimensional heat recovery (q*) and Non-dimensional pressure drop (P*) of analytical and CFD models

of L/d = 5, 10 and 20 with Reexh from 20,000 to 400,000. D/d=1.25, centerline water inlet / outlet placement Type 1

with Rewater = 8,300.


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EFFECT OF WHRU LENGTH TO TUBE DIAMETER (L/d) RATIO

Results from both analytical and CFD models shown in Figure 4 shows that higher L/d increase heat recovery under

similar exhaust gas flow conditions. This was expected as more available area for heat transfer is available with

higher L/d. Increasing L/d from 10 to 20 increases heat recovery by 68% to 81% whereas a reduction of L/d to 5

decreased heat recovery by 41% to 46%. Better heat recovery performance was observed at lower exhaust gas

Reynolds numbers. The maximum possible heat recovery was achieved if the length of the WHRU was longenough, i.e., L/d=1,000. However, once the maximum heat recovery is achieved (q*=1), further increase of WHRU

length did not improve the heat recovery. On the contrary, additional pressure drop at the exhaust gas side was

incurred. This causes back pressure, negatively impacting the performance and fuel consumption of the power plant

upstream. Results in Figures 4 shows that increasing the L/d from 10 to 20 amplified P* by 96%; while a decrease

of L/d from 10 to 5 reduced P* by 45%. This result was expected as frictional loss is proportional to WHRU

length. It was also observed that P* decreases as the exhaust Reynolds number increases. For the all CFD models,

P* leveled off at Reynolds numbers of 150,000 to 200,000. This is expected for flows that are modeled over

smooth surfaces. CFD models provided valuable insights into the longitudinal and radial temperature distributions

within the counter flow water jacket WHRU. The temperature profiles and contours of the exhaust stack for

WHRUs with L/d ratios of 5, 10, and 20 at exhaust Reynolds numbers of 20,000 and 400,000 are shown in Figure 5

and 6 respectively.

L/d=5 L/d=10 L/d=20

Fig 5: Temperature profile for WHRU with L/d = 5, 10 and 20, D/d = 1.25, t* = 0.0625 at Reexh

= 20,000 and

400,000; centerline water inlet / outlet placement Type 1. Re water = 8,300.

L/d = 5 L/d = 10 L/d = 20

Figure 6: Temperature contours of exhaust stack for WHRU with L/d = 5, 10 and 20 at D/d=1.25, t*=0.0625 at Reexh= 400,000 with centerline water inlet/outlet placement Type 1. Re water = 8,300.


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Generally, higher exhaust gas Reynolds numbers increase tube temperatures and amplify temperature differences.

The sharpest temperature gradients were observed at Reexh = 400,000. Temperature profiles on Figure 5 shows that

regions of adverse temperature differences shift along the WHRU with different L/d. A WHRU with L/d = 5, the

largest difference occurs at the lower half of the WHRU where the exhaust inlet and water outlet are situated. When

L/d is increased to 10, this difference is most prominent in the middle. WHRU with L/d = 20 has the most adverse

temperature differences occur at the upper end of the exhaust stack; near the location of the exhaust outlet and the

water inlet. This insight allows potentially problematic areas to be identified for mitigating measures.

Figure 7: Water streamlines depicting uneven water flow distribution inside water jacket of WHRU of L/d =10,

D/d=1.25 with centerline water inlet/outlet placement Type 1 at Reexh = 400,000, Rewater = 8,300.

Another key finding was the uneven flow distribution within the water jacket. This was revealed by examining the

streamlines within the water jacket region (Figure 7). Most of the water entering the water jacket was channeled

towards the 12 oclockprofile, as opposed to the 6 oclock profiles (where the water inlet and outlet were located).Flow distribution along the 3 oclock and 9 oclock profiles were balanced. As such, uneven heat transfer resulted in

adverse temperature profilesbetween the 6 oclock and 12 oclock profiles . This finding was evident in all WHRU

L/d ratios; as shown in Figure 5 and 6.

This difference was particularly pronounced at Reexh= 400,000. The 6 oclock and 12 oclock temperature profiles

WHRU of for L/d = 5, 10, and 20 at Reexh of 400,000 are plotted in Figure 8 for better comparison. Adverse

temperature profiles causes localized thermal stresses due to differential expansions in the tube. Differential stresses

are produced both longitudinally as well as radially along the tube and severely impact the reliability of WHRUs.

During transient loadings such as startups, shutdowns, or load changes, these stresses could potentially be amplified

due to transient and unsteady temperature differences.


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Figure 8: Comparison of temperature profile at 6 oclock and 12 oclock at Re exh = 400,000, Rewater = 8,300.for

WHRU of L/d = 5, 10, and 20

EFFECT OF WATER INLET AND OUTLET PLACEMENT

In order to mitigate the adverse temperature profiles within the WHRU and improve reliability, WHRU models with

three different types of water inlet and outlet placements as shown in Figure 9 were investigated. Type 1 and 2 are

commonly found, unlike Type 3; where the water inlet and outlet are shifted from the centre line in attempt to

improve the flow distribution within the water jacket. Results in Figure 10 shows that there are no significant

difference in heat recovery performance for WHRUs with water placement Type 1 and 2. However, heat recovery

improved by up to 19% when a WHRU equipped with placement Type 3 was used. Results in Figure 10 also

showed that the type of water placement did not have any significant impact on the exhaust side pressure drop. A

WHRU with placement Type 3 produced exhaust side back pressure 1% to 3% lower compared to WHRU with

placement Type 1 and 2.

