Jean- Christophe Yann…

My former students, Flavien, Karthik, Antoine and Raji, also deserve my gratitude for all work realized during their internships. Without their contribution, it would be impossible to obtain all the results of proper quality.

I would like to thank my colleagues and friends from the lab for good moments and hours of discussions about life, culture

 3.8. Sensible and latent heat transfer near meniscus region: (a) schema [3.17] and (b) temperature profile [3.18] Evaporator and condenser are usually placed on the opposite sides of PHP (adiabatic zone between them) which leads to differences in saturation conditions between two zones and pressure driving displacement of the liquid-vapor pairs along the channels as a result.

Consequently, heat transfer process, related to PHP operation, consists of both latent and sensible heat forms, which allow increasing thermal performance. Simple design without any porous structures, moving parts and sub-systems coupled with high heat transfer performance makes PHP a very promising technology for future exploitation onboard space apparatuses.

However, considering available works in literature, complex mechanisms of thermohydrodynamic processes inside PHP is not fully studied and any precise suitable design recommendations are still not available, together with the contradictory believes of  Second, the maximal rejected heat flux in regards to the surface superheat is a one order of magnitude higher for two-phase systems compared to the liquid forced convection and at least three orders of magnitude higher than gas forced convection;  Finally, recent electronic devices require thermal management systems (TMS) capable to reject heat fluxes densities higher than 100 W/cm 2 . Thus, this part will focus on two-phase cooling solutions and a main attention will be paid to the high performances they represent.

Any cooling solution of a space electronic equipment is based on the design features that determine compactness, low power consumption and weight, as well as a decrease in the cost of the product. Thus, the choice by the designer of the thermal management system type is guided by the applications they aim for. So, understanding not only the principles of functioning of various types of TMS, but also, more importantly, their functional limitations, as well as the necessary and sufficient conditions that determine the stable operation of the TMS is of major interest. There are numerous different approaches of the TMS classification mainly caused by the development of recent TMS solutions. For a general understanding of the types and location of the two-phase heat transfer systems, it is proposed to consider the classification according to the principle of their functioning, shown in Fig. 2.2. Based on this classification, latter devices will be described in this chapter.

Mechanical Pump Driving Loops

Microchannel Heat Sink

The microchannel heat sinks were initially proposed as single-phase systems by Tuckerman and Pease in 1981 for the high-performance microprocessors cooling [2.4]. The primary component of a microchannel cooler (i.e. Microchannel Heat Sink) is shown in Fig. During two last decades, two-phase microchannel systems have attracted much attention as a method for high-heat flux electronic cooling. It has been shown that flow through large surface area (related to the device volume) microchannels boosts the heat transfer coefficient in single-phase flow; additional gains in the heat transfer coefficient can be obtained by allowing the liquid to vaporize along the walls of the microchannel [2.6].

These systems also possess the advantages like compactness and minimal coolant usage that make them attractive for electronics cooling applications.

Spray and Jet-Impingement Cooling Systems

The mechanism of the Spray Cooling could be seen in Fig. 2.4. Here, fluid is introduced to a nozzle located opposite to the heated surface. With appropriate driving pressure (provided by a pump), the fluid leaving the nozzle breaks up into individual droplets that deposit on the hot surface. The fluid forms a film on the surface and heat is transferred both by sensible mode (at the moment of the droplets deposition) and latent mode (occurring during the film evaporation).

Along with the other potential benefits of two-phase cooling, spray cooling can provide a uniform heat flux and temperature across the heating surface. Usually, spray coolers also provide low flow rate requirements due to the precise fluid-delivery and heat removal mechanism. Despite all these advantages, spray cooling is considered excessively complex and unfeasible for many applications [2.7]. Jet Impingement could be divided in the three following basic forms: free jet (liquid jet in a vapor or gaseous ambient), submerged jet (liquid jet in a liquid ambient), and confined jet (liquid jet confined between the nozzle and heat source).

A free jet-impingement cooling is usually used for two-phase applications. The primary components of this approach can be seen in Fig. 2.5. A continuous jet of the fluid is forced through a jet nozzle that impinges on the heated surface. With the applied heat, film evaporation occurs on the surface removing heat from the surface. Previous research works have shown that heat transfer coefficient is significantly higher near the jet flow than around it [2.10]. Thus, the concentration of the heat transfer coefficient and associated heat removal within the impingement zone could provoke significant temperature gradients within the device. This non-uniformity of temperature can cause thermal stress in electronics and their potential failure.

Passive Two-Phase Cooling Devices

The Thermosiphon

Thermosiphon is a high performance passive two-phase gravity-assisted heat transfer device (Fig. 2.6), which was firstly proposed in 1838 by Perkin [2.11]. At the beginning of the XX th century, a thermosiphon, called the Perkin's tube, was used as a heater in baking ovens. After, during Second World War, high demand in such devices as cooling system for gas turbine blades, significantly increased. In USSR thermosiphons were widely used in Long's thermal piles during construction in permafrost regions. From 1980, thermosiphons have been used in different applications, as well solar heating, construction, metallurgy, power electronic cooling etc. due to their simple structure, low cost and high performance.

Usually, the thermosiphon is made of an evacuated sealed smooth or grooved tube, closed from both sides and partially filled with a working fluid at ~20% of filling ratio (an optimal value) [2.12]. Conventional thermosiphon is composed of evaporation, adiabatic and condensation sections (Fig. 2.6). As a gravity assisted device, the condenser should be located above the evaporator to allow the return of liquid condensate from condenser to evaporator.

The heat flux from the source, applied to the evaporator (located on the bottom side of the tube), provokes boiling in the fluid volume. Generated vapor moves to the condensation zone thanks to the density and pressure difference between the evaporation and condensation zones. Condensed vapor flows back to the evaporator in the form of liquid film on the wall of the thermosiphon thanks to gravity assistance. The thermosiphon is a one of the most effective passive two-phase heat transfer devices, but heat transfer rate strongly depends on the geometrical parameters, fluid properties and filling ratio but also on the orientation. Performances of the thermosiphon are significantly affected by the inclination angleoperation failure could be reached in the horizontal position or in position with the evaporator located above the condenser.

Sometimes, high applied heat fluxes lead to reach the boiling crisis and dry-spots generation on the evaporator wall resulting in heat transfer performances degradation. To prevent this phenomenon, sintered wick deposition on the evaporator walls could be used to homogenize the fluid distribution (Fig. 2.6b) [2.13].

Another way to improve performances of such two-phase heat transfer devices consists in the use of the gravity assisted natural circulation two-phase loop, also called Closed Loop Thermosiphona particular version of the thermosiphon, described above. As for conventional thermosiphon, this system consists of the condenser located above the evaporator and connected via two tubes, called "riser" and "down-comer". Generally speaking, the terms "riser" and "down-comer" describe the fluid flow direction through these connection tubes. As shown in Fig. 2.7, in the bottom side of the closed loop thermosiphon, boiling in the fluid volume occurs and vapor flows through the left tube ("riser") to the condenser. After condensation, the working fluid flows to the evaporator through the right tube ("down-comer"). The main difference of the Closed Loop Thermosiphon from the conventional thermosiphon is the separation of the paths of the upward two-phase liquid/vapor flow and the descending condensate flow. This modification of the thermosiphon allows to significantly increase heat transfer performances of the device.

The Heat Pipe

The passive two-phase heat transfer device capable to work with different orientation (even, for some cases, opposite to gravity vector) have been developed in the middle of the XX th century. First concept of the device, later called Heat Pipe, has been proposed by Gaugler [2.15] in the context of the refrigeration machines with the evaporator located above the condenser. Almost twenty years later, in the framework of the American Space Program, Grover re-introduced the heat pipe as a device very similar to the device of Gaugler [2.16].

Typical heat pipe (Fig. 2.8a) looks almost identical to thermosiphonincluding evacuated sealed tube with the same evaporation, adiabatic and condensation zones and partially filled with working fluid. However, a main difference between these two devices is in the way the liquid returns from the condenser to the evaporatorcontrary to the thermosiphon, the heat pipe is not a gravity-assisted device and liquid returns to the evaporator thanks to a capillary structure (also called the "wick"), usually deposited on the walls along the heat pipe. The capillary structure could be in a form of the sintered powder, open grooves and mesh structure or their combinations, as presented in Fig. The heat pipe has a wide range of applications due to its gravity independence and wide range of the construction materials. Today, heat pipes are widely used in terrestrial and space applications in the thermal management systems. Compatibility with a wide range of fluids allows operation from the cryogenic (using helium as a working fluid) to the extreme temperatures (filled with liquid metals).

The Variable Conductance Heat Pipe

Some applications need to guarantee a constant temperature of the heat source whatever the heat flux value is. Obviously, evaporator temperature of the thermosiphon and heat pipe strongly depends on the applied heat flux: for this reason, a Variable Conductance Heat Pipe, presented in Fig. 2.9, has been developed.

Unlike the heat pipe, the variable conductance heat pipe uses a reservoir, filled with a noncondensable gas to modulate the device conductance. The working principle of the VCHP is very similar to the conventional heat pipe, except the varying length (and effective heat transfer surface) of the condenser.

(a) (b)

Figure 2.9. The working principle of a variable conductance heat pipe [2.19].

As seen in Fig. 2.9b, with low heat load (or without it), non-condensable gas fills a significant part of the heat pipe, preventing an optimal heat transfer rate. But increase of the temperature on the evaporator surface leads to the vapor pressure rise and moves the noncondensable gas into the reservoir, liberating a larger condenser surface area to interact with the vapor, resulting in a higher condensation rate (see Fig. 2.9a). So, heat transfer rate could be passively controlled to keep the cooling device at a relatively constant temperature in the range of ±1 o C [2.20].

Despite an advantage of the variable conductance heat pipe in the temperature control, the design and fabrication of such device is more difficult and costly comparing to the one of conventional heat pipe.

The Micro Heat Pipe

The miniaturization of electronic components and higher and higher heat generation have led to the introduction of the concept of a Micro Heat Pipes by Cotter in 1984 [2.21].

The micro heat pipe is defined as a heat pipe in which the mean curvature of the liquid vapor interface is comparable in magnitude to the reciprocal of the hydraulic radius of the total flow channel. Typically, micro heat pipes have convex but cusped cross sections to keep a capillary pumping in the corners, with hydraulic diameter in range of 10 to 500 μm. The typical micro heat pipe and its profiles are shown in Fig. 2.10.

Figure 2.10. Cross-section geometries of the micro heat pipes [2.22].

In analogy with conventional heat pipes, micro heat pipes have also evaporator, adiabatic and condensation zones. The working principle is very similar to the heat pipe with the open groovescentral part of the micro heat pipe occupied by a vapor, but corners (grooves) filled with liquid which flows from the condenser to the evaporator thanks to the capillary pressure difference.

Micro heat pipes are able to transfer radial heat flux densities up to the 90 W/cm 2 in horizontal inclination and up to 150 W/cm 2 for vertical orientation with the condenser located above the evaporator [2.23]. But, due to the variation of the liquid cross section with the decrease of the menisci radii along the capillary flow, the liquid pressure losses become very high and increase much faster than the driving capillary pressure (increasing with decreasing radii). Thus, the remove heat and the transport distance of these systems remain very low compared to the others (a few Watts and centimeters only).

The Loop Heat Pipe and Capillary Pumped Loop

Last few decades, the increase of the demands in the heat transfer devices capable to work with long distances and under different gravity conditions have been observed. Despite the capillary pumping effect inside the conventional heat pipe, the length increase leads to the augmentation of the pressure losses in the liquid flow, which provoke a significant heat transfer degradation due to lower working fluid flow rate to the evaporator. Thus, a lot of attention has been brought to the novel two-phase loops, combining the benefits of the closed looped thermosiphon and conventional heat pipe, as well high heat transfer rate, low hydraulic resistance and ability to work under different gravity conditions. As a result, two types of two-phase loops have been developed in the second part of the XX th century.

First presented two-phase loop was a Capillary Pumped Loop, developed by Stenger in 1966 for NASA on-orbit spacecraft missions [2.24]. Later, in 1972, a little different device, called Loop Heat Pipe, has been developed in the former USSR [2.25].

The main physical principle of operation is the same as that of heat pipes, but the vapor and liquid paths are here separated (Fig. Two-phase capillary pumped loops have several advantages compared to the conventional heat pipe. The porous wick provides thermal insulation and hydraulic separation (hydraulic lock) between liquid and vapor phases, preventing phase change on the vapor-liquid interface; they increase the design flexibility thanks to the separation of the evaporator and condenser; the capillary structure is present just in the evaporator which drastically reduces pressure losses and increases the capillary limit, and the separation of liquid and vapor lines prevent shearing of liquid particles by vapor flow (entrainment limit).

Despite these advantages, capillary pumping loops are complicated to be systematically implemented due to reservoir separation from the evaporator and necessarily presence of the temperature control heater.

The Pulsating Heat Pipe

The first concept of the Pulsating Heat Pipe has been proposed in 1971 by Smirnov [2.27]. The proposed pulsating heat pipe, described in Fig. 2.12, consists of two reservoirs (evaporator and condenser), connected by a capillary tube to ensure vapor liquid separation in the slug form. The evaporator and connecting tube are initially filled with working fluid, the condenser is partially filled with non-condensable gas and liquid, as shown in Fig. 2.12 (stage 1). Evaporation, caused by the applied heat load, leads to the pressure gradient generation and following liquid pushed to the condenser by piston effect, provoking compression of the passive gas (Fig. 2.12, stage 2). Then liquid evaporates and connection tube is filled only with vapor. Simultaneously, when vapor gets in the condenser, it condensates on the internal surface of this zone, provoking a local decrease of pressure. At the same moment, depending on the pressure distribution in the various sections of the device, the potential energy, stored in the volume of passive gas compressed in the condenser, will push the liquid back to the evaporatorreturning to the initial conditions.

Figure 2.12. Single-channel pulsating heat pipe [2.27].

Despite the singe-channel pulsating heat pipe introduction in 1971, this concept was not used until the end of the XX th century. In 1990, Akachi [2.28] introduced a new concept of pulsating heat pipes. Relatively new, this type of heat pipe represents a unique serpentine capillary tube bended in few/many turns from hot to cold sources, and partially filled with a working fluid at liquid/vapor saturation state (Fig. 2.13). The fluid at saturation state is alternately distributed in the form of liquid plugs and vapor bubbles inside the tube, due to surface tension forces in the tube of capillary dimension.

The physical principle of the PHP is based on the phase change induced motions of working fluid from evaporator to condenser. With heat power applied, the volume of the vapor bubbles increases in the evaporation zone, generating outward forces. The opposite phenomena occurs in the condensation zonevapor bubble volume decreases with heat rejection, generating inward forces. These two forces induce the motion of the all fluid from the evaporator to the condenser. Due to the exploitation of both sensible and latent heat transfer modes, PHP shows a high heat transfer capability, compared to classical heat pipes.

In addition, PHP has a simple structure and the ability to operate under different gravity levels and different positions. However, the complexity of the physical phenomena inside pulsating heat pipes creates a big challenge in applications design. A detailed presentation and overview of this system will be drawn in Chapters 3 and 4 of this document.

Conclusions

A brief review of the most efficient two-phase thermal management approaches has been done in this chapter. The comparison of these system regarding heat transfer capability presented in table 2.1. Active cooling solutions (pump forced liquid circulating loops) ensure the high-performance heat rejection, uniform surface temperature and gravity independent operation. Despite these advantages, complexity of physical phenomena involved in the system and high power consumption for the fluid pumping represents a big challenge to use such systems in space applications.

Passive two-phase systems could operate stably with high level of heat rejection without supplementary energy consumption. Unfortunately, a more simple and stable devicethe thermosiphon -could not operate without gravity assistance, which makes this type of device not applicable for space application. Other passive two-phase systems are successfully used for the space missions during last few decades. However, new challenges in the TMS for space electronic cooling need the more efficient and simple devices. The pulsating heat pipe stands out thanks to its high performance, simple structure and ability to work under different gravity conditions. However, despite its advantages in terms of simplicity and heat transfer, its large-scale development is not yet very relevant, unlike other two-phase technologies such as the capillary pumped loops or the thermosiphon, more mature from a technological point of view. Certain aspects related to the operation of the pulsating heat pipes are not yet fully understood due to relatively novel technology and physical complexity. In further chapters, some experimentations will be led on given PHP in ground and space conditionsto try to better understand the physical phenomena and the induced performances of such devices.

Chapter 3 Fundamental Mechanisms of the PHP Operation

In this chapter will be given a detailed description and analysis of the thermohydrodynamic working principles of the pulsating heat pipe, their types and differences. Due to the particular shape of the channels in the flat plate pulsating heat pipes studied in this thesis, peculiarities of the flow in rectangular capillary channels will be also discussed.

Pulsating heat pipe working principles

Beyond the different geometries and dimensions, according to the arrangement of the tube, pulsating heat pipes could be first categorized into the two principal configurations, according to [3.1], presented in Fig. However, up to now, several other characteristics can be used to categorize the PHPs: tubular or flat plate design, 2D or 3D structure, etc. The principal difference between tubular and flat plate pulsating heat pipes requires to be specified due to the direct influence on the device operation. If tubular PHP can be presented as a tube bended into a serpentine (Fig. The main parameter that distinguish flat plate device from tubular one is the geometrical continuity between channels leading to very low transverse thermal resistance: thermal spreading occurs and tends to strongly decrease the thermal gradients between channels.

This causes homogenization of the pressure differences in the channels, which are the main drivers of oscillations under slug flow regime, particularly in horizontal inclination [3.3].

Even with these construction differences, all PHP types have common characteristics, in their general structure and, more important, in their operation. These basic characteristics can be summarized by:  A unique channel bent in a serpentine with at least a few "U-turns". This channel could be with connected ends (CLPHP) or not (OLPHP). The channel surface remains smooth no additional treatment is necessary;

 Channel is partially filled with working fluid, which is compatible with pulsating heat pipe material and corresponds to the necessary thermophysics demands;

 Presence of at least one evaporation zone (hot source -heat absorption) and one condensation zone (cold source -heat rejection);

 Acceptable presence of the adiabatic zone between evaporator and condenser, but not required.

In addition to previously mentioned criteria for PHP performance, the correct operation of the device is insured by the initial liquid distribution inside the PHP in the form of the liquid plugs and vapor slugs chain. This distribution is insured by the capillary forces domination on gravitational forces and leads to the complete transverse tube blocking by the liquid plug which separates neighboring vapor slugs. The meniscus region exists on both ends of each liquid plug, caused by the interaction forces at the solid/liquid/vapor interfaces.

Depending on the surface properties and vapor slug length, the meniscus can be surrounded by thin liquid film. In case of isothermal conditionsno heat load and no heat rejection - But device performance and operation strongly depend on the continuous nonequilibrium conditions inside the system. So, pulsating heat pipe operation can be described with the local thermodynamic states for different zones, as shown in Fig. 3.4. According to Khandekar et al. [3.6], initial temperature and vapor quality in the evaporator and condenser zones are known (or can be assumed); consequently the state at the outlets of the evaporator and condenser are known too. If "point A" can be considered as a starting at the evaporator inlet, the following explanations can be given to the involved thermodynamic transformations:

A→B: (outlet of evaporator): constant pressure heat input combined with an isentropic compression due to bubble expansion;

B→D: isenthalpic pressure decrease in the adiabatic section; D (inlet of condenser) →E (outlet of condenser): due to the complicated nature of the thermodynamic process between the condenser inlet and outlet, it can be simplified to constant pressure condensation with negative isentropic work;

E→A: an isenthalpic pressure drop in the adiabatic section completes the cycle;

Because of the numerous assumptions made in this description, thermodynamic analysis is insufficient to study real PHP system, but it helps to a better understanding of the PHP operation.

Physical phenomena in PHP.

Pulsating heat pipe is a high-performing device with a very simple design, as was mentioned in previous chapter. Despite these advantages, understanding of PHP operation is still quite complex and requires intensive investigations: many parameters (design, fluids, filling ratio, inclination etc.) influence the device operation. In general, these parameters can be classified in few groups: First of all, the main task of a heat transfer device is high efficient thermal transport and stable operation. So, from this point of view, higher thermal conductivity of the PHP material is always an advantage. On the other side, pulsating heat pipe exploits both sensible and latent heat of the working fluid, which influences are greatly higher than material heat conduction, especially in cases with long adiabatic zone. Value of effective thermal conductivity (𝜆 𝑒𝑓𝑓 = 𝑄 ̇𝐿/[𝐴 𝑡 (𝑇 𝑒𝑣 -𝑇 𝑐𝑜𝑛𝑑 )] with, respectively, 𝑄 ̇ the applied heat load, L a characteristic length between hot and cold sources, At the PHP solid transverse cross section and (𝑇 𝑒𝑣 -𝑇 𝑐𝑜𝑛𝑑 ) the temperature difference between evaporator and condenser zones) for conventional pulsating heat pipes classically goes up to 10 kW.m -1 .K -1 [3.7] and up to 20 kW.m -1 .K -1 for cryogenic PHPs [3.8].



Diameter influence on the liquid distribution inside PHP

Due to the strong dependency of the flow pattern from capillary forces, one of the main parameters influencing the pulsating heat pipe operation is the channel diameter. Indeed, surface tension forces become predominant when channel diameter becomes smaller and two-phase flow changes into a slug-plug -or Taylor-flow pattern, defined as the basis of the PHP operation.

The scientists and engineers who worked on mini-and microchannel heat transfer also If a vapor slug is introduced in a vertical channel with a small diameter (smaller than a critical value defined below thanks to Eq. (3.1)-(3.3)), the bubble will not rise up by buoyancy. This means that liquid plugs and vapor slugs can be formed without stratification. All these configurations have been classified thanks to dimensionless numbers describing the ratio between capillary and other influences forces (gravity, inertial, viscous forces). Originally, Hosoda et al. [3.11] proposed to use dimensionless Bond number (𝐵𝑜 = (𝜌 𝑙 -𝜌 𝑣 )𝑔𝐷² 𝜎 ⁄ ), representing the ratio between capillary and gravity forces, to identify the diameter range, providing the transition to slug-plug flow. A critical dimeter based on Bond number can be calculated as follows:

𝐷 cr,Bo = 2 √ 𝜎 𝑔(𝜌 l -𝜌 v ) (3.1)
It has to be noted that this criterion has been widely used in the literature for PHP design.

However, as discussed by Mameli et al. [3.12], this criterion is obviously not applicable for conditions where gravity acceleration is no more relevant or applicable in such equation, like high accelerations or microgravity conditions. From the knowledge of velocity of the liquid plugs and vapor slugs motions (around 0 -1 m.s -1 for PHPs), liquid inertia has been accepted as a predominant force between the two phases due to the higher density of the liquid phase.

From this point, a new criterion based on the Weber number, based on the ratio between inertial and capillary forces (𝑊𝑒 = 𝜌 𝑙 𝑈 𝑙 2 𝐷 𝜎 ⁄ ), has been proposed by Gu et al. [3.13] and defines a new critical diameter from the critical value known as We,cr = 4:

𝐷 cr,We = 4𝜎 𝜌 l 𝑈 l 2 (3.2)
Initially related to the microchannel two-phase flow, and based on Garimella number Ga (𝐺𝑎 = 𝑅𝑒√𝐵𝑜), a more relevant criterion has been proposed later by Harichian and Garimella [3.14] to define the diameter limitation for reduced gravity fields, from the found critical value of Ga = 160:

𝐷 cr,Ga = √ 160𝜇 l 𝜌 l 𝑈 l√ 𝜎 𝑔(𝜌 l -𝜌 v ) (3.3)
According to presented criteria, the above mentioned simplified explanation of the diameter effect can be classified quantitatively.

Fluid flow and heat transfer in PHP

Considering the fluid distribution pattern inside the PHP, the motion of the liquid plug could be described by momentum equation (Eq. (3.4)). The main forces acting on the slugplug flow system originate from the surface tension, viscosity, inertia, tangential shear stresses and the force associated with the disjoining pressure at the molecular level. So, the integrated momentum equation for a moving liquid plug of length Ls over the channel with cross-section area S is given by:

𝑑(𝑚 𝑙 𝑢 𝑙 ) 𝑑𝑡 = [(𝑃 𝑟 -𝑃 𝑎 ) -Δ𝑃 𝑐 ± 𝜌 𝑙 𝑔𝑑ℎ]𝑆 -(𝐹 𝜏 + 𝐹 𝑡𝑢𝑟𝑛𝑠 ) (3.4)
where ml represents the liquid plug mass (ml = S•Ls) and ul the liquid slug velocity. ΔPc is the capillary pressure drop arising from hysteresis between advancing (θa) and receding (θr) contact angles during the motion of the plug (ΔPc = (2σ/R)•(-cos(θa) + cos(θr)). This additional resistance is additive and gets amplified if several slugs are simultaneously located in a channel (see Fig. 3.7). The cumulative pressure gradients damp oscillations and, when contact angle hysteresis is too large, an excessive number of liquid plugs may rapidly block the flow, especially if the driving force is not sufficient [3.4]. The term ρlgdh represents gravity pressure drop, which can be positive, negative or equal to zero according to the inclination of the tube and the direction of the liquid plug motion. There are a lot of contradictions in the heat transfer mechanismes inside pulsating heat pipe: it was repeatedly established that liquid evaporation/condensation contributes mainly as a means to slug-plug flow motion and the main part of heat is transferred by the sensible heat of the liquid plugs flowing from evaporator to condenser and vice-versa. Recently, the predomination of the latent heat transfer in pulsating heat pipes has been experimentally proven (between 55% and 80% of rate compared to sensible one) (Fig. 8.a) [3.16; 3.17].

Infrared measurements of the single vapor slug flow in a capillary channel heated by Joule effect (Fig. 3.8b) showed that the external temperature of the tube wall cooled only by classical convective cooling (single liquid liquid flow) increases transiently with time (Fig.

3.8c) [3.18]; in the region named "close to the meniscus", experimental temperature slightly decreases and separates from the numerical curve (after 3.6 s); the third zone corresponds to the liquid thin film evaporation phase: a sudden drop of the wall temperature is observed, followed by a plateau. The wall temperature in this area seems to be almost constant in time due to high heat transfer rate (the temperature minimal value is very close to the vapor saturation temperature). evaporating thin film in heated copper tube and (c) infrared temperature profile [3.18].

The last zone corresponds to the sharp linear temperature increase ("dried zone") due to complete liquid film evaporation and sensible heat transfer with vapor. These results confirm the above-mentioned assertion that latent heat transfer via the liquid films accounts for a considerable portion of the total heat transfer rate a very high efficiency of pulsating heat pipes.

Peculiarities of slug-plug flow in channels with rectangular shape

Due to specific production methods, the flat plate pulsating heat pipes are most of the time manufactured with milled channel with sharp angles (rectangular, triangular, trapezoidal etc.). These sharp angles induce capillary pressure imbalance between the edges and the corners and, as follows, non-symmetrical cross-sectional fluid distribution. The capillary pressure gradient induces the liquid flow along the channel even in dry-out conditions or with non-moving bubbles (no liquid deposition on the walls). Such efficient self-induced transport mechanism of liquid by the capillary effect is naturally absent in circular tubes.

The schema of a bulk meniscus with thick capillary films in the corners and dried edges of a vertical square channel, during evaporation, is illustrated in Fig.

Influence of orientation on the two-phase flow pattern inside FPPHP

The observations of the hydraulic behavior inside flat plate pulsating heat pipes are still limited but enough for the qualitative comparison of the fluid flow regimes occuring in different device orientations.

In vertical position (bottom heated mode, BHM), different modes can be observed depending on the operating conditions. For example, with low filling ratios (below 20%), the device can operate like an interconnected looped thermosiphon [3.20]. In this mode, vapor rises through the center of the channel and the condensed liquid returns to evaporator in a form of liquid film falling along the corners. Same effect was observed also with heat load augmentation [3.5]. Such operation is a result of the channel shape (with sharp angles) and was not observed in previous studies with circular-shaped channels [3.21]. This trend was confirmed even for a filling ratios close to 50% and could be caused by capillary pressure inbalance in the channel corners [3.22, 3.23].

Sometimes, the annular flow was prevalent in the central and lateral channels and distributed along the whole length of the FPPHP, and the excess liquid was trapped in adjacent channels, without contributing in heat transfer (dotted yellow rectangles in Fig. According to [3.20], higher filling ratios lead to the transition from thermosiphon mode to pulsating slug-plug flow regime. The same effect was observed for heat load increasethe flow pattern inside the FPPHP became a mix between classical thermosiphon annular (counter-current) flow in some channels and capillary slug-plug flow in others. In this case, liquid plugs are moving from the evaporator (Fig. 3.10b), with a random and rapid velocity [3.22]. This flow pattern change leads to the thermal characteristics improvement by the sensible heat contribution in overall heat transfer.

The flow pattern in horizontal position can be only slug-plug (if the internal diameter is small enough to prevent liquid stratification) due to absence of the gravity forces influence.

Despite the reported stable and efficient FPPHP operation in horizontal position [3.3],

sometimes unstable operation and low thermal performance, caused by intermittently longterm stopover phenomena (momentary stops of flow oscillations) or even early dry-out of the evaporator, was observed during device operation, tested in horizontal orientation [3.21, 3.24, 3.25, 3.26]. The insufficient pressure perturbations from one channel to another can cause this operational stops.

Succesive photos of a FPPHP (8 turns, 2×2 mm 2 , water) tested in horizontal position at 30 W of heat load are presented in Fig. 3.11 [3.24]. Oscillations occur only for short time periods and completely stopped at steady-state. Such operation can be explained by the thermally linked neighboring channelsthermal equilibrium in the neighboring channel leads to the driving pressure diminution, unlike tubular PHP with naturally separated channels.

In dried-out configuration, long liquid slugs are accumulated in the cold edge of the channels (condenser zone), forming U-shaped liquid columns of approximately the same length, as shown in Fig.

Conclusions

The fundamental operational principles of the pulsating heat pipes, as well peculiarities of the flat plate devices and their differences, and main influencing parameters, have been discussed in this chapter. Despite the fact that physical mechanisms of conventional (tubular)

pulsating heat pipe operation are well known, particular and contradictory phenomena occur in flat plate devices due to the thermal link between channels and their specific shape (sharp angles presence leads to the liquid film distribution thanks to capillary pressure unbalance between corners and edges of the transverse cross section).

The operation of an oscillating heat pipe depends on a complex combination between different parameters, since they are all more or less coupled to each other. The current level of knowledge (especially on FPPHPs) does not yet allow to create an engineering recommendations to be used in industrial level or create an adequate modelling tool.

Chapter 4 Parameters Influencing on the PHP Thermal Performance

In previous chapter, the global mechanism of the PHP functioning was presented;

influence of some geometrical parameters (channel shape/diameter; tubular/flat plate) and device orientation on flow and heat transfer behavior inside was discussed. Besides already discussed studies in last part, the influence of the fluid thermophysical, operating conditions and geometrical parameters on the pulsating heat pipe thermal transport characteristics is required. The comparison of the parameters influence on the thermal performances (based on the operational temperatures and thermal resistances) of tubular and flat plate pulsating heat pipes will be done in following chapter.

Channel diameter and shape

As previously mentioned, the internal diameter of the PHP directly influences the device operation due to the fluid distribution along channel, relating to the interfacial interaction with the PHP channel walls and gravity viscous effect.

Orientation

Due to the PHP's basic operational principles, device orientation is one of the most important parameters influencing the performance and operation stability. It is known that gravity assists the liquid reflow to the evaporator for device positioned in vertical or favorably inclined orientation with condenser located above the evaporator. However, in opposite orientation, with evaporator located above condenser, the gravity forces are directed away from the evaporator, which means that the driving pressure forces should be high enough to ensure the liquid flow. There is particular case on which the gravity almost not influences the fluid flow: i/e in horizontal orientation. In this position gravity force is directed perpendicular to the flow.

In order to understand the orientation influence on PHP operation, many studies have been done for PHPs tested with different geometrical parameters and different fluids.

According to previous studies [4.3; 4.4; 4.5], the PHP operates with best thermal performance in vertical bottom heated mode (also called "favorable orientation"), but change of inclination angle does not influence significantly the device operation, until the horizontal position for which the heat transfer efficiency drastically decreases (Fig. 4.4).

Lengths of the PHP respective zones

There are not a lot of available literature on the influence of evaporator, condenser and adiabatic lengths on pulsating heat pipes thermal transport characteristics. However, the length of PHP does not affect just the working fluid amount, but also the fluid flow inside the tubes, and so heat transfer performances. On one hand, it is clear that the length of each section of the device (evaporation, adiabatic and condensation) affects pulsating heat pipe operation in their own way. For example, decreasing evaporator and condenser lengths leads to the local heat transfer efficiency diminution due to a lower heat transfer area. On the other hand, it is known that condensation heat transfer rate in pulsating heat pipes is lower than evaporation ones. Due to the wickless structure, PHP is not subjected to the entrainment limit, but the boiling limit might occur and manifest itself by an unacceptable overheating of the evaporator due to the lack of cooling working fluid, tending to an eventually dry-out.

Thus, heat transfer area of the condenser needs to be high as much possible [4.6]. Concerning adiabatic zone, its length mostly influences the flow behavior inside the channelthe longer the channel the higher flow resistance caused by viscous forces, which requires highest driving pressure to overcome the flow resistance. Qu et al. [4.1] have tested closed loop pulsating heat pipe to find an optimal evaporatorto-condenser length rate. Fig. 4.5a shows the evolution of the thermal resistance of a pulsating heat pipe formed of a seven centimeters long condenser (over a total length of 22 cm) and filled with water and ethanol at FR = 50%. Evaporator lengths varies from 3 cm to 7 cm by step of 2 cm. Among all of the evaporator lengths, the largest thermal resistances of the PHP appears for the lowest 3 cm length, indicating that the over-short evaporator length makes the device work at a low efficiency. Thermal resistance curves clearly show the best performances of both PHP (filled with water and ethanol) for the 7 cm-length evaporator.

Obtained results look very different from tests performed by Charoensawan and Terdtoon The question that rises from these last two results is whether the increasing performances come from the increase of the evaporator length, the condenser length or from the decreasing adiabatic length. The influence of the length of each PHP section on its operation and heat transfer performance is obvious, but almost unstudied. The absence of relevant studies on this topic creates a huge interest for future works.

Number of turns

By default, pulsating heat pipe is a closed or coupled tube with several numbers of turns.

Parallel channels quantity increase leads to the increase the PHP volume capacity and heat transfer area in evaporation and condensation zones. Moreover, thermal performances becomes almost independent from orientation with higher numbers of turns [4.10]. On the opposite, operation failed for all tested PHPs with the smallest number of turns (5 turns).

Filling ratio

Since the pulsating heat pipe operation is based on the oscillatory movement of the liquid-vapor pairs (plugs and slugs), the necessary condition is to have enough liquid inside the PHP that allows this two-phase distribution. The ratio between the volume of liquid phase, being inside the PHP, and the total inner volume of the channel is called "filling ratio"

(FR). Considering the following two extreme cases: 0% and 100% of filling ratio, it becomes obvious that, in the first case (FR = 0), if there is no fluid inside PHP, the latter will operate in dry mode which means pure conduction heat transfer; but if PHP is filled with 100% of liquid, and knowing that the liquid is incompressible, there will be no place for bubbles formation, so the PHP operates in the single-phase thermosiphon mode. Thus, the filling ratio range from 10% to 90% is considered as a true range for the PHP operating as a pulsating device [4.11].

Despite numerous studies on the filling ratio influence, it remains impossible to clearly define the optimal value because the optimal filling ratio strongly depends on operating conditions, such as orientation, working fluid properties, distance between evaporator and condenser, their respective lengths, etc. Most of the time, a FR value close to 50% has been found by researchers, but with a variation in range depending on every operating conditions.

At the same time, an optimal charge ratio exists for each PHP setup; for many typical experiments (circular cross-section in a planar array with less than 20 parallel channels), the optimum filling ratio has been found of around 40% [4.12]. FR = 10,[START_REF] Khandekar | Thermo hydrodynamics of closed loop pulsating heat pipes[END_REF]50,70,85%), and tested in vertical position (BHM). There are different lowest thermal resistances with different heat inputs and filling ratios. For the PHP charged with ethanol, the lowest thermal resistance at low heat load has been found for FR = 50% and the highest for FR = 85%; but for higher heat loads, PHP charged at FR = 70% has shown the best thermal performances. Similarly to the case with ethanol, PHP charged with acetone operates better at lower heat loads with filling ratio of 35% and presents worst thermal performances for FR = 85%. But at higher heat loads, the PHP has shown better performances at FR = 70%. Yang et al. [4.13] evaluated filling ratio influence on the thermal performance of two PHPs (1 mm, and 2 mm, internal diameters) operated in vertical (BHM and THM), and horizontal orientations and filled with ethanol as working fluid (see Fig. 4.9a). For horizontal and vertical THM orientations, the optimum filling ratio was of about 50% -65%. In vertical BHM position, the optimum filling ratio was of about 40-70%, and 15 % was found for lower heat load, corresponding to a minimum thermal resistance. This is attributed to the gravity assisted thermosiphon mode of operation, i.e. the device does not act as a PHP in this situation, but as an interconnected array of closed two phase thermosiphons. However, with heat load increase, filling ratio has to be increased and the device starts to perform in selfexcited thermally driven oscillating/pulsating mode. In this case, the maximum thermal performance is not very sensitive to FR in the range between 40% and 70%. Finally, Kamijima et al. [4.15] evaluated the filling ratio influence on micro-FPPHP with channel internal diameter of 0.35 mm and filled with Fluorinert FC-72. For low heat input (up to 10 W), the thermal conductivity of micro-PHP increases with the filling ratio increase (Fig. 4.9b) but, the effective thermal conductivities for FR = 39%, 44%, and 63% decrease when heat load exceeds 10 W, indicating the occurrence of the dry-out. The micro-FPPHP filled at FR = 48% and 54% can operate till almost 25 W with continuous effective thermal conductivity increase. The optimal FR in this case tends to increase the heat power range of operation.

However, filling ratio is also related to the liquid plugs and vapor plugs proportions: on one hand, with low filling ratios, the liquid plugs are shorter and wall friction force is smaller leading to easier oscillations. But continuous channel wall rewetting in the evaporator becomes more difficult and dry-out appears more frequently. On the other hand, high filling ratio leads to higher friction forces, that means the necessity in higher pressure difference between neighboring vapor slugs to drive the flow [4.16]. It can be concluded that the optimal filling ratio is the result between liquid and vapor plugs distribution and respective lengths in the device, depending on operating conditions.

Working fluid

First of all, the working fluid should correspond to all design, project and safety demands, i.e. Assuming that a thin liquid film surrounds each vapor bubble, vapor is in saturation state, and pressure can be expressed using the Clapeyron equation and the ideal gas law corrected with compressibility factor Z:

𝑃 𝑣 = ( 𝜕𝑃 𝜕𝑇 ) 𝑠𝑎𝑡 𝑍𝑅𝑇 𝑠𝑎𝑡 2 ℎ 𝑙𝑣 (4.1)
Previously, it was noted that pressure difference between two vapor slugs surrounded a liquid plug (Pr -Pa) is a main driver for oscillation motion in PHP. As the experimental timeaverage temperature difference between two vapor slugs is less than 2 °C [4.19], the terms of Eq. (4.1) can be assumed to be equal to the values corresponding to the average saturation temperature of the two vapor plugs and can be evaluated at Tsat = (Tsat,r + Tsat,a)/2. Then, the driving pressure for the fluid motion (from Eq. (3.4)) can be simplified as follows:

∆𝑃 𝑣 = (𝑃 𝑟 -𝑃 𝑎 ) ≈ 2 ( 𝜕𝑃 𝜕𝑇 ) 𝑠𝑎𝑡 𝑍𝑅𝑇 𝑠𝑎𝑡 ℎ 𝑙𝑣 (𝑇 𝑠𝑎𝑡,𝑟 -𝑇 𝑠𝑎𝑡,𝑎 ) (4.2)
Assuming that slug-plug motion in PHP is a fully developed laminar flow, the viscous friction of vapor is negligible and fluid velocity is related to the sensible heat transfer rate due to the motion of liquid slug (ul ≈ Qsens/ρlcplS (Tev -Tcond) where Qsens=γsensQ is experimentally obtained with sensible heat transfer rate, related to total heat transfer), the frictional pressure drop is determined as follows:

∆𝑃 𝜏 = 32𝜇 𝑙 ( 𝐿 𝑠 𝐷 2 ) 𝑢 𝑙 ( 𝛾 𝑠𝑒𝑛𝑠 𝑄 𝜌 𝑙 𝑐 𝑝𝑙 𝑆(𝑇 𝑒𝑣 -𝑇 𝑐𝑜𝑛𝑑 ) ) (4.3)
Since the figure of merit is proportional to the ratio of driving pressure ΔPv on the frictional pressure drop ΔPτ (see Eq. (4.5)), the latter has to be calculated:: However, if analysis provided by authors is suitable for their PHP under tested parameters and has a good agreement with experimental results, the best thermal performance for larger pulsating heat pipes was obtained with water as a working fluid.