Placement Type 1 Placement Type 2 Placement Type 3Centerline water inlet and outlet

on 6 o'clock side

Centerline water inlet at 6 o'clock

side and outlet at 12 o'clock side.

Laterally shifted water inlet and outlet

at 6 o'clock side.

Figure 9: Three types of water inlet and outlet placements for counter flow water jacket WHRU of L/d=10,

D/d=1.25 t*=0.0625.


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Figure 10: Comparison of non-dimensional heat recovery and (q*) and dimensional pressure drop (P*) forwater

inlet / outlet placement Types 1, 2, and 3 for WHRU of L/d=10, D/d=1.25, t* = 0.0625. Reexh from 20,000 to

400,000, Rewater = 8,300.

Results in Figure 11 show large temperatures differences between the 6 o'clock and 12 o'clock profiles for water

Type 1 and Type 2 placements. Sharp temperature gradients were also present near the exhaust outlet and water inlet.

The placement of the inlet and outlets on opposite sides of the WHRU (Type 2 placement) resulted in temperature

profiles which were more adverse, negatively impacting the WHRUs reliability. This finding is interesting

considering that placement Type 1 and 2 are very common in the heat exchanger industry.

In contrast, placement Type 3 which had water inlet and outlet located off centre produce gradual temperature

profile radially and longitudinally. Inspection of the water streamlines in the water jacket and corresponding

temperature contours in Figure 12 revealed that placement Type 3 improved water flow distribution in the water

jacket significantly. Placement Type 3 can be used to mitigate adverse temperature profiles and the negative effects

of differential expansion and contraction in order to improve the WHRU's reliability.

Placement Type 1 Placement Type 2 Placement Type 3

Figure 11: WHRU temperature profiles of WHRU with for water inlet / outlet placement Type 1, 2 and 3 with

L/d=10, D/d=1.25, t*=0.0625. Reexh = 400,000, Rewater = 8,300.


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Placement Type 1 Placement Type 2 Placement Type 3

Figure 12: Water streamlines and exhaust stack temperature contours for WHRU with water placement type 1, 2 and

3 L/d = 10; D/d=1.25, t*=0.0625 at Reexh = 400,000; Rewater = 8,300.

CONCLUSION

The paper presents the effects of exhaust tube length-to-diameter (L/d) ratio and coolant inlet and outlet placementson temperature profile, heat recovery performance, and exhaust side pressure drop in a counter-flow water Jacket

WHRU. Non-dimensional parametric studies of the WHRU were conducted using analytical and CFD models.

Results indicate that heat recovery increases with higher L/d. However, once the maximum possible heat recovery is

achieved, further increase in L/d only increases the exhaust gas back pressure of the power plant upstream.

Analytical models were able to predict the heat recovery, back pressure performance but were unable to provide

insights on temperature profiles and local fluid behavior within the WHRU. The CFD models generated provided

insights of the flow distribution and corresponding temperature profiles allowing potentially problematic areas to be

identified for mitigating actions.

One key finding from the study is the significance of water inlet and outlet placement toward the control of the

temperature profile within the WHRU. WHRUs with water inlets and outlets located in the centerline of the WHRU

(Type 1 and Type 2) experienced poor and uneven distribution of flow in the water jacket. The finding related to

centerline water inlet and outlet placement was consistently found in WHRUs with different L/d ratios. Poor

distribution of water flow resulted in adverse temperature profiles between the 6 oclock and 12 oclock positionsand sharp temperature gradients at the upper section of the WHRU where the exhaust outlet and water inlet are

located. Adverse temperature profile are known to produce differential expansions and thermal stress, initiating

failure mechanisms such as fatigue and stress corrosion cracking that increase the probability of premature WHRU

failures. The poor distribution of water flow also reduced the heat recovery potential of the affected WHRU. This

finding is crucial as centerline water inlet and outlet placements are very prevalent in cylindrical-shaped WHRUs

and heat exchangers. The problem was subsequently resolved by shifting the water inlet and outlet off center in

placement Type 3. This change improved the flow in the water jacket and eradicated the adverse temperature

profiles. Heat recovery performance of the WHRU also increased as much as 19% with no increase in exhaust side

pressure loss.

The study provides insights into the complex relationships involved in the design of a WHRU. Designers or program

managers of waste heat recovery projects need to consider performance and reliability from a total system point of

view. The increase in heat recovery is often linked to an increased exhaust side pressure drop, which in turn impacts

the power plant upstream in a detrimental manner. In order to further optimise reliability and performance, follow-

on study of the water inlet and outlet of water jacket WHRUs could be conducted looking at the effects of shapes,

vertical or horizontal locations, and sizes of the water inlet and outlet.

ACKNOWLEDGEMENTS

The study was part of U.S. Navy WHR Capability Roadmap undertaken by the Naval Postgraduate School. The

support and guidance from Dr Sanjeev B. Sathe and Dr Knox Millsaps from the Mechanical and Aerospace

Engineering Department of Naval Postgraduate School during the course of study is gratefully acknowledged.


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Performance and Reliability of WHRUs

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