∆𝑃 𝑣 ∆𝑃 𝜏 = ( 𝜕𝑃 𝜕𝑇
Therefore, it can be concluded that the figure of merit can finally be proposed as an initial qualitative guideline for working fluid selection to be used in micro PHP. Anyway, validation of proposed approach needs much more approbations with different geometries.

Binary mixtures

Passive heat transfer enhancement, based on fluid modification, is one of the more practical techniques thanks to its easily implementation and low cost. As mentioned above, the PHPs operation behavior and performance are greatly influenced by the working fluid properties. Some of them directly affect both flow behavior and thermal performances. the positive influence of some mixtures on their PHP start-up and dry-out delay. However, authors also concluded on the dominating importance of the filling ratio, mixing ratio and binary mixture type on the PHP operation. A more detailed information can be found in the review paper of Ayel et al. [4.24].

Results for a PHP with channel diameter of 2 mm tested with water, methanol, acetone and their mixtures, at FR = 50%, obtained by Pachghare and Mahalle [4.25], are presented in Fig. 4.12. It can be seen that the PHP filled with water-methanol mixture (Fig. 4.12a)

gives better thermal performance than both pure working fluids, especially water. It is also clear that, during starting at low heat loads (from 10 W to 30 W), the thermal resistance of the pure working fluid drastically varies and, with heat load augmentation, the thermal resistance curves converge for all fluids. However, after 80 W of heat load, water-methanol binary mixture filled PHP dries out.

In the case of water-acetone mixture (Fig. 4.12b), the PHP presents lower thermal resistance than with pure water, but higher than with pure acetone. Heat load augmentation does not provoke the dry-out at high heat fluxes. This configuration does not seem interesting in that it only leads to intermediate thermal performances between both pure fluids.

Self-Rewetting Fluids

Heat transfer degradation of PHPs often happens due to partial or full dry-out of the evaporator zone, particularly when using water as working fluid: while very performing through its thermophysical properties, its bad wettability, compared to other fluids, generally plays a negative role on heat and mass transfer in such systems. In order to overcome this problem, binary aqueous mixtures, based on the addition of a low quantity of high-weight alcohols, also called Self-Rewetting Fluid (SRWF), could improve wettability and rewet the heat transfer surface. Abe et al. [4.26] measured the surface tension of alcohols and their aqueous mixtures and concluded that SRWFs surface tension decreases with increasing concentration and become more hydrophilic. Moreover, above a certain threshold temperature, the surface tension increases when temperature rises (Fig. 4.13): it would be conducive for the colder fluid moving spontaneously toward the heated side. Therefore, as shown in Fig. 3.16a, the self-rewetting fluid can flow from colder region towards the higher temperature one on the channel inner surface thanks to the Marangoni effect.

It is known that, in confined two-phase flows, the long vapor slugs are not always surrounded by liquid films, leading to the generation of a dry patch on the heated surface.

But, with SRWF, this phenomenon can be prevented and the evaporation heat transfer can be improved in the PHP, as schematically shown in Fig. 4.14b. This phenomenon causes a rapid increase in vapor pressure in the heating section due to high concentration of evaporated heavy alcohol within. On the other hand, under microgravity, the PHP filled with SRWF operated much better than with pure water: the maximum evaporator temperature rise during transient microgravity phases for SRWF-filled PHP was of around 11 K at 200 W of applied power, whereas but this temperature difference was of around 30 K for water. This happened due to stopover phenomena for water-filled PHP, during which all liquid phase was accumulated in the condenser section and the absence of gravity forces did not push the liquid phase towards the evaporator. Almost the same trends in thermal performance were obtained in [4.31]: in vertical BHM mode, PHP filled with water operated better than with butan-2-olwater solution, especially for low heat loads. With heat load augmentation, the thermal resistances became almost equal. But, in horizontal orientation, the PHP filled with water and tested at 20 o C of condenser temperature did not operate at heat loads higher than 100 W, and SRWF-filled device operated in all heat loads range (from 50 W to 200 W).

Moreover, the thermal resistances for pure water were at least three times higher than for SRWF.

Based on this literature survey, an attempt will be made to clarify some questions, related to the different parameter influence on the flat plate pulsating heat pipe, especially in the microgravity conditions.

Surfactants

The positive effect of the fluid modification has been described above, but zeotropic binary mixtures are characterized by significant changes in boiling temperatures and transport properties. So, it is almost impossible to selectively change specific properties like, for example, surface tension, which directly influences liquid/surface wettabilityvery important feature influencing boiling and evaporation heat transfer.

In pharmaceutic and cosmetic industries, particular substances, called "surfactants", are widely used to stabilize creams, foams etc. and prevent fluids segregation by the interfacial tension reduction on the interface between two immiscible liquids.

Borrowing this approach, a very little rate of surfactant added to a working fluid can significantly decrease the total surface tension (Fig. 4.15a) and improve the channel walls wetting of the working fluid while keeping other thermophysical properties constant, as well density, saturation temperature, viscosity (Fig. 4.15b), latent heat of vaporization etc.

In literature, using of the surfactant aqueous solutions as working fluids generally lead to better PHP operation and also prevent from dry spots formation [4.32]. Wang et al. [4.33] used sodium stearate surfactant solutions with concentrations of 10, 20 and 40 ppm, and found that the thermal resistance values of the PHP with 40 ppm of surfactant has significantly been decreased compared to pure water. Bastakoti et al. [4.34] tested the PHP filled with cetyltrimethyl ammonium chloride (CTAC) surfactant solution and found a lower thermal resistance compared to the one obtained with pure water. Bao et al. [4.35] observed that the presence of surfactant had a positive effect on lower start-up, higher dry-out and heat fluxes compared to pure water. Should be noted that all these tests were done for the PHP operated in BHM. However, testing their device in horizontal orientation, Ayel et al. [4.38] found unstable PHP operation with water/Tween 20 and water/Tween 40 solutions: the device operated quite better than one filled with water, but often sudden and great evaporator temperature rises happened (up to 50 K higher than average operational temperature). The foam generation and temporary channel blockage can also explain this phenomenon.

Microgravity experiments on pulsating heat pipes

The gravity directly influences the two-phase fluid dynamics (during pool boiling, flow boiling and boiling in confined space) and, as a consequence, on the device thermal performance and, finally, makes existed prediction tools unusable for space applications Note that, on Fig. 4.21b, the FPPHP tested in horizontal inclination doesn't seem to be influenced by changes of gravity levels, proving that such a system can be an efficient candidate for thermal control in different space applications.

The influence of the channel diameter on the thermal behavior and fluid flow inside flat plate pulsating heat pipes have been studied for four semi-transparent devices by Ayel et al. [4.47;4.48;4.49] during ESA 64 th PFC. Each tested FPPHP was made of copper plate with engraved square channels with diameters range from 1.5 to 3 mm, and 11 U-turns in the evaporator zone. A borosilicate glass was glued on the grooved side of the copper plate to allow visualizations. The refrigerant FC-72 was used as working fluid (FR = 50%). Initially, FPPHPs were installed in vertical position (BHM). During normal gravity periods, the FPPHP with highest channel diameter showed the best thermal performance. Microgravity onset led to the flow regime change into slug-plug for all channel diameters. Dry-out periods were observed with following overall fluid motion reactivations phases. PHPs with 2.5 mm and 3 mm channel diameters showed almost the same thermal performance, but PHPs with lower diameters operated with less efficiency.

Recently, Ando et al. [4.50;4.51] performed on-orbit tests of a flat plate pulsating heat pipe with check valves and filled with HFC-134a at filling ratio of 45%. They found successful operation of their PHP with heat transfer performance matching to preliminary ground tests. No failure or performance degradation were observed for almost four years of operation. However, some start-up difficulties were seen under certain initial conditions.

Further analysis indicated that there is a risk of start-up with most check valves positioned close to the evaporation zone. However, if all check valves are located near the condenser section, liquid slugs could be supplied to the evaporator section even if the liquid was localized in the condenser section, and the start-up is successful. Later, Taft and Irick [4.52] showed stable operation of their PHP filled with butane during six weeks of space flight. Nevertheless, the influence of many other parameters is still not clear or sometimes contradictory. Numerous studies on space applications of pulsating heat pipes have been done, but still with no any precise design recommendation.

Conclusions

Review

Based on this literature survey, an attempt will be made to clarify some questions, related to the different parameters influence on the flat plate pulsating heat pipe, especially operating under microgravity conditions.

Chapter 5 Experimental Set-up for Ground and Parabolic Flight Tests

As mentioned in the previous chapter, gravity significantly affects the two-phase flow inside pulsating heat pipes, especially during conditions with gravity field absence.

Furthermore, gravity absence makes difficult (and sometimes impossible) to use grounddeveloped models and correlations to precisely predict the flow regime inside the channels and, as a consequence, heat transfer performances. Therefore, there is a strong demand on experimental validation of developed models in microgravity conditions. So, for these reasons an experimental rack has been developed to allow flat plate pulsating heat pipe tests under normal and microgravity condition.

In this chapter, detailed description of the experimental system and tests sections used for the flat plate pulsating heat pipes investigation during on-ground and several parabolic flight tests will be reported. Likewise, architectural differences between tested FPPHPs, the cooling systems for on ground and parabolic flight tests will be described. Then, the main measurement techniques and data treatments, as well uncertainty analysis will be presented.

The Parabolic Flight

Nowadays, several approaches are available for microgravity environment simulation

The Experimental set-up

Based on the technical requirements of the Zero-G Flights operator (Novespace) and on the scientific demands, the experimental system has been developed and built in order to respond to flat pulsating heat pipe investigations during on-ground and parabolic flight tests.

Used experimental set-up consisted of two main parts: command and test racks (Fig. 5.3).

The test rack has been designed in order to allow thermohydraulic investigations of the Since the experimental rack mostly consists in data acquisition equipment and test section, the command rack is used for test manipulations and includes all control, security and power equipment: three PCs (one for Labview® Software data acquisition management, and the two others for high speed -and IR-cameras control), two DC power sources assuring the FPPHP evaporator heating, two DC power sources for DAQ system and synchronization module, accelerometer, gear pump controller and security box.

The Test Section

The data obtained during 64 th , 69 th , 71 st mm will be analyzed in chapter 7). All four tested PHPs were made from a copper substrate with a length of 204 mm and varying plate width/thickness (corresponding to the channel diameter). A unique groove with a rectangular shape was machined in this plate forming a serpentine with 11 U-turns on one side (evaporator zone) or 22 parallel channels. Copper plate was covered by a borosilicate cover plate on the grooved side, forming the entire device (Fig. 5.4). The sealing of the grooved copper plate with borosilicate plate was realized thanks to glue deposition between them (with thickness of 0.5 mm). An elastic silicon glue (NUSIL® CV7 2289 1P) was used to overcome the possible deformation caused by the thermal expansion differences between copper and borosilicate (3.3×10 -6 K -1 for borosilicate and 17×10 -6 K -1 for copper). Unfortunately, this kind of glue is porous and air permeation inside the PHP from ambient is possible, due to very low internal pressure, compared to atmospheric one. Since the presence of non-condensable gases strongly influences the PHP operation, a containment channel connected to a high-volume vacuumed reservoir was provided ("Extension vacuum reservoir", Fig. 5.4). This kind of test cell modification keeps the functioning area of the device "protected" from ambient air permeation. A heating wire with diameter of 1 mm was embedded in a copper plate (height of 10 mm and width equal to evaporation zone of each PHP) which was brazed on the back side of the PHP to provide heat supply to evaporator.

A second massive copper plate (10 mm thickness, 100 mm height and width equal to the device one) with machined serpentine channel through which the cooling liquid circulates, was soldered on the back top-side of the PHP and used as a cold source for heat rejection (condenser).

69 th ESA Parabolic Flight Campaign

Later Based on the previous experience, and in the purpose of having a thermal expansion coefficient as close as possible to borosilicateor sapphireones, together with a high value of thermal conductivity, molybdenum was chosen as material for pulsating heat pipe metallic plate and sapphire glass as a cover plate (thermal expansion coefficients: 3.2×10 -6

K -1 for sapphire and 5×10 -6 K -1 for molybdenum; and thermal conductivity of 138 W.m -1 K - 1 for molybdenum). In analogy with copper-based flat plate pulsating heat pipes, tested during 64 th ESA PFC, a unique serpentine groove of rectangular shape (depth of 2.5 mm and width of 3 mm) was milled in the molybdenum plate (200×80×3 mm 3 ). Regarding the changes in width of the device (restricted to 80 mm), the number of U-turns in the evaporator was also reduced down to seven (or 14 parallel channels).

Grooved molybdenum plate was covered with a sapphire plate (80×200×5 mm 3 ) using rigid epoxy glue to guarantee perfect sealing at the plate boards, and between adjacent channels relative to one-another. The thickness of the glue layer was set at 0.5 mm (same than for the previous case), providing a depth of the channel of 3 mm, giving them square shape which corresponds to the chosen hydraulic diameter of 3 mm. Epoxy glue could be used here thanks to the same levels of thermal expansion of both materials: indeed, although rigid, compared to elastic -silicon-typeglues, this type of glue turns out to be perfectly leakproof, particularly to air permeation. Photograph and scheme of the developed FPPHP can be seen in Fig. 5.5. Temperature measurements in the evaporator zone were provided by T-type thermocouples installed in grooves (milled on the back side of the FPPHPs, on the ribs between adjustment channels). Pressure values inside the FPPHP were acquired thanks to the pressure sensor connected to the port at bottom rib of the device. These means of the data extracting are the same for all tested PHPs.

The filling Procedure

After test cell manufacturing, the system needs to be partially filled with liquid with the required quantity. The latter is determined by the filling ratio (FR), corresponding to the volume of fluid injected into the device compared to the total internal volume of the channel (see chapter 4). Previous literature survey and past laboratory work have shown that the optimal value of the PHP filling ratio has been generally found close to 50%. On this basis, a first step consists in preparing the filling reservoir with the required quantity of the fluid.

As all procedures have been performed at room temperature (Troom ≈ 20 o C), the required mass of fluid can be calculated as follows:

𝑚 𝑙,𝑓𝑖𝑙𝑙 = 𝜌 𝑙 (𝑇 𝑟𝑜𝑜𝑚 )𝑉 𝑡𝑜𝑡 × 𝐹𝑅 100

(5.1)

where Vtot, and l, are respectively the internal volume of the PHP (defined during device conception) and the fluid density at Troom. After fluid mass calculation, the filling reservoir isvacuumed down to less than 0.03 mbar with a molecular pump of a Axiden® ASM 142 leaks detector. After evacuation, the filling reservoir was weighed and filled with the previously prepared liquid (with a small higher amount than calculated). For water as working fluid, the following steps are: measuring the liquid mass inside the reservoir; heating the reservoir above 125 °C to ensure continuous liquid boiling inside and to dissolve gases separation; after boiling (and external wall temperature verification to be sure that the internal pressure becomes above atmospheric one), a first degassing is done by opening the valve under an extractor hood; then, the reservoir is immersed into a cryostat bath and cooled at -20 o C; once turned into solid phase, a vacuum pump is connected to the reservoir for 30

second to evacuate all residual vapor and non-condensable gases (NCG). This set of procedures (heating-cooling-evacuating) is repeated three times before pulsating heat pipe filling. For other working fluids, the solidification process was not performed, but only cooling the reservoir before heating it again. The last action in reservoir preparation is the liquid saturation pressure verificationa pressure sensor is connected to the reservoir set into the thermostat bath, which allows to control temperature inside the reservoir. If the saturation pressure curve is equal to the theoretical one (Fig. 5.8), then it is assumed that there are no NCGs inside the reservoir or with negligible quantity.

After saturation pressure verification, the pulsating heat pipes can be filled: the FPPHP is connected to the filling sub-system, consisting in a rotary vacuum pump, a leak detector and a filling reservoir, schematically presented in Fig. ). If the value of the helium permeation in the system is lower than 10 -9 mbar.ls -1 for the glued PHP and 10 -10 mbar.ls -1 for the soldered one, the device can be filled with earlier predefined working fluid. inside them. Finally, the two valves between reservoir (V3) and the FPPHP (V0) are opened, then the fluid penetrates into the PHP thanks to gravity assistance (reservoir installed above FPPHP) and pressure difference between reservoir and FPPHP.

To fill PHP for tests with non-condensable gases presence, the rotary pump needs to be replaces by the NCG reservoir (500 cm 3 of volume) connected to the ambient through needle valve (valve V1 is closed) after residual liquid evacuation and before leaks detection. Then, vacuum pump of leaks detector used for NCG reservoir vacuuming (needle valve, valves V0, V2 are closed). After this, previously, described procedure of leaks detection can be done. If system is leaks-free, FPPHP can be filled with NCG (valves V2 and V3 are closed), the needle valve opens slightly to fill predefined quantity of gases (controlled with pressure sensor). Then pressure is reached necessary value, needle valve and valve V1 closed. Finally, previously explained last step of filling procedure can be done.

The cooling loop

Since the developed FPPHPs were tested in two operating conditions: on ground and during parabolic flights, two different cooling sub-systems were used to ensure stable and efficient heat rejection from the condenser cooling plate.

Data acquisition and control

On one hand, temperature measurements play a key role in the analysis of heat transfer devices. Due to some design differences, the number and positions of the thermal sensors varies from each tested pulsating heat pipe. Despite this, the quantity of thermocouples for each device (four in evaporation zone as minimum) is assumed to be sufficient for thermal Due to flight trajectory, the gravity levels are not permanent during whole tests, so measurements of the gravity acceleration become necessary to provide the adequate analysis of the obtained data. The measurements of the accelerations in three orthogonal directions are realized thank to a "homemade" accelerometer.

The heat load applied to the evaporators were ensured by two ELC® ALR3220 electrical power sources (DC, 0-30 V, 0-20 A; dV=0.03%, dI=0.05%). Measurement of the output electrical tension was done at both ends of the electrical heating resistant wire, whereas current was directly by the power source.

To Moreover, this program interface ensures the system security provoking electrical heater shutdown if one or several measured temperatures reach a value higher than 115 o C.

This threshold value has been chosen to prevent from thermal stresses and deterioration of all the elements constituting the experimental bench and prevent from significant pressure augmentation inside PHPs (related to the design). According to Novespace's specifications, the experimental bench was also equipped by a redundant physical security box with maximum temperature set at 125 ˚C.

Working fluid

The Additional attempt to improve heat transfer characteristics of the FPPHPs has been done, using different solutions of water and alcohol, and also with surfactant additives. The changes in the device heat transfer capability are expected mostly due to wettability improvement and rewetting phenomena (for self-rewetting fluids).

Experimental procedure

Ground tests

Before every experimentations, the PHP had to be positioned in a predefined orientation vertical or horizontal. Temperature of the secondary cooling fluid had also to be fixed thanks to the cryothermostat controller. Then, once the device temperature stabilized, a 50Wheat load level was applied. After a while (thermal regime establishment between 30 minutes and 1 hour), heat load was increased by steps of 50 W. Same procedure was repeated until 200 W. By the end of the test period with heat load of 200 W, it was then reduced down to 100 W to ensure repeatability tests.

Once the first test sequence is over, the condenser temperature or orientation could be changed for the following test.

Parabolic flight tests

During parabolic flight campaigns, FPPHP have been tested with different heat loads for each series of parabolas. Before first set of parabolas, all systems need to be verified and previously determined heat load should be applied to reach the stable PHP operation. After announcement of the pull-up phase (beginning of the parabolic maneuver), data acquisition was startedtemperature, pressure and heat load values recording, as well video from highspeed camera. Due to lack of time between parabola's sets to save recorded video on hard drive, the video registration was stopped just after the fourth parabola and saving procedure started. All other data continued to be registered until the end of the series of fifth parabolas.

Immediately after flight stabilization, the acquisition was stopped and new values of heat load were set.

Data reduction and uncertainty analysis

Usually, experimental studies are accompanied by numerous measurements of different parameters, which can be difficult in the term of the widescale representation and thermal efficiency comparison. However, thermal resistance, as an important parameter indicating the device capability to thermal transport, and can be used for the FPPHP efficiency analysis.

It is defined as follows:

𝑅 𝑡ℎ = (5.4)

Resulting uncertainty of the heat load varies in the range of 0.08% to 0.15%. The uncertainty for the thermal resistance has been found for the highest heat loads (with higher temperature differences between evaporator and cooling water) and was equal to 1.7%, whereas it reached values up to 26.7% for the lowest heat loads and temperature differences (complete range of uncertainties can be found in Appendix C).

Results presentation and test repeatability

To facilitate the understanding of the graphics on the temperature behavior presented in the next chapter, first results for the water-filled FPPHP tested on ground in vertical position (BHM) will be here described as an example. The temperature and pressure responses as The modifiable parameter during ground tests is the heat load, but for parabolic flight it is the gravity acceleration. So, results representation will be slightly different. Fig. 5.15 depicts the typical temperature, pressure and gravity profiles during one set of five consecutive parabolas. In analogy with Fig. 5.13, curves colored in shades of red, orange and grey are respectively representing evaporator, adiabatic temperatures and pressure. All parabola's set test are performed with the same heat load, so the curve of heat power is not necessary in the graphs, but the blue line, representing cooling water mean temperature, has been added due to the absence of precise cooling liquid temperature regulation (provided by thermostat during ground tests). The dashed grey curve shows the gravity variation during the five consecutive parabolas and allows to recognize all transitions caused by the gravity.

At the beginning of the parabola set, the FPPHP operates in vertical position (BHM) due to normal gravity onboard of aircraft. In Fig. 5.15, just after parabola starts, when microgravity conditions are reached, the evaporator temperatures drastically increase due to channel wall dry-out in the evaporator zone and liquid accumulation in the condenser zone (as already mentioned in section 3.2.4). This phenomena was observed for all pure fluids, tested during this work. Adiabatic temperatures smoothly increase in the microgravity phase beginningand heat is transferred mostly by conduction. Similarly to the adiabatic

Conclusions

The experimental system for ground and microgravity experiments, allowing both, qualitative and quantitative characterization of the flat plate pulsating heat pipe have been This chapter will be devised in two main parts:

 in the first one, results will be described for FPPHP tested on ground in order to evaluate the influences of orientation, operating conditions (condenser temperature, input heat load and filling ratio) and fluid thermophysical properties on both FPPHP behavior and thermal performance under normal gravity conditions;

 the second part will describe results for the FPPHP tested onboard of aircraft during parabolic flights under microgravity conditions (~0.01g) and filled with different fluids.

In addition, results from specific ESA 74th PF campaign, for Lunar (~0.18g) and Martian (~0.38g) gravity conditions, will be briefly presented. Influence of noncondensable gases on FPPHP operation during normal gravity and microgravity periods will also be discussed.

Ground tests

As mentioned above (sections 3.2.4 and 4.2), the influence of gravity on the operation of a pulsating heat pipe is a crucial step in determining prospective applications for this device, especially for aerospace applications. This issue has been intensively studied to demonstrate gravity effects on PHP performances and overall operational behavior.

Previous studies on pulsating heat pipes have shown great influence of orientation (inclination) on both thermal transfer characteristics and operation behavior [6.1; 6.2; 6.3; 6.4; 6.5]: gravity assistance during vertical operation for bottom heated mode (BHM), as well as over inclinations of PHP (respecting BHM) [6.6], leads to significant decrease of thermal resistance compared to horizontal FPPHP operationin vertical orientation, gravity vector helps the liquid plugs to flow back to the evaporator zone.

In this section, ground based experimental results for the copper flat plate pulsating heat pipe, developed and described in Chapter 5, and tested in both BHM vertical and horizontal orientations are presented. Operating conditions, used working fluids and filling ratios are shown in table 6.1. Next augmentation of heat load up to 100 W leads to low temperatures (at the level of ~28 o C) and higher pressure amplitude oscillations during total period before next heat load step. Absence of significant temperature oscillations is due to gravitational forces dominance on surface tension forces and ensuing FPPHP operational changesfrom slug/plug pulsating flow to annular, as previously observed and reported by [6.2]. FPPHP operational behavior for 150 W and 200 W of heat load is very similar to the case at 100 W of applied power with average evaporator temperature augmentations till ~33 o C and ~37 o C, respectively. One can argue that, at these operating conditions, the FPPHP operates very efficiently. For methanol-filled FPPHP (FR=50%), the temperature and pressure responses are presented in Fig. 6.3. Temperature behavior seems to be similar to the case of water-filled FPPHP for 50 W of applied power but with higher evaporator temperature increasement From Fig. 6.4, it seems that thermal resistance for methanol-filled device at 50 W of heat load and Tcool = 20 o C is two times higher than for water-filled and three times higher than for ethanol-filled FPPHP, which seems to be in good accordance with evaporator superheats observed for this initial heat load in Fig. Empty FPPHP

Influence of the filling ratio

Some authors [6.7; 6.8] previously reported that filling ratio has significant influence on both pulsating heat pipe thermal performance and operation behavior. Usually, during the tests, flat plate pulsating heat pipe is filled with a filling ratio of FR=50%. Knowing that tested FPPHP is not symmetric (condenser is two time larger than evaporator) and that, during dry-out periods, liquid collects in the condensation zone, studied FPPHP with filling ratio of FR=59% allows to reach halfway between the condenser and the evaporator (Fig. 6.5), have been carried out. filled with water at filling ratios of 50%, and 59%, respectively. First of all, in analogy with results for filling ratio of 50%, thermal resistances decrease with heat load augmentation for both cooling temperatures at filling ratio of 59%.

Thermal resistance of FPPHP filled with water for FR=50% and tested at condenser temperature of 20 o C (solid blue line, Fig. 6.6) and heat load of 50 W is almost two times higher than for the case at FR=59% (dashed blue line, Fig,6.6). Following increase of heat load leads to the values alignment until almost fully coincidence of thermal resistance values at 150 W and 200 W of heat loads.

Condenser temperature rise shows opposite effect (for heat load equal to 50 W)thermal resistance for higher filling ratio (dashed red line, Fig. 6.6) at 50 W of applied power is almost four time higher than value for FPPHP filled with lower amount of fluid. However, with heat load augmentation (in the range of 100 W to 200 W), any influence of filling ratio on FPPHP thermal performances has been noticed. These disagreements with thermal resistances at 50 W of heat load could be provoked by FPPHP unstable operation, but with heat load augmentation thermal responses seem to become more stable and reproducible.

Influence of the binary mixtures

Nowadays, many studies on enhancement of heat transfer performances of the pulsating heat pipes are carried out. Based on PHP working principles concerning liquid plug motions, wettability and surface forces can significantly affect the two-phase transfer characteristics Due to these facts, improvement of wetting properties relating to pulsating heat pipes becomes an interesting approach. Nowadays, many techniques are used to change wetting properties as well both solid surfaces and/or fluids modifications. Indeed, using fluids mixtures in the context of wetting and interfacial properties seems to be an interesting approach. This subsection will briefly present results for FPPHP tested with alcohol aqueous solutions, self-rewetting fluids and water-surfactant solutions. Thermal behaviors will be analyzed and compared with the ones obtained with pure water.

Fluids modification leads to changes of many thermo-physical parameters (thermal and transport properties) so, in this study, fluids with relatively close values of surface tension have been selected. Table 6.2 presents main thermophysical properties of tested fluids (density, viscosity and thermal conductivity of water -0.5% Tween 40 © solution assumed equal to pure water due to small amount of surfactant). Comparison of thermal performances for the FPPHPs filled with water and water surfactant solution (0.5% Tween 40 © ) is presented in Fig. 6.9. Thermal resistances for tests conducted at Tcool = 20 o C of cooling water showed a lower thermal performance of water/surfactant solution than for pure water at the initial level of heat load (50 W).

Following overall thermal performance changes accompanied with heat load augmentation have been observed.

Test case with Tcool = 40 o C showed the same: higher thermal resistance value for waterfilled FPPHP at initial heat load of 50 W. This can be explained by the temperature augmentation during non-oscillating device operation, which could be seen also for watersurfactant solution, but with much less temperature augmentation (temperature peak in Fig.

A.1.13). Temperature high frequent and low amplitude oscillations take place with applied heat power increase (almost similar for two cases) which leads to the alignment of thermal resistance values.

Influence of the non-condensable gas presence

Generation of non-condensable gas inside two-phase heat transfer devices often occurs in a unpredictable way; it is mainly due to chemical reactions in liquid-solid interface, or to some some leaks through the soldered connections, etc. Non-condensable gas (NCG) presence inside two-phase heat transfer devices can become a crucial issue influencing the system operation and reliability [6.12]. NCG generation is one of the most important reasons of heat pipe heat transfer capability degradation and failure [6.13; 6.14]. However, not many papers exist concerning NCG influence on PHP operation. NCG amount directly influences the slug/plug flow dynamics and leads to a remarkable decrease of heat transfer capability with large amount of NCG [6.15; 6.16]. Injection of NCG in PHP does not prevent their operation, but produces a rise of overall system pressure, operational temperature resulting to thermal resistance increase [6.17]. Resulting system pressure and saturation temperature can be expressed as follow [6.18]: where Pv and Pgas are the partial pressures of vapor (corresponds to the saturation pressure) and NCG gases, respectively. Finally, the real saturation temperature of the system consisting NCG will be:

𝑃 𝑠𝑦𝑠 = 𝑃 𝑣 + 𝑃 𝑔𝑎𝑠 ( 6
𝑇 𝑠𝑎𝑡,𝑠𝑦𝑠 = 𝑇 𝑠𝑎𝑡 [𝑃 𝑠𝑦𝑠 ] (6.2)
From Eq. ( 6.1) and Eq. ( 6.2) clearly seen, that internal pressure of the system increases with NCG presence, leading to the saturation temperature augmentation.

In this section will be discussed the effect of NCG on the flat plate PHP operation in vertical orientation (BHM). The system pressure with NCG addition corresponds to 15 kPa at 20 o C, contrary 2.3 kPa for pure water. However, average evaporator temperature increase have been found around ⁓10 o C for heat loads in the range of 50 W -150 W.

Results for FPPHP tested in horizontal position.

The fact that, as previously mentioned, gravity effect on the FPPHP operation is minimal in horizontal orientation (gravity vector perpendicular to the flow direction) makes the tests in such position closer to the space conditions than tests in vertical BHM.

As mentioned in section 3.2.4, the fluid flow pattern in horizontal orientation can be only slug flow, and the motor of the fluid flow comes from the pressure instabilities in the various vapor bubbles. In such configuration, surface tension forces become significant in the momentum balance, and wettability also plays a role in heat and fluid flow behavior of PHPs.

In analogy with section 6.1.1, in this section results will be presented in form of temperature histories to evaluate a thermal transient behavior of the FPPHP and in form of thermal resistance curves to compare influence of different parameters (condenser cooling temperature, filling ratio, fluid properties etc.) on the FPPHP thermal transport efficiency.

Qualitative analysis of the FPPHP filled with pure fluids

Typical transient long-time test run results for the same working fluids tested under the same conditions presented in section 6.1.1.1, are presented in Fig. 6.11 -6.13 in the forms of temperature and pressure histories. All data of this section can be found in appendix A.2 (Fig. A.2.1 to A.2.16).

For reported operating conditions, except for some oscillations noticed at the beginning of heat power level of 50 W, FPPHP filled with pure water operates "dry" after 800 s (Fig.

6.11

), what can be provoked by liquid accumulation in the condensation zone (see Fig. Temperatures and pressure behavior of ethanol-filled FPPHP are presented in Fig. 6.12.

As for the case with water, some oscillations with following dry-out are noticed for first heat power load (50 W). Regular high amplitude and high frequency temperature oscillations can be observed for higher heat loads. It is interesting that sometimes high amplitude (~40-50 o C) temperature oscillations alternate with lower ones (20-25 o C) but with higher frequency. In such configuration, one can assess that the FPPHP filled with water operates at 40°C cooling temperature but in a degraded mode. Such high temperature oscillations are detrimental to prolongated and durable use of FPPHPs as heat transfer device.

Operation behavior for the FPPHP filled with methanol (Fig. 6.13) looks quite similar to the case of ethanol, except period for applied power of 50 Wif, for ethanol, dry-out has occurred, stable and low amplitude temperature oscillations have been observed for methanol, showing very good thermal operation in such conditions. Test periods for higher heat loads characterized by regular high frequency temperature oscillations and overall These values mean the presence of the alternate slug-plug flow through pressure sensor in the center of evaporator. No significant pressure oscillation for FPPHP filled with ethanol and operated in vertical position was foundcan be concluded that device working in thermosiphon mode. However, FPPHPs tested in horizontal orientation and filled with ethanol and methanol operated in clear oscillation mode starting from 100 W of heat load for both devices (high amplitude pressure oscillations can be found in Fig. 6.12 (starting from 1800 s) for ethanol and Fig. 6.13 (starting from 2200 s)).

Influence of the condenser cooling fluid temperature

As shown in Fig. 6.11, FPPHP filled with water and tested at Tcool = 20 o C of cooling water temperature operates in the "dry evaporator" mode and results have been obtained

only for 50 W and 100 W of heat load. Thermal resistances for these two heat power are presented in Fig. 6.14 (blue solid line), where thermal performance of water-filled FPPHP slightly degrades with heat load augmentation, in contrast with case of ethanol and methanol filled devices. If the thermal resistance for ethanol at the lowest heat load (50 W) is a bit higher than for water (0.55 K/W vs. 0.49 K/W), it significantly decreases with heat power augmentation.

This is due to the transition from "dry" operating mode to "oscillatory" one which has not been reached for pure water. It is noticeable that thermal resistance almost converge until the same value for all three cases at highest heat loads (150 W and 200 W).

Influence of the filling ratio

Filling ratio influence on thermal resistance of the FPPHP filled with water and tested in horizontal position is presented in Fig. 6.15. Augmentation of filling ratio from 50% to 59% at cooling liquid temperature of 20 o C leads to startup at 100 W of applied power with stable but high amplitude temperature oscillations up to ~70 o C (Fig. . From Fig. 6.15, it is clearly seen that thermal resistance for higher filling ratio decreases with heat load augmentation. Also, higher filling ratio leads to temperature oscillations stabilization for higher cooling temperature (Tcool = 40 o C) which leads to higher thermal resistance decrease compared with FR = 50%. For Tcool = 40 o C, influence of filling ratio on thermal resistance is not as significant as for Tcool = 20 o C but stabilization of temperature oscillation has been observed (Fig. .

Influence of the binary mixtures

If for FPPHP tested in vertical orientation any significant effect of fluid thermophysical properties modification on thermal performance was not noticed (even for some cases FPPHP filled with water operated quite better than water-based mixtures), ethanol and methanol additives positively affected FPPHP operation in horizontal conditions. FPPHP filled with alcohol aqueous mixtures operates stable with frequent and low amplitude temperature oscillations for both fluids and both cooling fluid temperatures, even better than FPPHP filled with pure ethanol and methanol (Fig. A.2.1,A.2.2,A.2.9,. Empty FPPHP With condenser cooling temperature of 40 o C, the thermal resistance decreases (Fig. 6.18) with heat load increase for both fluids. But, if for water the FPPHP startup and stable "oscillatory" operation occurs at 100 W of applied power, for water/surfactant case, operation is not stablewith rare high temperature peaks and almost smooth -driedperiods, which could be provoked by foam generation (see section 4.6.3) and augmentation of flow resistance inside the channel.

Influence of the non-condensable gases presence

The temperature and pressure histories for water-filled FPPHP with non-condensable gases presence and tested at Tcool = 20 o C are presented in Fig. 6.19. Period for 50 W of heat load characterized by almost horizontal (500 -800 s) pressure and temperature curves with values of ⁓30 kPa and ⁓72 o C, respectively. Same behavior have been observed for the device filled without NCG presence (1100 -3000 s in Fig. 6.11) but with lower pressure and evaporator temperature values (⁓6.5 kPa and ⁓50 o C, respectively). However, with heat load augmentation till 100 W, start-up is reached and device operates with high amplitude and frequent pressure oscillations, contrary the case with pure water, where any oscillatory operation was not found. 2) and overall system temperature increase.

Conclusions on ground tests

During ground tests, the parametric study of the FPPHP filled with different fluids was provided. In vertical position, PHP operates almost all time as an interconnected thermosiphon. Device, operated in horizontal conditions operated in the oscillatory mode then star-up is reached (usually at 100 W of applied power), except water-filled FPPHP tested at Tcool = 20 o C, which has worse thermal performance in these conditions. Alcohol aqueous mixtures using as a working fluids lead to the thermal performance improvement, comparing to pure liquids (especially water). Condenser fluid cooling temperature augmentation also leads to FPPHP heat performance improvement. Device, filled with water-surfactant solutions shown very unstable operation despite lower thermal resistances compared with pure water. Non-condensable gases presence leads to overall system thermal performance degradation, but provokes stable oscillatory operation of tested FPPHP.

Parabolic flight tests

Characterization of the FPPHP for different gravity levels has been done during 69 Tables with experimental parameters provide following information about tested FPPHP and test conditions: metal-base plate material (Momolybdenum and Cocopper) in which channels are engraved and material of covering plate (sapphire -Sa, and copper -Co, for example); working fluid and filling ratio; applied heat loads, and gravity acceleration levels.

The FPPHP has initially been installed in a vertical position related to the floor of the aircraft. As it was previously discussed, in vertical position gravity vector assists the liquid plugs to flow back to the evaporator zone (in bottom heating mode) but, for parabolic flight tests, gravity level influences liquid plugs movements just during the periods between parabolasduring microgravity periods position of device is irrelevant. During this study, tests have been performed for FPPHPs in bottom heated mode (BHM) under different heating configurations.

After parabola "0" of each test day, power has been provided to the heaters. After the end of each set of five parabolas, value of heat load has been changed to planned one for next parabolas set (or kept the same if necessary) to reach quasi-established thermal regime.

The detailed description of the parabolic flights and test procedure are presented in sections 5.1 and 5.4.2.

Overall thermal behavior of the FPPHP tested during PFCs

During the short periods of microgravity (~20 s), both thermal and hydraulic regimes are not established due to high thermal inertia of the FPPHPs. First of all, Fig. 6.20 represents temperature and pressure variations for the molybdenum device covered with sapphire plate, during a set of five consecutive parabolas inducing normal, hyper-and micro-gravity conditions, for 100 W of applied heat power. Evaporator temperature curves are colored in the shadows of red, adiabatic zone temperature is in orange and cooling one in blue. Gray curve represents values of the system pressure and dashed grey one level of gravity acceleration. First, it should be noted that hypergravity phases have practically not significant influence except a slight decrease of the temperature and pressure profiles compared to normal gravity conditions. A little difference between the evaporator temperatures can be observed for normal and hypergravity conditions. But, with the onset of microgravity, a decrease of this gap and a homogenization of the temperature profile occur, and is maintained until the gravity level normalization. Then, a temperature rise is observed for the evaporator temperatures during each microgravity phases. The latter is due to all liquid collection in the condensation zone and so to evaporator dry-out (Fig. 6.21)heat being transferred from evaporator to condensed only by FPPHP heat conduction through the metallic and sapphire plates. As could be seen in Fig. 6.21a, during normal gravity, semi-annular/annular flow patterns are predominant for ethanol-filled FPPHP, but during almost all parabolas (without temperature drops) liquid remains accumulated in the condenser zone in the form of liquid plugs. At the same time, without liquid re-flow into evaporator, dry-out occurs with low amplitude liquid plug oscillations.

Sometimes, evaporator temperatures suddenly and arbitrarily drops during microgravity periods, as well observable instantaneous pressure rise which corresponds to some changes in flow regime and fluid entering in the evaporator causing sensible and latent heat transfer between superheated walls and subcooled liquid also called "re-activation" phases (Fig.

Influence of fluid thermodynamic properties on FPPHP behavior during parabolic flight campaign.

Results for FPPHP filled with water, water/16% ethanol, water/28% methanol and water/5% 1-butanol solutions and obtained during 71 st and 74 th ESA Parabolic Flight

Campaigns will be presented in this section. Note that only copper/copper FPPHPs have been used for these tests campaigns, so visualizations were obviously impossible. Thus, following analyses and explanations will be drawn based on our experience of previous visualization campaigns on such device. Fig. 6.22 -6.25 represent temperatures and pressure behaviors for all four tested cases at 50 W of applied power (figures for 150 W of applied power are available in Appendix B). In analogy with previous sections, curves, representing evaporator temperatures, are colored in the shadows of red, adiabaticyellow and blue for secondary cooling liquid. Pressure is represented by grey curves and gravity acceleration by grey dashed curves.

. During normal gravity periods, FPPHP filled with water operates stably with low temperature oscillations (in the range of the measurements uncertainty). Some temperature changes have been observed due to changes of secondary cooling fluid temperature. Should be noted, that secondary fluid temperature is not constant, but slightly increases with time, due to heat sink direct link with ambient air in the aircraftso, cooling temperature depends not only on heat load, but also on temperature of surrounding air. That said, hypergravity periods are accompanied with small temperature drops and pressure augmentation. Temperature histories for device filled with water/ethanol mixture are presented in Fig. 6.23. Because it was the first test for the second day of flights and heat load has been applied just before end of parabola "zero", temperature increases in a similar way than the one observed in vertical on-ground testswithout oscillations, showing that the start-up did not really occurred. It seems that temperature level becomes stable just before first parabola, and the transition between microgravity and beginning of hypergravity phases provokes the change of device operation mode accompanied with temperature drop and following stabilization with lower temperature level and low temperature oscillations amplitude. Then, microgravity periods, are characterized by evaporator temperature augmentation but this time with more frequent temperature fluctuations, attributed to reactivations phases, than with water case. Probably, for this case, liquid was not accumulated in the condenser zone and dry-out was not reached, but FPPHP operates in the pulsating mode with reduced From Fig. 6.24 it could be seen that for FPPHP filled with water/methanol mixture, high amplitude fluctuations are present for normal gravity periods. However, and unfortunately, any temperature or pressure drops have not been observed for microgravity phases. This fact indicates that for this fluid at 50 W of applied power, dry-out without any flow reactivations takes place during each parabola. (Fig. 6.25), especially in horizontal orientation. Thermal behavior for the FPPHP filled with this solution and tested during parabolic flight at 50 W of applied heat power is very similar to the case with pure water -during normal gravity period any significant temperature and pressure oscillations were not observed. Moreover, during all five parabolas any fluid flow reactivation accompanied with temperature drop have not been registered. Nevertheless, despite similarities in behavior, the overall evaporator temperature is almost 10 o C lower for 1-butanol solution than for pure water. Better channel walls wetting and higher thin-film evaporation rate could have caused this. Heat load augmentation up to 150 W leads to overall evaporator temperature increase and provoke oscillations with higher amplitude during normal gravity periods. Note that for the other cases, temperatures peaks are almost equal for all parabolas (during one set), but for water/1-butanol solution, the peak temperature values varies from one parabola to another.

.4. Influence of the non-condensable gases presence on FPPHP operations under microgravity conditions

All results obtained for coppersapphire (Co/Sa) FPPHP during ESA 71 st Parabolic Flight Campaign are presented in the form of temperatures and pressure evolutions, as well as differences between average evaporator temperature during normal gravity periods and peaks during parabolas, to explain FPPHP operating behavior for different gravity levels with and without non-condensable gas injection.

Temperatures, pressure and acceleration profiles for the FPPHP tested at 100 W applied heat load on a set of five subsequent parabolas are shown in Fig. 6.26 for the PHP filled with deionized water, and Fig. 6.27 for the PHP with water with NCG, respectively. [6.18], NCG injection to the close equilibrium two-phase systems leads to total system subcooling, which means that boiling will start with higher surface superheat than for pure water. Gas presence in slug form, in the evaporation and condensation zones, also reduces heat transfer effectiveness due to heat transfer surface occupation and liquid/vapor contact prevent with it.

For higher heat loads (100 W and 150 W, Fig. 6. 27 and B.2.6), the presence of NCG seems to provoke regular temperature and pressure fluctuations, whatever the gravity levels, which means alternate movements of liquid and gaseous (vapor/air) phases in form of slug/plug flow, unlike the case with pure water (Fig. 6.26). Values of evaporator temperature differences ("ΔT") between average temperatures during normal gravity and maximum temperatures during microgravity periods are presented in Fig. 6.28 as functions of parabola's number. Tests for low heat loads show almost the same values, with and without NCG, which can be explained by short microgravity periods and high PHP thermal inertia. Following power increase shows progressive and noticeable rise of temperature peaks for pure water (almost 7 times higher at 150 W than at 50 W); but these rises are much higher than those with NCG. Note that for the 4 th parabola with instabilities, at 100 W, overshoot value is lower than for other peaks, but still higher than with NCG. On one hand, the curves of ∆𝑇 do not almost vary (horizontal line) for PHP with pure water, which demonstrates a certain regularity. On the other hand, for PHP with NCG, (some values at 100 W or 150 W are lower than at 50 W). Fig. 6.29 presents pressure and temperature curves in the evaporator near pressure transducer location. Fig. 6.29a represents two consecutive parabolas (3 rd and 4 th ) for pure water at 100 W applied of heat power. During the first one, smooth temperature peak and evaporator dry-out have been observed, while, during the second one, instabilities have been noticed. During normal gravity periods, pressure oscillations are relatively low (note that oscillations here correspond to the pressure transducer uncertainty range). Some pressure fluctuations are however noticeable just as temperature ones, due to some spontaneous fluctuations in gravity levels (see dotted grey curve). During microgravity periods (see first parabola), pressure reaches peak value with following sharp drop, but on the second parabola short pressure stabilization with subsequent increase until parabola's end can be noticed.

Results comparison for micro(µ), Lunar and Martian gravity levels.

This subsection will present discussion on gravity level influence on water filled FPPHP operation behavior for 150 W of applied power, as a more representative case. All tests have been performed during the ESA 74 th Parabolic Flight Campaign.

Thermal behavior of water-filled FPPHP tested under microgravity conditions (presented in Fig. 6.30) have been casually discussed in previous sub-section. As shown in Fig. 6.30, normal gravity conditions are characterized by evaporator temperature smooth oscillations. In the same way, during microgravity periods, FPPHP mostly operate in "dried" mode, accompanied with high temperature rise, with short-term reactivations (parabolas 2, 3 and 5). For lunar gravity conditions, evaporator temperature increases (~2-3 o C) are observed (Fig. 6.31). The maximum peak temperature is about three times lower than for microgravity.

High and very frequent temperature oscillations, as well pressure oscillations, have been Fig. 6.32 represents temperature thermal behavior of device tested under Martian gravity conditions. It is clearly seen that there are not any significant temperature or pressure changes caused by martian gravity. It seems that the level of gravity does not influence that much the operation of the FPPHP, at least on a range from the Martian (0.38 g) to the hyper (1.8 g) gravity levels. Unfortunately, this point has not been treated in the frame of this work, and would require further study. smooth increase of the system pressure caused but evaporator dry-out and vapor superheating. However, some microgravity periods were accompanied with sudden temperature and pressure drops, which mean flow changes inside the PHP (will be analyzed in next chapter). For the device filled with water-ethanol mixture, these flow re-activations were observed much often than for other tested fluids. Presence of the NCG provokes not only the overall temperature augmentation, but also induces stable oscillatory operation of the device during both normal and microgravity phases with temperature increase (during microgravity) quite lower than for pure water (similarly to behavior observed and explained in section 6.1.2.5). Concerning Parabolic Flight Campaigns, during microgravity periods, significant increase of evaporator temperature has been observed for all tested fluids. This could be a result of liquid accumulation in the condenser zone and of evaporator dry-outheat being transferred just via thermal conduction through the FPPHP solid material plate.

Conclusion on parabolic flight tests

Interesting results have been found for FPPHP filled with water when non-condensable gases are present-despite the overall thermal performances degradation (caused by system sub-cooling), pulsating heat pipe with NCG injection shows steady oscillatory operation with stable slug/plug flow pattern even during microgravity periods, unlike cases with pure liquids, with means absence of the stopover phenomena.

Using water/alcohol mixtures leads to more frequent fluid flow reactivations, accompanied with temperature drops, during microgravity phases, especially for higher values of heat load (150 W).

Brief analysis of the gravity level influence (tests for Lunar and Martian gravity levels)

has shown that minor gravity values (0.18g) is enough to maintain continuous oscillation mode of FPPHP operation with some evaporator temperature augmentation (quite less than for microgravity periods) during reduced gravity phases. Almost any changes for FPPHP operation during Martian gravity phases, compared to normal gravity conditions, have not been observed.

Two-phase flow determination inside the FPPHP during microgravity

The presence of the whole menisci in every channels and liquid accumulation in the condenser zone are clearly visible in temperature decrease. Menisci movements and re-activation mechanisms, as well stopover mechanism, will be explained in the next sections.

Video post-processing tool

The general idea of the image post-processing in current work is the determination of As the infrared data can be exported in numerical format (matrix of temperature values), the temperature drop can easily be recognized and position can be extracted in the form of pixel's number. This procedure, repeated for every pictures in the matrix, allows to define the positions of menisci for the specified period in every channels. Because the purpose was to visualize the whole channels of the entire FPPHP, the main disadvantage of this method is a very low resolution of the infrared camera (resulting resolution is 2-3 pixels for a 3 mm channel's width) which did not allow to provide the analysis for most of registered videos due to vibrations occuring during parabolic flightssometimes zone of interests being displaced to the channel wall resulting in numerous parasitic interfaces detection.

On the basis of this observation, very similar approach has then been developed for visual images from high speed camera. As shown in fig. 7.3, the menisci are contrasting with channel (greyscale image with light intensity between 0-255) -interfaces exist as some change in intensity across the channel and their position can be extracted using the horizontal gradient filter.

The filtered image now contains interfaces, but other horizontal gradients caused by changes in lighting and general camera noise can also be observed. These have lower intensities than the interfaces and are mostly one or two pixels in thickness (similarly to some interfaces).

Finally, an approach based on the OpenCV [7.12] library has been implemented to track the menisci along the channels. OpenCV (Open Source Computer Vision Library) is a library of programming functions mainly aimed at real-time computer vision and consisting most of basic image manipulation tools. The library is cross-platform and free for use under the open-source license. To process rough images, first, they have been extracted from raw videos in the form of image frames sequences and then converted to black and white. After that, feature definition and object tracking tools, available in the OpenCV® tool-set, have been used for the menisci tracking. The main purpose of these tools is the manual definition of the vapor/liquid interfaces and their following automatic tracking through time. One of the advantages of this approach is the possibility to define the flow transition from slug flow to annular flow pattern -or another oneif interface cannot be defined anymore (due to highly deformed menisci). Therefore, in the following, critical velocities/acceleration values can be defined as the last values obtained before menisci "disappearance" or their deformation, corresponding to transitional flow patterns.

To validate the above developed approach, available images from previous studies of tubular pulsating heat pipe [7.13] have been used for the generation of the composite image (Fig. 7.4). Then, the image of menisci is isolated and placed over an extracted backdrop of the fluid regions in the image. The menisci are then moved with a given velocity profile (sinusoidal). The developed framework, used to determine menisci velocities from synthetic images, allowed comparing the latter with programmed velocity profile and showed great consistency in the results (see Fig. 7.5). Unfortunately, it must be said that, during normal and hyper gravity conditions, analysis is impossible via all methods above due to the absence of menisci in such annular flow patterns. Thus, analysis could be done just for stopover and flow re-activation during microgravity periods.

Motion of liquid plugs

For better readability of the acquired data, an extra focus will be done on the examination of two nearby liquid plugs: LP67 and LP89, which are located in the center of the FPPHP. Fig. 7.6 depicts these two liquid plugs, as well as the channel numbers. to influence of small pressure variations in the evaporator caused by vapor superheating, which was already observed and described in previous studies [7.14]. Note that the mirror effect between the displacement of the left-hand side meniscus is well reproduced by that of the right-hand side one, corresponding to the same liquid plug, the latter being incompressible (a small difference can be attributed to measurement error and a very slightly higher velocity for the receding meniscus compared to the advancing one due to the downstream deposition of the thin liquid film which causes a local low loss of mass).

However, at the beginning of microgravity phases, interfaces oscillations amplitudes are quite high due to influence of inertia forces and, after stabilization, surface tension became dominant.

Despite the rise of the interfaces motion amplitude during re-activation periods (Fig. 7.7, red field), liquid slugs do not attain the evaporator zone (due to the low filling ratio, i.e.

40%), which is not represented in the y-axis of Fig. 7.7 (situated between 0 cm and 4 cm positions). This means that the fluid flow transition is probably not triggered by evaporation in the evaporator but in the superheated adiabatic zone. On the other hand, high menisci motion amplitudes lead to the liberation of the wall surfaces in the condenser (located between 12 cm and 20 cm in y-axis of Fig. is equal to 0.15 m/s, which is one and half times higher than the commonly known operating velocity of 0.1 m/s in the literature [7.15]. It is also higher for ethanol compared ot the ones obtained with FC72. It must be noted that, sometimes "transition velocity" corresponds to the menisci deformation, but not to their total "disappearance". However, reaching these velocities gives us a flow pattern transition represented in Fig. 7.1b (right part).

Non-dimensional analysis of flow pattern transition

The obtained transition velocities allow evaluation of commonly used criteria for different, but not very far from the reference value. Note that any dependencies with applied power have not been founded. Earlier, modified non-dimensional numbers, based on operation limit map to quantify the effect of interface disturbance propagation and surface tension to actual flow conditions such as velocity and acceleration, have been proposed by [7.18]. This operation limit map is On placing the non-dimensional numbers obtained through previous post-processing in the operation limit map and verifying with flow pattern correlations, all the points from postprocessing are most likely to fall nearly to the upper coalescence and transition regions (Fig. 7.12) irrespective of the working fluid or channel diameter. This discrepancy may arise due to that estimated modified non-dimensional numbers did not account the bubble length to channel diameter ratio (used in [7.18]). However, more test points are necessary to determine their location in the flow map and to validate this approach.

Flow influencing forces analysis

The following momentum equation [7.19] could be used to calculate the magnitude of the driving pressure as well as the impact of the influencing forces on the motion of liquid plugs (Fig. 7.14):

𝑚 𝑙 𝑑𝑢 𝑙 𝑑𝑡 = 𝐹 𝑑𝑟𝑖𝑣 -𝐹 𝑣𝑖𝑠 (7.7)
where Fvis and Fdriv are viscous, and driving forces, respectively. These forces are calculated with eq. (7.8). Gravity forces are not taken into account in this calculation due to the microgravity conditions. Finally, from the knowledge of every post-processed parameters for every liquid plugs, it becomes possible to evaluate the driving force, Fdriv, from Eq. (7.7) and (7.8).

Due to specific interest in flow transition and almost similar motion amplitudes and frequencies for all liquids plugs, calculated forces magnitudes just for one "surrounding"

and two central liquid plugs (for channels 1-14, 6-7 and 8-9, see Fig. characterized by the significant amplitude rise of all forces oscillations for high heat loads, which was also reported for the transition from the stopover period to a large-amplitude oscillations for horizontal micro FPPHP [7.20], but with quite higher values (which could be caused by the microscale channel dimensions, i.e. of 1×0.5 mm²). Fluctuations for lower heat powers are less noticeable.

It is important to note that the viscous forces are the weakest ones, while not being negligible for both cases during dry-out phases; but this influence comes out just before transition (as can be seen in Interesting fact is that the transition takes place not at the same time for each liquid plugs, but with short time delay, that may be due to the vapor compressibility (dash line in Fig. 7.14).

Knowing the driving forces, the resulting values of the driving pressures ∆𝑃 𝑑𝑟𝑖𝑣 (eq. 7.12) are plotted in Fig. 7.15. Note that the capillary pressure drop, ∆𝑃 𝑐 , is directly included in the driving pressure term in eq. (7.12) due to the fact that its value is not estimable in these measurements. In analogy with the previous figures, driving pressures are plotted for the liquid plugs 6-7, and 8-9, in Fig. 7.15 for the periods from start of microgravity (at t = 0 s) until the re-activation phase with applied power of 100 W (Fig. 7.15a) and 150 W (Fig. 7.15b), respectively. In parallel, the experimental absolute pressure (Pexp), in the evaporator zone between these two liquid plugs, is also plotted in the same graphs. One can see that the absolute pressure shows increased instabilities during the re-activation phases, synchronously with the calculated driving pressure fluctuations from the post-processing of the image analysis. Here it can be seen that during total period until re-activation, values of driving pressure for the two liquid plugs seems to be almost similar with a slight time lag, due to the vapor compressibility (as shown previously). Amplitudes of pressure oscillations increase up to the value attained at the re-activation time and do correspond to significant absolute pressure surge between two liquid plugs. Interesting to note, that the final values of driving pressure (Pexp) (corresponding to the flow pattern transition) are almost similar for both values of applied heat power and are of about 350-380 Pa. Note that this pressure difference is one order of magnitude higher than the maximum capillary pressure drop, from minimum receding contact angle to maximum advancing contact angle (∆𝑃 𝑐,𝑚𝑎𝑥 = 2𝜎 𝑅 ⁄ ), of about 20-28 Pa when calculated in the temperature range of Fig. 6.20. This means that the driving pressure difference is largely counterbalanced by the inertial and viscous forces before the rupture of the menisci at the transition between slug flow and annular flow.

Conclusions

The object tracking based visual analysis methodology for liquid-vapor interfaces position determination has been developed, validated and implemented to study the flow behavior during stopover and re-activation periods under microgravity conditions.

Experimental investigation of these phenomena inside flat plate pulsating heat pipes filled with ethanol (3 mm diameter) and FC72 (1.5 mm and 3 mm diameters) and operating under microgravity is also performed.

Microgravity periods, as previously reported [7.14], are accompanied by the stopover phenomena and evaporator dry-out that leads to overall thermal transport degradation. It was Chapter 8

Conclusions and Perspectives

Conclusions

The work presented here can be divided in two main parts respectively describing the parametric studies on the FPPHP and the analytical analysis of the hydraulic behavior of the FPPHP during stopover and flow re-activation periods. The parametric studies has consisted in the evaluation of the device orientation, gravity, working fluid, filling ratio, condenser temperature and non-condensable gases presence influence on the FPPHP operation. For the analytical part, only the videos obtained during 64 th and 69 th parabolic flight campaigns for three different FPPHPs filled with ethanol and FC72 have been used for analysis.

Related to the parametric studies, the following conclusions can be presented:

 The FPPHP tested in vertical orientation (bottom heating mode) mostly operates as an interconnected lopped thermosiphon (except water-filled device) due to relatively low surface tension and liquid stratification (hydraulic diameter is higher than static capillary limit, linked to the Bond number). Condenser temperature increase leads to significant temperature rise in evaporator zone. Higher filling ratio almost has no effect on the device performance. Non-condensable gases addition into the system leads to the total thermal performance degradation but also to the high-amplitude temperature and pressure oscillations appearances.

 Water-filled device tested in horizontal orientation at low cooling temperature has not reached start-up and operated in "dry" mode. Heat load increase leads to the total system failure. Other tested fluids show high amplitude and frequent evaporator temperature oscillations. Performance improvement and earlier start-up observed with condenser temperature increase as well for the case with higher filling ratio. Non-condensable gas presence provoked stable oscillatory operation for water-filled PHP, which was failed with pure water.

 Microgravity periods during parabolic flights are characterized by the significant increase of evaporator temperature for all tested fluids, which could be explained by dry-out and insufficient parabola duration for start-up (like temperature augmentation before start-up for horizontal orientation). The non-condensable gas addition leads to the continuous oscillatory operation, without significant difference in temperature and pressure amplitudes, comparing to the normal gravity periods.

In the last part of this thesis, post-processing of the video obtained during microgravity periods and its analysis has been presented. Evaporator dry-out starting at the beginning of parabola has been confirmed. However, during some parabolas, stop-over periods are interrupted by a flow pattern transition from slug flow to annular flow with uniform liquid distribution inside pulsating heat pipe. This flow pattern change is of particular interest to evaluate conventional criteria of flow transition regarding microgravity conditions. As a consequence, analyses performed for this particular case can be summarized as follows:



Résume

L'intégration de dispositifs électroniques de haute puissance et dissipant des centaines de watts par centimètre carré au sein des systèmes spatiaux provoque une hausse significative des besoins en systèmes de refroidissement de haute performance, légers et efficaces. En raison de ces exigences spécifiques, l'utilisation des systèmes de refroidissement diphasiques représentent un grand intérêt pour ces applications spatiales. Les caloducs oscillants, dispositifs passifs et simples de transfert de chaleur diphasique sont devenus une excellente solution pour les systèmes de management thermique spatiaux. Actuellement, le fonctionnement de ce type d'échangeurs de chaleur n'est pas entièrement étudié, notamment en conditions de microgravité. C'est pourquoi, dans le cadre des projets successifs INWIP et TOPDESS de l'ESA, l'évaluation des performances des caloducs oscillants plats et des spécificités opérationnelles sous différentes conditions de gravité a été défini comme principal objectif du présent travail de thèse.

Les études expérimentales menées au cours de ce travail ont porté sur la caractérisation thermique des différents caloducs oscillants testés au sol et au cours de quatre campagnes de vols paraboliques de l'ESA. Des études paramétriques des performances de ces systèmes ont été effectuées en configuration verticale et horizontale, en utilisant différents fluides de travail (mélanges alcool-eau, fluides remouillants ou solutions de surfactant), différentes températures de condenseur, plusieurs concentrations de gaz non condensables et différents niveaux de gravité (terrestre, martien, lunaire, hyper et microgravité). L'amélioration des performances, parfois inattendues, ont été observées sous certaines conditions opératoires. Les études analytiques de l'écoulement diphasique dans le caloduc oscillant, concernant notamment le passage de l'arrêt de fonctionnement aux oscillations de forte amplitude (« phases de réactivation ») ont été réalisées grâce à des visualisations haute fréquence. Pour finir, les critères de transition d'écoulement du régime bulles/bouchons au régime semi-annulaire/annulaire a été étudiée dans le cadre d'une l'analyse non-dimensionnelle des principaux paramètres de l'écoulement.

Mots-clés: Caloducs, Composants électroniques--Systèmes de refroidissement, Écoulement diphasique, Gravité réduite, Transfert de chaleur, Caloduc oscillant, Écoulement bulles/bouchons, Écoulement annulaire

Abstract

Integration of high power and performance electronic devices in modern spacecraft systems with heat generation of hundreds of watts per square centimeter crucially increases needs in high performance, low weight, energy efficient and reliable thermal management systems. Due to these specific demands, two-phase heat rejection devices are of major interest for space applications. The Pulsating Heat Pipes, as simple passive twophase heat transfer devices, have become an excellent solution for space cooling systems. Today, operation of this kind of heat transfer devices is not fully studied, especially in microgravity. So, in the context of ESA consecutive projects INWIP and TOPDESS, evaluation of the flat plate pulsating heat pipe performances and operational peculiarities under different gravity conditions has been defined as the main aim of the present PhD work.

Experimental studies conducted during this work concerned the thermal characterization of different flat plate pulsating heat pipes tested on ground and during four ESA Parabolic Flight Campaigns. Parametric investigations have been carried out for vertical and horizontal orientations of the device using different working fluids (including alcohol aqueous solutions, self-rewetting fluids and surfactant solutions), condenser temperatures, non-condensable gases concentrations and gravity levels (Earth, Martian, Lunar, hyper and microgravity). Obvious, and sometimes unexpected, performances improvement has been observed under certain conditions. Analytical studies of the two-phase flow inside the FPPHP, especially concerning the transition from stopover to high-amplitude oscillations (fluid flow "re-activation phases") have been done thanks to highfrequency visualizations. Finally, the flow transition criterion, from the slug-plug flow regime to semiannular/annular flow, was studied thanks to dimensionless analyzes considering the main parameters of the flow. Keywords: Heat pipes, Electronic apparatus and appliances -Cooling systems, Two-phase flow, Reduced gravity environments, Heat-Transmission, Pulsating heat pipe, Slug flow, Annular flow

3. 7 .

 7 Pressure drop in Taylor bubble flow assuming (a) no pressure in individual bubbles and (b) liquid plug advancing angle lower than π/2 [3.4]. .............................................

  Abbreviations

  researchers [1.4; 1.5]. Therefore, this technology needs further investigations related to the understanding improvement of the phenomena inside the PHP. This thesis has been realized in the context of two successive research projects funded by European Space Agency -INWIP and TOPDESS -which are dedicated to the development of the high efficient and passive thermal management technologies for the space applications. The first project -Innovative Wickless Heat Pipe Systems for Ground and Space Applications (INWIP)aims at designing, manufacturing and testing innovative wickless heat pipes with enhanced performances and reliability for both space and ground applications. Based on the research consortium experience and fields of interests, research work is focused on the numerical and experimental investigations of three two-phase heat transfer devices: wickless hybrid thermosiphons, grooved heat pipes and pulsating heat pipes. Parametric studies of these systems include the performance evaluations of the several working fluids, functionalized surfaces (hydrophobic and hydrophilic), gravity influence (different orientations during on-ground tests and reduced gravity during parabolic flights). This project is a first preparatory step for the further systematic experimentation on the Thermal Platform1 TP1 onboard the International Space Station, investigating particular grooved and pulsating heat pipes filled with ordinary and self-rewetting fluids. The second project -Two-Phase Passive Thermal Devices for Deployable Space Systems (TOPDESS), continues developing the research work started through the previous ESA MAP INWIP project, but this time with the ambitious purpose of dealing with deployable and/or flexible two-phase wickless devices for developing the applicative aspects. Nevertheless, the deployability or flexibility of such systems will not be studied in this thesis. We will mainly focuses on heat and mass transfer inside flat-pate pulsating heat pipes. In particular, tests of the improved version of the Flat Plate PHP, developed during the INWIP project, should be done in order to meet the following objectives:  The high-speed grayscale and infrared visualizations of the entire channel array for an accurate detection of the liquid-vapor interfaces and measurements the liquid bulk temperature (to determine the local heat transfer coefficient);  A parametric study concerning the wettability of the fluid/metal couple, by using different surface coatings (oxidation, etc.) and/or fluid thermophysical properties by adding specific substances such as alcohols (self-rewetting fluids) or surfactants. Therefore, the main purposes of this thesis are to characterize the heat transfer performance of the flat plate pulsating heat pipe operating under normal and reduced gravity, understanding the two-phase flow behavior inside the FPPHP and increasing the heat transfer characteristics using working fluids with different transport properties. Cooling Systems for Ground and Space Applications To better introduce the pulsating heat pipe and describe motivation of this work, a short overview of the existing thermal management devices is presented in this chapter. As shown in Fig. 2.1, recent cooling solutions can be classified in the four main categories as a function of the amount of rejected heat flux and surface superheat [2.1]:  Natural convection;  Gas forced convection;  Liquid forced convection;  Two-phase cooling. It seems that two-phase cooling devices have several advantages compared to the standard single-phase cooling systems:  First, a fluid's latent heat of vaporization can be a few orders of magnitude larger than the specific sensible heat;

Figure 2 .

 2 Figure 2.1. Comparison of the various thermal management approaches [2.1].

Figure 2 . 2 .

 22 Figure 2.2. Classification of the two-phase heat transfer devices [2.2; 2.3].

2. 3 .

 3 Electronic devices are typically attached to the bottom side of the microchannel heat sink under the ribs in the central part. Open top side is usually covered by a planar plate with the inlet and outlet ports. With appropriate driving pressure, the fluid is driven through the microchannels and heat is transferred from the electronic component through the microchannels into the fluid.

Figure 2 . 3 .

 23 Figure 2.3. Typical microchannel heat sink with parallel channels [2.5].

Figure 2 .

 2 Figure 2.4. Schematic view of the spray cooling process [2.8].

Figure 2 .

 2 Figure 2.5. Schematic of the jetimpingement cooler [2.9].

Figure 2 . 6 .

 26 Figure 2.6. Thermosiphon schematic representation [2.13].

Figure 2 .

 2 Figure 2.7. Schematic of a thermosiphon loop [2.14].

Figure 2

 2 Figure 2.8. (a) Heat pipe working principle [2.17] and (b) types of capillary structures [2.18].

  2.11). At the evaporator, a porous wick acts as a capillary pump. The liquid evaporates by absorbing a heat flow from the evaporator. The vapor flows to the condenser through the vapor line and condenses in the condenser. The condensate returns to the evaporator through the liquid line. The control reservoir helps to absorb fluctuations and fluid expansion that may occur during loop operation. The reservoir also modulates the operating conditions of the loop, temperature and saturation pressure in particular. It is the position of the reservoir relative to the evaporator that leads to the distinction between the two types of two-phase loops: Loop Heat Pipe (Fig.2.11a) and Capillary Pumped Loop (Fig.2.11b).

Figure 2 .

 2 Figure 2.11. Schematic working principle of the loop heat pipe (a) and capillary pumped loop (b) [2.26].

Figure. 2 .

 2 Figure. 2.13. Basic construction of a closed looped pulsating heat pipe [2.29].

Figure 3

 3 Figure 3.1. Schematic representation of the different PHP configurations [3.1]: (a) Closed Loop Pulsating Heat Pipe, (b) Open Loop Pulsating Heat Pipe.

3 .Figure 3 . 2 .

 332 Figure 3.2. Difference between (a) tubular [3.2] and (b) flat plate PHP.

  Fig. 3.3.

Figure 3 . 3 .

 33 Figure 3.3. Taylor bubble chain in a vertical tube (left), the forces acting on the liquid plug and transfer processes in the single-cell scale (right) [3.4].

Figure 3 .

 3 Figure 3.4. Pressure-enthalpy diagram of the PHP operation [3.6].

Design parametersoo

  Open or closed loop configuration o Device material and type (tubular PHP or flat plate PHP) o Channel diameter and shape (alternate or not) o Number of turns o Length of the evaporation, adiabatic and condensation zones o Presence of valves (check valves or Tesla valves) Heat sink heat transfer coefficient and temperature

  tried to classify the channels, based on their diameters and/or on the possibility of the existence of a Taylor flow inside them. A few main classifications for the transition from macroscale to microscale flows, based on the hydraulic diameter, have been proposed during few last decades. First, Mehendal et al. [3.9] recommended a size-based classification as follows:  Microchannels (D = 1-100 µm)  Mesochannels (D = 100 µm to 1 mm)  Macrochannels (D = 1-6 mm)  Conventional channels (D > 6 mm) Later, Kandlikar [3.10] implemented the following classification, based on the channel diameter range and more reliable:  Microchannels (D = 50-600 µm),  Minichannels (D = 600 µm to 3 mm)  Conventional channels (D > 3mm) Unfortunately, transition criteria based just on channel dimensions do not reflect the influence of channel size on the physical mechanisms and do not take into account fluid properties. A more general definition should address to the interaction of different forces, as well gravitational, interfacial and viscous ones.

Fig. 3 .

 3 Fig. 3.5 and 3.6 schematically present the fluid distribution inside vertically oriented PHPs with different diameters under adiabatic and operating condition. Two cases in Fig.

3. 5 Figure 3 Figure 3 . 6 .

 5336 Figure 3.5. Effect of channel diameter on fluid distribution inside the circular tube under adiabatic and operation conditions: diameter (a) much and slightly (b) higher than critical one [3.5].

Figure 3

 3 Figure 3.7. Pressure drop in Taylor bubble flow assuming (a) no pressure in individual bubbles and (b) liquid plug advancing angle lower than π/2 [3.4].

Figure 3 . 8 .

 38 Figure 3.8. Sensible and latent heat transfer near meniscus region: (a) schema [3.17]; (b)

  3.9. The experimental comparison on the self-induced evaporation of n-hexane in capillary channels with square and circular cross-section, open to stagnant ambient air, confirmed a tendency indicating that evaporation in a channel with corners can be several orders of magnitude faster than in a circular channel with the same dimensions. From other side, prediction of the liquid film length and thickness and following modelling of the heat transfer became a very complicated task.

Figure 3 .

 3 Figure 3.9. Liquid film distribution in a vertical rectangular square channel [3.19].

3 .Figure 3 .

 33 Figure 3.10. Visualization of the fluid flow in adiabatic section: (a) annular flow, Q = 40 W and (b) semi-annular flow, Q = 100 W (copper FPPHP, 2×2 mm 2 , ethanol, FR = 50%, BHM) [3.22].

  3.11c. Due to liquid accumulation on one end of the device, the channels in the evaporator zone are totally occupied by vapor exposed to dried wall without liquid film, resulting in vapor superheating. In this conditions heat transfer from evaporator to condenser are done mainly by conduction with relatively poor thermal performance compared to the normal operation. When the fluid starts to flow, both ends of a U-shape liquid column oscillate in opposite directions. If the amplitude of oscillations is sufficient, liquid flow into the evaporator and vapor flow into the condenser occur, completing the circulation cycle.

Figure 3 .

 3 Figure 3.11. High speed visualization of the FPPHP operating in horizontal position (a) initial state; (b) after 1 min of heating; (c) steady-state operation (2×2 mm 2 , Q = 30 W, water, FR = 70%) [3.24].

  Figure 4.1. Effect of channel diameter on the PHP thermal performances for different working fluids: (a) water and (b) ethanol (FR = 50%, vertical BHM, Tcond = 25 o C) [4.1].

Fig. 4 .Figure 4 . 3 .

 443 Fig. 4.2b and 4.2c illustrate the temperatures behavior for both PHPs (circular and square cross-sections) as functions of time for different heat power input levels. The startup (red flash in Fig.4.2b) of the circular cross section PHP is observed since 200 W applied heat power; but, for the square cross section PHP, it happens earlier at 80 W applied power (red flash in Fig.4.2c). A second activation was observed at 160 W, accompanied by the evaporator and adiabatic temperatures decrease, and condenser temperature increase. From this heating power level, the amplitude of the oscillations decreases. In addition, thermal resistances of PHP with circular channel are quite higher than thermal resistances for square channel PHP, starting from 80 W heat applied power (Fig.4.3).

Figure 4 . 4 .

 44 Figure 4.4. Effect of inclination angle on the PHP thermal resistance: (a) conventional PHP (D = 2 mm, ammonia, FR = 50%, Tcond = 28 o C) [4.4].

  Figure 4.5. Influence of the (a) evaporator (FR = 50%, vertical BHM, Tcond = 25 o C) [4.1] and (b) condenser (1.5×1.5 mm 2 , acetone, FR = 50%, horizontal, Tcond = 20 o C) [4.8] lengths on the PHP thermal performance.

Figure 4 . 6 .

 46 Figure 4.6. Adiabatic length influence on the PHP thermal resistance (1.6×1.7 mm 2 , ethanol, FR = 50%, Tcond = 20 o C) [4.9].

Figure 4 . 7 .

 47 Figure 4.7. Influence of the number of turns on PHP performance (water, horizontal, Tcond = 25 o C) [4.7].

Fig. 4 .

 4 Fig. 4.8 describes the effective thermal resistances of a PHP (1 mm diameter, 150 mm length, 20 turns) charged with ethanol (Fig. 4.8a, FR = 50, 70, 85%) and acetone (Fig. 4.8b,

  Figure 4.8. Thermal resistance for PHP tested in vertical position and filled at different ratios with (a) ethanol and (b) acetone (1×1 mm 2 , BHM, Tcond = 11 o C) [4.14].

  satisfy the requirements related to the materialfluid compatibility, PHP channel diameter, working temperatures, pressure etc. But from the operational point of view, the fluid should have high value of the slope of vapor-liquid saturation line ((dp/dT)sat)[4.17] to ensure stable liquid-vapor pairs oscillations inside PHP.Kearney et al. [4.18] tested PHPs filled with ethanol, Novec 649 and Novec 7200 (their liquid/vapor saturation slopes are shown in Fig.4.10a in comparison with water). Based on the (dp/dT)sat values and experimental results, it seems that Novec 649 is the best working fluid (from the three tested) to ensure stable operation of PHP (Fig.4.10b). To confirm this, the temperature differences between evaporator and condenser are presented in Fig.4.10b for PHPs filled with ethanol and Novec 649 (FR=30%). It is clearly shown that, in horizontal inclination, the PHP filled with ethanol showed worse performance than with Novec 649; moreover, PHP filled with Novec 649 continued its operation till 0.8 W.cm -2 of heat flux, while PHP filled with ethanol failed at 0.6 W.cm -2 . In vertical orientation, PHP filled with Novec 649 again showed better thermal performances than with ethanol.Later, based on previous investigations and numerous assumptions,Kim and Kim. [4.19], proposed a figure of merit for the working fluid selection related to their micro FPPHP (80×35×1.7 mm 3 , D=0.67 mm).

  Figure 4.10. (a) (dp/dT)sat as function of temperature for various working fluids and (b) temperature difference between evaporator and condenser (1.55×1.9 mm 2 ,FC-72, FR = 30%, Tcond = 18 o C) [4.18].

  Figure 4.11. (a) Comparison of the thermal resistance of the MPHPs with the inverse of the pressure ratio (Eq. (4.4)) at various operating temperatures; and (b) change in the values of the proposed figure of merit (Eq. 4.6) for micro-FPPHP with operating temperature and working fluid (1×0.5 mm 2 , FR = 50%, BHM) [4.19].

  However, it is possible to change the properties of the different fluids by mixing them in various volume ratios to improve the performances of PHPs. During the last decade, authors studied PHPs operation with classical binary mixtures in different proportions, as well water-acetone, water-ethanol, ethanol-acetone [4.20; 4.21; 4.16], and even immiscible water-HFE7100 mixtures[4.22] and pentane/heptane/methanol mixtures[4.23]. Authors reported

  Figure 4.12. The binary mixture influence on the PHP performance: (a) methanol-water and (b) acetone-water mixtures (D = 2 mm, FR = 50%, BHM) [4.25].

Figure 4 .

 4 Figure 4.13. Definition of self-rewetting fluid [4.27].

  Figure 4.16. (a) Surfactant concentration influence on the measured equilibrium surface tension and (b) variation of the measured dynamic viscosity versus temperature [4.36].

  [4.39]. Therefore, it is necessary to perform experiments directly in microgravity conditions.As discussed in the Chapter 2, heat pipes with capillary structure are widely used in space applications since the second part of 20 th century. Despite their proven operation, the thermal performance of conventional heat pipes is severely degraded in weightless environment [4.40] compared with ground conditions.

Figure 4 .

 4 Figure 4.18. Temperature histories for FPPHP (1×1 mm 2 , R-114, FR = 50-70%, Tcond = 10 o C) tested by Gu et al. [4.41].

Figure 4 .

 4 Figure 4.20. Cross-sectional view of the PHP with separating grooves [4.5].

  of the basic research studies on pulsating heat pipes operation under different conditions, i.e. geometrical parameters, orientations, working fluids, and gravity, concerning the device thermal behavior has been done in present chapter. Currently, pulsating heat pipes are considered as a very promising thermal management technology. Since first device introduction byAkachi [4.10], numerous experimental and theoretical studies have focused on the influence of different parameters on pulsating heat pipe operation and performance: channel shape and diameter, number of turns, device orientation, filling ratio, working fluid thermophysical properties and even microgravity. Some authors focused on the fluid motion inside the PHP to better understand the physics of the liquid-vapor pair motions.

  and scientific tests performing, as well drop tower, parabolic flights, sounding rockets or experiments on-board satellites and the International Space Station (ISS). The comparison of these approaches in the context of the gravity level vs. duration is resumed in Fig. 5.1. The parabolic flights have been chosen as a compromise between duration (around 20 seconds per parabola), accessibility (few campaigns per year) and opportunity to perform controlled experiments: researchers are able to interact with the experimental system and change parameters during experiments if necessary. Furthermore, in comparison with other "ground-based" microgravity simulation environment (drop tower and sounding rocket), parabolic flights do not impose very strong limitations related to the device mass and volume, so experiments could be more flexible.

Figure 5 . 1 .

 51 Figure 5.1. Available reduced gravity platforms for scientific approaches [5.1].

Figure 5 . 2 .

 52 Figure 5.2. Parabolic flight trajectory [5.2].

FPPHPsFigure 5 . 3 .

 53 Figure 5.3. Parabolic flight experimental system: (a) top view of the test rack and (b) command rack.

  and 74 th ESA Parabolic Flight Campaigns are used for the analysis in the present work, as well from ground tests for the FPPHPs used in 71 st and 74 th ESA PFCs. Because FPPHPs used during all these tests have different geometrical and fluid/solid thermophysical properties, it seems necessary to present each test section in chronological order. 5.2.1.1. 64 th ESA Parabolic Flight Campaign Despite the fact of analyzing videos of two-phase flows inside only two out of four tested FPPHPs, an experimental investigation on four one-side transparent flat plate pulsating heat pipes with different channel diameters (1.5, 2, 2.5 and 3 mm) was performed during 64 th ESA PFC (only data for the FPPHPs with channel diameters of 1.5 mm and 3

Figure 5

 5 Figure 5.4. Copper-Borosilicate FPPHP used for 64 th ESA Parabolic Flight Campaign.

  , in the context of the ESA project INWIP and preparation of the forecast FPPHP tests onboard the ISS, the dimensions of the test cell have been modified respecting installation requirements -200 mm high and 80 mm wideprovided at the time by ESA specifications for integration in ISS Heat Transfer Host 1.

Figure 5

 5 Figure 5.5. Molybdenum-Sapphire FPPHP for 69th ESA Parabolic Flight Campaign.

Figure 5 . 6 .

 56 Figure 5.6. Copper-Copper FPPHP for 71 st and 74 th ESA Parabolic Flight Campaigns.

Figure 5

 5 Figure 5.7. Cross-section cut of a welded flat plate PHP sample.

  5.9. All three components are connected by a Swagelok® valve to a T-connection and are used respecting the manipulation priority. The first step is to evacuate residual liquid and/or gases from the device, and create the vacuum inside it with the rotary vacuum pump, connected to the FPPHP (opening valves V0 and V1). Once the device empty and vacuumed (the valves V0 and V1 are closed), leaks tests of all connections and PHP interfaces (both glued and soldered) have been performed thanks to Pfeiffer Adixen ASM 142 helium-sensible leaks detector (opening valves V0 and V2

Figure 5 . 8 .

 58 Figure 5.8. Validation of the filling fluid saturation pressure (ethanol).

Figure 5 .

 5 Figure 5.10. Parabolic flight cooling system description.

Figure 5 .

 5 Figure 5.11. Aluminum cooling plate with fans.

  performance characterization. As mentioned above, T-type thermocouples with accuracy of ±0.7 o C have been used for the temperature measurements (their installation schema for each FPPHP presented in Fig.5.12). A CompactRIO NI-9213 module have been used to record temperatures.On the other hand, one of the most important PHP operational parameter is the absolute pressure and fluctuations. First, internal pressure monitoring makes possible to control the leaks and non-condensable gas presence inside the PHP by comparing it to the saturation pressure function of temperature. Then, pressure measurements allow to have an idea of the fluid flow behavior inside the PHP. A GE ® PTX5076 pressure sensor (±200 Pa accuracy) is connected to connected to (or to one of) the central U-turn(s) of the evaporator zone (except for the 64 th ESA PFC FPPHP where the pressure sensor was connected to the central surrounding U-turn of the condenser zone) (see figure5.12 for positions). The sensor membrane was centered relatively to the middle of the device to prevent the liquid column pressure influence. Grabbing of the pressure values from the transducers were realized thanks to CompactRIO NI-9215 voltage measurement module.

  Figure 5.12. Temperature and pressure sensors placements in the FPPHPs tested during (a) 64 th , (b) 69 th and (c) 71 st and 74 th PFCs (blue and red lines correspond to condenser and evaporation sections, respectively).

  importance of the working fluid selection on the pulsating heat pipe operation has been mentioned in the previous chapter. This step is based on various parameters as well chemical compatibility with the device material, good thermal (high sensible and latent heat) and hydraulic (low viscosity and surface tension) properties ensuring high heat transfer rate and stable oscillatory flow. Also, due to peculiarities of this work related to the microgravity tests, low toxicity as well relatively low operational temperature and pressure are required. According to the Merit number proposed by [4.19] and shown in Fig. 4.11b (see Chapter 4), the best fluids for PHPs are toxic and/or have a high saturation pressure in the tested temperature range (20-100 o C) as well ammonia, R22 and R134a. These choices in this work have been made in favor of water, FC72, ethanol and methanol. Despite their low "ranking" related to other fluids, most of them correspond to the necessary requirements of design parameters, non-toxicity (except for methanol) and low pressure in the specified temperature interval. Moreover, these fluids have been widely used for electronic cooling (FC72), experimental studies (water and ethanolclassical fluids) and for space applications (methanol).

  functions of time and related to the heat load are shown in Fig.5.13. Evaporator temperatures are plotted in different shades of red, adiabatic ones are in orange, the pressure curve is plotted in grey and heat load is represented by the dashed green line. Left ordinate-axes for temperature curves have the same scale for every graphs (0-120 o C), while right one can vary depending on the pressure values (hPa or kPa) in order to improve graphics readability.It seems that Fig.5.13 can represent the main FPPHP operation conditions for major number of tested cases. For low heat load applied (50 W), the FPPHP filled with pure water operates in dry mode (liquid accumulates in the condenser zone). In these conditions, the FPPHP transfers heat mostly thanks to the device thermal conduction (discussed in the previous chapter). With heat load augmentation, evaporator (and adiabatic) temperatures increases with significant pressure fluctuations, which means short-term "out" from stopover operational mode. Sometimes this phenomenon is observed for the FPPHP tested with other fluids, in different position and heat load.

Figure 5 .

 5 Figure 5.13. Transient temperature and pressure responses of the copper FPPHP (also ESA 71 th PFC) related to the heat load increase (water, FR=50%, Tcool = 20°C, vertical BHM).

Figure 5 .

 5 Figure 5.14. Temperature and pressure histories for FPPHP: (left) original and (right) repeated tests (ethanol, FR=50%, Tcool =20 o C, vertical BHM)

  slightly increases probably due to evaporation of the residual liquid in the evaporator and vapor superheating. Sometimes, the evaporator temperature drops during parabola. At the same time, pressure and adiabatic temperature surges can be observed, which illustrates the sudden change of the FPPHP operational modetransition from stopover to oscillations, due to the two-phase fluid flow "re-activation".

Figure 5 .

 5 Figure 5.15. Transient temperature and pressure responses of the molybdenum FPPHP (used for ESA 69 th PFC) related to the gravity acceleration changes (ethanol, FR=40%, Q = 100 W, vertical BHM during normal gravity).

  described in this chapter. The prototypes have been developed for the tests during four ESA Parabolic Flight Campaigns, based on the security requirements and zone of scientific interestscoupled thermal and hydraulic studies of the FPPHPs under different during on-ground tests and parabolic flight campaigns (ESA 69th, 71st and 74th Parabolic Flight Campaigns) on flat plate pulsating heat pipe will be presented and analyzed in this chapter.

Figure 6 . 1 .Fig. 6 .

 616 Figure 6.1. Temperature and pressure responses for FPPHP filled with water (FR=50%; Tcool = 20 o C, vertical BHM).

Figure 6 . 2 .

 62 Figure 6.2. Temperature and pressure responses for FPPHP filled with ethanol (FR=50%; Tcool = 20 o C, vertical BHM).

Figure 6 . 3 .

 63 Figure 6.3. Temperature and pressure responses for FPPHP filled with methanol (FR=50%; Tcool = 20 o C, vertical BHM).

(

  Fig. 6.4 represents a comparison of thermal resistances as functions of applied heat power for FPPHP filled with water, ethanol and methanol, with filling ratio of FR = 50% at range of heat loads of 50 W -200 W and condenser temperatures of 20 o C and 40 o C (Fig.A.1.6-A.1.8). Here, higher values of thermal resistance mean lower thermal performances and, so, lower values correspond to better heat transfer characteristics of FPPHP.

6 . 3 .

 63 Lower thermal resistance value at 50 W of heat load is observed for FPPHP filled with ethanol. Following heat load increase leads to thermal resistances decrease for all tested fluids at Tcool = 20 o C. Then, their convergence almost at the same point (R = 0.07 K/W) for 200 W of applied power is observed.

Figure 6 .

 6 Figure 6.4. Thermal resistances for FPPHP tested at Tcool = 20 o C and Tcool = 40 o C of cooling water temperature (FR = 50 %, vertical BHM).

Figure 6 .

 6 Figure 6.5. Schematic representation of the liquid level inside FPPHP (from the top condenser) for filling ratio of FR = 50% and FR = 59%.

Fig. 6 .

 6 Fig. 6.6 shows the comparison of thermal resistances for flat plate pulsating heat pipe

Figure 6 . 6 .

 66 Figure 6.6. Effect of filling ratio on thermal resistances of FPPHP filled with water (vertical BHM).

  high influence on flow resistance in mini-and microchannels. In addition, wettability has influence on liquid film formation close to the liquid plugs and can affect heat transfer performances.

Fig. 6 .

 6 Fig. A.1.1, A.1.2, A.1.6, A.1.9 and A.1.10).

Figure 6 . 7 . 6 -

 676 Figure 6.7. Thermal resistances for FPPHP filled with ethanol and methanol aqueous solutions at Tcool = 20 o C and Tcool = 40 o C (FR = 50%, vertical BHM).

Figure 6 .

 6 Figure 6.8. Thermal resistances for FPPHP filled with SRWFs at Tcool = 20 o C and Tcool = 40 o C (FR = 50%, vertical BHM).

Figure 6

 6 Figure 6.9. Thermal resistances for FPPHP filled with Tween 40 © aqueous solution at condenser temperature of Tcool = 20 o C and Tcool = 40 o C (FR = 50%, vertical BHM).

Figure 6 .

 6 Figure 6.10. Temperature and pressure responses for FPPHP filled with water and NCG (FR = 50%; Psys = 15 kPa; Tcool = 20 o C, vertical BHM).

Fig. 6 .

 6 Fig.6.10 presents temperature and pressure behavior for FPPHP filled with NCG presence. The regular and high amplitude temperature oscillations can be seen starting from 100 W of heat load (low amplitude oscillations present at 50 W also), contrary case with pure water (Fig.6.1), where temperature increases significantly during period for 50 W of applied power. Moreover, great oscillations of pressure values testifies the alternate presence of liquid and gaseous phases in the evaporation zone (near to the pressure sensor probe),

  3.11c). During the beginning of heat power step of 150 W, evaporator temperature exceeds have been stopped. Anyway, presented cases could characterize an overall behavior of FPPHP operating with water in horizontal position.

Figure 6 .

 6 Figure 6.11. Temperature and pressure responses for FPPHP filled with water (FR = 50%; Tcool = 20 o C, horizontal).

  can be observed. Sometimes (especially for highest heat loads) higher temperature rises with following drops have been registeredthis means a short-term periods of flow oscillations stopover.

Figure 6 .

 6 Figure 6.12. Temperature and pressure responses for FPPHP filled with ethanol (FR = 50%; Tcool = 20 o C, horizontal).

Figure 6 .

 6 Figure 6.13. Temperature and pressure responses for FPPHP filled with methanol (FR = 50%; Tcool = 20 o C, horizontal).

Figure 6 .

 6 Figure 6.14. Thermal resistances for FPPHP tested at Tcool = 20 o C and Tcool = 40 o C of cooling water temperature (FR=50%, horizontal).

Figure 6 .Figure 6 .

 66 Figure 6.15. Effect of filling ratio on thermal resistances of FPPHP filled with water (horizontal).

Figure 6 .

 6 Figure 6.17. Thermal resistances for FPPHP filled with SRWFs at condenser temperature of Tcool = 20 o C and Tcool = 40 o C (FR=50%, horizontal).

  A2.4 and A.2.12 -Appendix A.2could be seen that, in horizontal position, FPPHP filled with self-rewetting fluid(2-butanol) operates in optimal mode with high frequent and low amplitude oscillations for both temperature levels of condenser cooling fluid. Heat load augmentation provokes overall non-significant evaporator temperature rise keeping the same operational mode. These evaporator temperature increases lead to smooth decrease of thermal resistance values.Operation behavior and evaporator temperature evolution for the FPPHP filled with water/surfactant solution are almost the same as the case with pure water at 50 W of heat load and at cooling fluid temperature of 20 o C (Fig. A.2.5 -Appendix A.2). But, contrary to the case of water-filled device, where heat load augmentation leads to total operation failure, startup is reached at 50 W of applied power for water/surfactant solution. Moreover, stable temperature oscillations have been observed only for 100 W of heat load (Fig. A.2.13 -Appendix A.2). Heat load increase up to 150 W led to temperature increase exceeding 100 o C.

Figure 6 .

 6 Figure 6.18. Thermal resistances for FPPHP filled with Tween 40 © aqueous solution at condenser temperature of Tcool = 20 o C and Tcool = 40 o C (FR = 50%, horizontal).

Figure 6 .

 6 Figure 6.19. Temperature and pressure responses for FPPHP filled with water and NCG (FR=50%; Psys = 15 kPa; Tcool = 20 o C, horizontal).

Fig. 6 .

 6 20 presents typical temperature and pressure behavior for the FPPHP tested during five subsequent parabolas for the molybdenum-sapphire FPPHP filled with ethanol (ESA PFC 69). Although only one case is presented, the main behavior observed for microgravity conditions are representative of all other levels of heat power applied and all other tested fluids. Figures for all other tests conditions could be found in Appendix B (Fig. B.1.1 to B.3.3). Here will be presented mostly qualitative analysis of the FPPHP behavior for tests conducted during parabolic flights with explanation of the main related phenomena.

Figure 6 .

 6 Figure 6.20. Transient temperatures, pressure and acceleration responses during a series of 5 successive parabolas for molybdenum-sapphire (Mo/Sa) FPPHP at heat load of 100 W (ethanol, FR = 40%, Tcool = ~25 o C).

Figure 6 .

 6 Figure 6.21. Flow pattern for FPPHP for (a) normal gravity and (b) microgravity.

Figure 6 . 22 .

 622 Figure 6.22. Temperature behavior of FPPHP filled with water (Co/Co, Q=50 W, Tcool = ~25 o C, ESA PFC 74).

  Temperature augmentation, provoked by liquid accumulation in the condenser zone and by evaporator dry-out (explained in the beginning of this section), have been observed for all parabolas. Device operation reactivation phases, followed by temperature sudden decrease , have been observed for first, fourth and fifth parabolas. Higher heat load does not almost influence temperature behavior during normal gravity (but with augmentation of overall evaporator temperature) and microgravity period are characterized by higher temperature peaks and rare reactivations (parabolas 1, 2 and 5 in Fig.6.30).

Figure 6 .

 6 Figure 6.23. Temperature behavior of FPPHP filled with water/16% ethanol (Co/Co, Q = 50 W, Tcool = ~25 o C, ESA PFC 74).

  Also, pressure oscillations during microgravity phases seem to confirm these trends.Higher heat load level of 150 W leads to small temperature oscillations amplitude augmentation and overall evaporator temperature rise (as it is the case with water) during normal gravity. Note that despite significant evaporator temperature augmentation (normally states about dry-out) during microgravity periods, regular reactivation have been observed for all parabolas (Fig.)it could be clearly seen with pressure oscillations, which mean oscillatory flow inside FPPHP.

Figure 6 . 24 .

 624 Figure 6.24. Temperature behavior of FPPHP filled water/28% methanol (Co/Co, Q=50 W, Tcool= ~25 o C, ESA PFC 74).

Finally

  , last tested fluid is 5% 1-butanol aqueous solution, being a self-rewetting fluid.During ground tests, water/1-butanol filled FPPHP have shown pretty good performances

Figure 6 .

 6 Figure 6.25 Temperature behavior of FPPHP filled water -5% 1-Butanol (Co/Co, Q=50 W, Tcool = ~25 o C, ESA PFC 71).

Figure 6 . 26 .

 626 Figure 6.26. Test results for the FPPHP filled with deionized water (Co/Sa, Q = 100 W, Tcool = ~27 o C, ESA 71 st PFC).

Figure 6 . 27 .

 627 Figure 6.27. Test results for the FPPHP filled with water and NCG (Co/Sa, 100 W, Tcool = ~28 o C, ESA 71 st PFC).

Figure 6 .

 6 Figure 6.28. Temperature overshoots during parabolas.

  ∆𝑇 show an unstable behavior but with much smaller values and distribution

Figure 6 . 29 .

 629 Figure 6.29. Pressure and temperature instabilities during two successive parabolas at 100 W for FPPHP filled with (a) DI water and (b) water -NCG.

Fig. 6 .

 6 Fig.6.29b presents pressure and temperature oscillations for water and NCG during the same 3 rd and 4 th parabolas. No significant changes in temperature and pressure are noticeable for the three gravity levels during first parabola, which means continuous slug/plug flow motion. Temperature rise, such as pressure increase, can be observed during the second parabola. However, unlike case with pure water, a significant pressure surge happens before end of parabola, corresponding to a change of flow pattern.

Figure 6 .

 6 Figure 6.30. Transient temperatures, pressure and acceleration responses during a series of 5 successive parabolas for copper FPPHP at heat load of 150 W (water, microgravity, ESA PFC 74).

  parabolas and, probably, can be explained by absence of full stopover present gravity force assists the liquid reflows back to evaporator.

Figure 6 . 31 .

 631 Figure 6.31. Transient temperatures, pressure and acceleration responses during a series of 5 successive parabolas for copper FPPHP at heat load of 150 W (water, Lunar gravity, ESA PFC 74).

  Experimental results obtained for FPPHPs filled with different fluids and tested during parabolic flights have been analyzed in this section. During microgravity periods, temperature augmentation of evaporator temperature was observed for all tested cases.During most of parabolas, evaporator temperature augmentation was accompanied by a

Figure 6 .

 6 Figure 6.32. Transient temperatures, pressure and acceleration responses during a series of 5 successive parabolas for copper FPPHP at heat load of 150 W (water, Martian gravity, ESA PFC 74).

Finally, tests for

  three different levels of gravity (micro, Lunar and Martian) shown evaporator temperature increase during lunar gravity phases, but with frequent flow reactivations (unlike microgravity); and any difference for FPPHP operating behavior for Martian gravity, compared to the terrestrial.In this chapter, influence of parameters on FPPHP operation, as well as orientation, cold source temperature, filling ratio, fluid thermodynamic properties and microgravity level, have been studied.Main conclusions are the following:Concerning ground tests in vertical position, any significant difference in operation behavior of FPPHP for all tested pure fluids has not been found (except different operational temperature ranges, and observed instabilities for device filled with 1-butanol and Tween 40 aqueous solutions). If for 50 W of applied power, thermal resistances greatly differ, no significant changes have not been observed for higher heat loads. Filling ratio augmentation has not a significant effect on FPPHP thermal performances.Operation behavior of water-filled FPPHP in horizontal position is very different from vertical orientation, especially for low cold source temperature. If for vertical position, no start-up has been reached just for 50 W of applied power, FPPHP in horizontal position operates in "dry" mode for 50 W and 100 W of heat loads. Total system failure has been observed with heat load augmentation. On the contrary, other tested fluids show high amplitude and frequent evaporator temperature oscillations. Condenser temperature augmentation leads to start-up for water-filled FPPHP and temperature oscillations homogenization for heat load range from 100 W to 200 W for mixtures. Higher filling ratio also leads to stable oscillatory regime establishment for FPPHP tested at both temperatures of cooling liquid.

  Figure 7.1. Transient operation during microgravity phase (ethanol; 3 mm; 69 th ESA PFC): (a) temperatures profiles of flow reactivation during the 5 th parabola of Fig. B.1.1; and (b) flow patterns one second before (left) and one second after fluid flow reactivation (right).

  Fig. 7.2 are shown the temperature profile along a single channel (corresponding to the yellow line) on which the significant temperature drop, corresponding to the vapor-liquid interface, can easily be detected.

Figure 7 . 2 .

 72 Figure 7.2. Infrared post-processing procedure on FlirIR software® (darker lines, and lighter lines on the left image correspond to the vapor and the liquid phase, respectively) [7.10].

Figure 7 . 3 .

 73 Figure 7.3. A cropped greyscale image of channel with vapor-liquid interfaces.

Figure 7 . 4 .

 74 Figure 7.4. Definition of the menisci position: channel with menisci (images from [7.13]).

Figure 7 . 5 .

 75 Figure 7.5. Developed tracking approach validation.

Figure 7 . 6 .

 76 Figure 7.6. Representing liquid plugs and channels numbering (ethanol; 3 mm; 69 th ESA PFC).

  Figure 7.7. Displacement of the vapor/liquid menisci for two central liquid plugs during microgravity period from dry-out to re-activation phase (ethanol; 3 mm; 69 th ESA PFC): (a) Fig. 6.20, third parabola (100 W); (b) Fig B.1.1, first parabola (150W).

  Figure 7.8. Transient velocities of two following liquid plugs menisci (Fig 6.20, parabola 1) (ethanol; 3 mm; 69 th ESA PFC).

Figure 7 . 9 .

 79 Figure 7.9. Channel distribution of Vtrans. (65 th and 69 th ESA PFCs)

7. 8 )

 8 during reactivation phases, are plotted in Fig.7.9. In most cases, maximum oscillation velocities correspond to the transition velocities just before the flow pattern transition/menisci deformation. The average transition velocity for all presented conditions

  confined flow prediction: Weber (𝑊𝑒 = 𝜌 𝑙 𝑈 𝑙 2 𝐷 𝜎 ⁄ ) and Garimella (𝐺𝑎 = 𝑅𝑒√𝐵𝑜) numbers, as functions of the channel number are presented in Fig. 7.10 and Fig. 7.11, respectively. First observations show a distribution of Weber number ranging from 1 to 25, but most of the points (about 80%) for ethanol-filled FPPHP are gathered in the zone between 2 and 5, very close to the reference value of 4. The average value is 3. Value of the Weber number for FPPHP filled with FC72 and channel diameter of 1.5 mm varies in the range of 0.1 to 11 with concentration of the points majority between the values of 1.5 and 2.5 (with an average value of 2). For the device filled with FC72 and channel diameter of 3 mm, Weber number values are mostly in the range of 4-8 withan average value of 6. All these values are quite

Figure 7 .

 7 Figure 7.10. Weber number based on the transition velocity (65 th and 69 th ESA PFCs).

Figure 7 .

 7 Figure 7.11. Garimella number based on the transition velocity (65 th and 69 th ESA PFCs).

  in conventional non-dimensional numbers) related to the gravity acceleration term g have been replaced by the acceleration term of liquid plugs (dul/dt). The modified numbers are calculated as follows:

Figure 7 .

 7 Figure 7.12. Flow map from [7.18] and adapted for the flow transition under microgravity for FPPHP with rectangular channels.

  Fig. 7.14.

  Fig. 7.14d, e and f). Comparing presented cases, the acceleration and driving forces values on the transition line are very close from case to case. The low influence of the frictional part on the force balance and very close (almost equal) behavior of the accelerating and driving forces corresponding well to the results obtained in [7.21].

  Figure 7.14. Forces influencing the liquid plugs motion during period prior to reactivation phases (ethanol; 3 mm; 69 th ESA PFC): for 100 W (a) channels 1-14; (b) channels 6-7; (c) channels 8-9; and for 150 W (d) channels 1-14 ; (e) channels 6-7; (f) channels 8-9.

  Figure 7.15. Comparison of driving forces values for two central liquid plugs (from Fig. 7, ethanol; 3 mm; 69 th ESA PFC): (a) 100 W and (b) 150 W.

  during some parabolas, stop-over periods occur and are sometimes interrupted by a flow pattern transition from slug flow to annular flow with uniform liquid distribution inside pulsating heat pipe, also called re-activation phases. The motion of liquid/vapor interfaces, as well as the force influence on fluid flow and flow pattern change, have been analyzed. Despite the different amplitude and oscillation frequencies during microgravity start and stopover periods, the values of transition velocities appear to be very close to each other. Frequently used non-dimensional criteria, based on Weber and Garimella numbers, have been evaluated for tested FPPHP under microgravity conditions. Experimentally obtained average values of We number for flow transition are very close to the reference critical value, whereas average values of Ga number are one-two orders of magnitude higher than reference critical value. Adaptation of the flow map correlations (based on modified Bond, Weber and Froude numbers including the actual fluid acceleration than a constant gravity acceleration value) developed in [7.18] placed experimentally obtained in this work transitional points in the zone of slug-plug dominant flow pattern, nearly to the coalescence limit and zone of slugplug to annular transition. This work refers to few specific cases, and future complex studies are necessary to enhance the characterization of flow behavior and flow transition inside of PHP under both ground and microgravity conditions. The quantitative analysis of the flow transition mechanisms and forces influencing them can help in design of the stable functioning pulsating heat pipe without stopover phenomena and give the necessary data for the modelling.

  Analysis of the liquid-vapor interfaces motions has shown very close values of the flow transition velocities for each studied PHP/fluid pair. No dependence of velocity on applied heat load has been found.  Non-dimensional criteria, based on Weber and Garimella numbers, have been evaluated using experimentally determined velocities. Obtained We numbers are very close to the reference critical value but average values of Ga number are one-two orders of magnitude higher than critical one.  Finally, proposed in [8.1] new flow pattern map with modified Bond and Froude numbers, including the actual fluid acceleration instead of a constant gravity acceleration value has been used to analyze the transitional points obtained in actual work. Their placement in this fluid properties independent map marks the distribution of the points in the zone of slug-plug flow dominant, nearly to the coalescence limit and zone of slug-plug to annular flow transition.Generally speaking, this work gave some answer related, in particular, to the operational performances of the flat plate pulsating heat pipe under different working conditions, as well influence of filling liquid properties. Moreover, the visual analysis of the two-phase fluid flow and its transition can be used in future modelling tools development resulting to the engineering design approaches. Finally, presented results and analysis shed the light on further researches, proposed in the next section.8.2. PerspectivesFirst, the accurate infrared measurements of the liquid bulk temperature should be realized in the context of the local heat transfer coefficient determination and liquid-vapor phase's separation detection (coupled with high-speed visualization).Working on inverse heat conduction problem has already started in collaboration with partners from University of Parma. Knowing that external side of the FPPHP is affected by the flow temperature inside the device, it became possible to determine the wall-to-fluid heat fluxes using inverse heat conduction problem resolution approach, previously validated by numerical simulations. This methodology can provide a better insight into the oscillatory typical flow for such heat transfer devices (Fig.8.1).

Figure 8 . 1 .

 81 Figure 8.1. The evaluated heat flux adiabatic zone (ethanol, FR = 50 %, horizontal, 200 W) [8.2].

Figure A. 1

 1 Figure A.1.1. Temperature and pressure responses for FPPHP filled with 20 % ethanol/water solution (FR = 50%, Tcool = 20 o C, vertical BHM).

Figure A.1. 2 . 3 .

 23 Figure A.1.2. Temperature and pressure responses for FPPHP filled with 28 % methanol/water solution (FR = 50%, Tcool = 20 o C, vertical BHM).

Figure A. 1

 1 Figure A.1.4. Temperature and pressure responses for FPPHP filled with 5 % 2butanol/water solution (FR = 50%, Tcool = 20 o C, vertical BHM).

Figure A.1. 6 .

 6 Figure A.1.6. Temperature and pressure responses for FPPHP filled with water (FR = 50%, Tcool = 40 o C, vertical BHM).

Figure

  Figure A.1.8. Temperature and pressure responses for FPPHP filled with methanol (FR = 50%, Tcool = 40 o C, vertical BHM).

Figure

  Figure A.1.10. Temperature and pressure responses for FPPHP filled with 28% methanol/water solution (FR =50 %, Tcool = 40 o C, vertical BHM).

Figure

  Figure A.1.12. Temperature and pressure responses for FPPHP filled with 5 % 2butanol/water solution (FR = 50%, Tcool = 40 o C, BHM).

Figure

  Figure A.1.14. Temperature and pressure responses for FPPHP filled with water (FR = 59%, Tcool = 20 o C, vertical BHM).

Figure

  Figure A.1.16. Temperature and pressure responses for FPPHP filled with water + NCG (FR = 50%, Tcool = 40 o C, vertical BHM).

Figure A. 2 . 2 . 3 .

 223 Figure A.2.2. Temperature and pressure responses for FPPHP filled with 28 % methanol/water solution (FR = 50%, Tcool = 20 o C, horizontal).

Figure A. 2

 2 Figure A.2.4. Temperature and pressure responses for FPPHP filled with 5 % 2butanol/water solution (FR = 50%, Tcool = 20 o C, horizontal).

Figure A. 2 . 6 .

 26 Figure A.2.6. Temperature and pressure responses for FPPHP filled with water (FR =50 %, Tcool = 40 o C, horizontal).

Figure A. 2 .

 2 Figure A.2.78 Temperature and pressure responses for FPPHP filled with methanol (FR = 50%, Tcool = 40 o C, horizontal).

Figure A. 2 .

 2 Figure A.2.10. Temperature and pressure responses for FPPHP filled with 28 % methanol/water solution (FR=50%, Tcool = 40 o C, horizontal).

Figure A. 2 .

 2 Figure A.2.12. Temperature and pressure responses for FPPHP filled with 5 % 2butanol/water solution (FR=50%, Tcool = 40 o C, horizontal).

Figure A. 2 .

 2 Figure A.2.14. Temperature and pressure responses for FPPHP filled with water (FR = 59%, Tcool = 20 o C, horizontal).

Figure A. 2 .

 2 Figure A.2.16. Temperature and pressure responses for FPPHP filled with water + NCG (FR=59%, Tcool = 40 o C, horizontal).

Figure B. 2 . 2 .

 22 Figure B.2.2. Temperature and pressure responses for FPPHP filled with 5% 1butanol/water solution (Co/Co, FR = 50%, Q = 150 W, microgravity).

3 .

 3 Figure B.2.3. Temperature and pressure responses for FPPHP filled with water (Co/Sa, FR = 50%, Q = 50 W, microgravity).

Figure B. 2 . 4 .

 24 Figure B.2.4. Temperature and pressure responses for FPPHP filled with water (Co/Sa, FR = 50%, Q = 150 W, microgravity).

Figure B. 2 . 6 .

 26 Figure B.2.6. Temperature and pressure responses for FPPHP filled with water + NCG(Co/Sa, FR = 50%, Q = 150 W, microgravity).

  Figure B.3.1. Temperature and pressure responses for FPPHP filled with water (Co/Co, FR = 50%, Q = 50 W, Lunar gravity).

Figure B. 3 . 2 .

 32 Figure B.3.2. Temperature and pressure responses for FPPHP filled with water (Co/Co, FR = 50%, Q = 50 W, Martian gravity).

3 .

 3 Figure B.3.3. Temperature and pressure responses for FPPHP filled with 20% ethanol/water solution (Co/Co, FR = 50%, Q = 150 W, microgravity).

Figure B. 3 . 3 .

 33 Figure B.3.3. Temperature and pressure responses for FPPHP filled with 28% methanol/water solution (Co/Co, FR = 50%, Q = 150 W, microgravity).

3 4 5 6 7 8

 345678 Le deuxième chapitre présente une revue des différents systèmes de refroidissement. Il n'existe pas de classification universelle des systèmes de contrôle thermique du fait de la complexité des différents principes de fonctionnement ainsi que de leur grand nombre. C'est pourquoi ici seuls les systèmes diphasiques sont présentés en tant que solutions de contrôle thermique parmi les plus efficaces et largement utilisées pour les applications terrestres et spatiales. Les systèmes diphasiques peuvent être divisés entre systèmes actifs et passifs. Dans les systèmes actifs, l'utilisation d'une pompe est nécessaire pour forcer le mouvement du fluide caloporteur dans la boucle. En revanche, le mouvement du liquide dans les systèmes passifs est assuré par la gravité ou la capillarité. Dans les systèmes actifs, la pompe doit être complétée par un vase d'expansion permettant l'amortissement des instabilités éventuelles liées aux dilatations thermiques du fluide. Malgré un fonctionnement efficace, les composants supplémentaires et la consommation d'énergie entraînent une baisse de la fiabilité de ces systèmes. En revanche, les sous-systèmes visant la fiabilité globale sont exclus dans les systèmes de refroidissement passifs, mais un autre moyen de mouvement de fluide est nécessairele fonctionnement des caloducs, LHP et CPL est permis grâce à la présence d'une structure capillaire ; les thermosiphons fonctionnant quant à eux grâce à la gravité (donc pas utilisables pour les applications spatiales malgré son efficacité). Les caloducs présentent un avantage certain de par leur simplicité par rapport aux boucles diphasiques mais leur utilisation reste limitée à cause de leurs basses performances en termes de transfert thermique, elles-mêmes liées aux limites de transport du fluide par la structure capillaire. Une autre technologie récente dans la famille des caloducs est le caloduc oscillant. Dans ces systèmes, le transfert de chaleur par voies latente et sensible couplé aux transferts de masse provoqué par une différence de pressions entre les zones de condensation et celles d'évaporation, ce qui entraîne la mise en mouvement global du fluide caloporteur. Bien qu'offrant des performances limitées en termes de densité de flux par rapport à d'autres technologies, les caloducs oscillants ont l'avantage d'être des systèmes simples à réaliser et pouvant fonctionner sous différentes conditions de gravité. Chapitre Les principes de fonctionnement fondamentaux des caloducs oscillants, les particularités de ces dispositifs lorsqu'ils sont fabriqués à partir de plaques planes usinées, et leurs différences ainsi que les principaux paramètres d'influence sont abordés dans le troisième chapitre. Malgré différentes géométries, le point commun de l'ensemble de ces caloducs est la présence d'une zone d'évaporation au contact d'une source chaude, d'une zone de condensation (source froide) et d'une partie adiabatique entre ces deux zones. Malgré le fait que les mécanismes physiques du fonctionnement des caloducs oscillants conventionnels (tubulaires) sont bien connus mais encore peu maîtrisés, des phénomènes particuliers et contradictoires se produisent dans les dispositifs à plaques planes (caloducs oscillants plats) en raison du lien thermique entre les canaux et de leur forme spécifique (la présence d'angles vifs conduit à une distribution non uniforme du film liquide lié au déséquilibre de pressions capillaires entre les coins et les bords de la section transversale). Le fonctionnement d'un caloduc oscillant dépend d'une combinaison complexe entre différents paramètres, majoritairement couplés les uns aux autres. Un bon fonctionnement est obtenu lorsqu'un écoulement dit bulles/bouchons (slug flow) parcourt les différentes boucles du caloduc. Les transferts de chaleur se font à la fois par chaleur sensible dans les bouchons de liquide et par chaleur latente au niveau de l'évaporation ou de la condensation des films de liquide entourant les bulles de vapeur. Ces transferts sont sensiblement différents dans les caloducs plats avec une section carrée ou rectangulaire à cause de la présence des films liquides plus épais dans les coins et des liens thermiques entre les différents canaux. Aussi, le faible niveau de connaissance de ces systèmes ne permet pas encore de faire un dimensionnement pour un prototype industriel. Chapitre Une analyse bibliographique des différentes expériences réalisées sur les caloducs oscillants tubulaires et plats est détaillée afin d'évaluer l'influence des différents paramètres dans le quatrième chapitre. Pour des puissances moyennes à élevées, les performances des caloducs oscillants plats ayant des canaux à section carrée sont meilleures que pour ceux à section circulaire. L'orientation joue également un rôle important. La configuration verticale ou inclinée avec évaporateur en bas est plus favorable que la configuration horizontale. Une faible longueur adiabatique entre l'évaporateur et le condenseur tend à diminuer la résistance thermique. Il en est de même avec l'augmentation du nombre de boucles. Le taux de remplissage doit aussi être optimisé pour favoriser les oscillations, éviter des frottements trop importants, mais aussi l'assèchement de l'évaporateur en cas de quantité trop faible. Les performances thermiques dépendent également des propriétés thermophysiques des fluidse et sont inversement proportionnelles à la pente de la courbe pression/température de saturation. Les mélanges binaires ont aussi largement été utilisés et certains comme le mélange eau-méthanol permettent d'améliorer les performances. L'amélioration de la mouillabilité des fluides en mélangeant l'eau à des alcools ou en utilisant des surfactants permet de réduire la formation de zones sèches au niveau de l'évaporateur. Plusieurs études ont été menées sur des caloducs dans des conditions de microgravité. Généralement les performances sont similaires en gravité normale ou microgravité en position horizontale si la taille des canaux est inférieure à la longueur capillaire. Les performances sont un peu réduites en microgravité par comparaison à la gravité normale en configuration verticale. Chapitre Le système expérimental, destiné aux expériences au sol et en microgravité, et permettant à la fois la caractérisation qualitative et quantitative des caloducs oscillants plats, est décrit dans le cinquième chapitre. Les prototypes ont été développés pour les expériences au sol et au cours de quatre campagnes de vols paraboliques de l'Agence Spatiale Européenne, sur la base des exigences de sécurité et de la zone d'intérêt scientifique -études couplées thermiques et hydrauliques des caloducs oscillants plats testés sous différentes conditions de gravité. Les caloducs testés sont soit ren cuivre, soit en molybdène et certains sont recouvert d'une plaque en saphir permettant les visualisations complètes de l'écoulement dans les canaux et des mesures de température par thermographie infrarouge. Le système expérimental est équipé de thermocouples de type T permettant de mesurer les températures au niveau de l'évaporateur, de la zone adiabatique et du fluide de refroidissement secondaire. Un capteur de pression est également positionné soit en partie basse au niveau de l'évaporateur, soit en partie haute au niveau du condenseur. Une estimation des incertitudes des mesures est présentée, ainsi que l'analyse de répétabilité, qui démontre également la fiabilité du système Chapitre Les résultats obtenus par les études expérimentales, s'intéressant à l'influence des différents paramètres (comme par exemple, la nature du fluide, le taux de remplissage, la présence de gaz incondensables, la température de la source froide, le positionnement du caloduc et le niveau de gravite) sont présentés dans le chapitre 6. Pour la plupart des essais, les évolutions temporelles des températures au niveau de l'évaporateur, de la zone adiabatique, du condenseur et de la pression sont présentées. Enfin, une notion importante est mise en évidence et représente un critère pour l'évaluation de performance des caloducs : la résistance thermique globale. Donc chaque configuration de caloduc a été évaluée et comparée grâce aux calculs de ces résistances thermiques en plus des évolutions temporelles des températures et pression. Lors d'essais au sol, une étude paramétrique du caloduc oscillant rempli par des fluides de différentes natures a été menée. En position verticale, le caloduc oscillant fonctionne presque tout le temps comme un réseau de thermosiphons interconnectés. En orientation horizontale, le système fonctionne en mode pulsé puis le démarrage est atteint (généralement à 100 W de puissance appliquée), sauf lorsqu'il est rempli d'eau et testé à 20 o C de source froide : il présente de moins bonnes performances thermiques dans ces conditions. Les mélanges aqueux d'alcool utilisés comme fluides de travail conduisent à une très nette amélioration des performances thermiques, par rapport aux fluides purs (en particulier l'eau). L'augmentation de la température de condenseur entraîne également une amélioration des performances thermiques du FPPHP, en raison d'un glissement des propriétés thermophysiques dans le bon sens pour la plupart. Le système, rempli avec des solutions eau-surfactant, a montré un fonctionnement très instable malgré des résistances thermiques plus faibles comparées à l'eau pure. En revanche, il est aussi démontré que la présence des gaz non condensables dans le système, bien qu'entraînant une dégradation globale des performances thermiques du système, provoque un fonctionnement oscillatoire plus stable du fluide dans le caloduc testé.En effet, la présence de gaz incondensables provoque non seulement l'augmentation globale de la température du système en fonctionnement, mais induit également un fonctionnement oscillatoire stable du fluide aussi bien pendant les phases de gravité normales que de microgravité, avec une augmentation de température (en microgravité) bien inférieure à celle observée avec l'eau pure.Pour finir, une campagne de vols paraboliques a permis d'effectuer une série de tests sur trois niveaux de gravité différents (microgravité, gravités lunaire et martienne). Les résultats ont montré une augmentation de la température de l'évaporateur pendant les phases de gravité lunaire, mais avec des réactivations de fluide fréquentes (contrairement aux phases de microgravité). En revanche, pour la gravité martienne, les performances obtenues restent tout à fait similaires à celles observées sous gravité terrestre, démontrant un niveau seuil d'influence de la gravité sur le fonctionnement des caloducs oscillants lorsqu'ils sont en position verticale favorable.ChapitreLe septième chapitre est dédié aux visualisations des écoulement réalisées en phases de microgravité, afin d'analyser plus spécifiquement les conditions de redémarrage du régime bulles/bouchons observé pendant certaines paraboles (phases de microgravité). La méthodologie d'analyse de visualisation, basée sur la technique de suivi d'objets pour la détermination de la position des interfaces liquide-vapeur (ménisques) a été développée, validée et mise en oeuvre pour étudier le comportement de l'écoulement pendant les périodes de « stop-over » et de réactivations du fluide dans des conditions de microgravité. L'étude expérimentale de ces phénomènes à l'intérieur prototypes remplis avec de l'éthanol (diamètre des canaux : 3 mm) et FC72 (diamètres des canaux : 1,5 mm et 3 mm) et fonctionnant en microgravité est également présentée. Les périodes de microgravité s'accompagnent de phénomènes d'assèchement de l'évaporateur qui entraînent une dégradation globale des performances thermiques. Il a été constaté que, pendant certaines paraboles, des périodes d'arrêt se produisent et sont parfois interrompues par une transition du mode d'écoulement bulles/bouchons à un écoulement de type annulaire, avec une distribution uniforme du liquide à l'intérieur du caloduc oscillant, également appelées « phases de réactivation ». Le mouvement des interfaces liquide/vapeur, ainsi qu'un bilan des forces s'exerçant sur les bouchons liquides, et le changement de mode d'écoulement, ont été analysés. Malgré les différentes fréquences d'amplitude et d'oscillations pendant les périodes de « stop-over » et « startup » pendant les phases de microgravité, les valeurs des vitesses de transition semblent être très proches les unes des autres. Des critères basés sur des nombres adimensionnels fréquemment utilisés pour ce type d'écoulement, les nombres de Weber et de Garimella, ont été évalués pour les caloducs oscillants testés lors des phases de microgravité. Les valeurs moyennes obtenues expérimentalement du nombre de Weber pour la transition d'écoulement sont très proches de la valeur critique de référence (We,crit = 4), tandis que les valeurs moyennes du nombre de Garimella sont supérieures d'un à deux ordres de grandeur à la valeur critique de référence (Ga,crit = 160). L'adaptation des corrélations pour une tentative de cartographie d'écoulement (basée sur les nombres de Bond, de Weber et de Froude modifiés, incluant l'accélération réelle du fluide plutôt que l'accélération gravitationnelle constante), initialement développée par nos collègues de l'Université de Brighton, a permis de positionner les points de transition obtenus expérimentalement lors de ce travail dans la zone d'écoulement à dominante slugplug, presque à la limite des zones de transition par coalescence et de transition d'écoulement slug-plug à annulaire. Un bilan des forces entre l'accélération des bouchons liquides, les forces visqueuses et les forces motrices dues à la différence de pression entre deux bulles vapeurs de part et d'autre de ces bouchons a été effectué. La force visqueuse est estimée à l'aide d'une loi de Poiseuille en écoulement stationnaire, et la force motrice est ainsi identifiée à partir du bilan de quantité de mouvement par déduction des autres termes. La force visqueuse s'avère plus faible que la force de pression motrice calculée, cette dernière étant aussi supérieure d'un ordre de grandeur aux forces capillaires. Chapitre Cette thèse a été menée dans le cadre de deux projets successifs ESA MAP INWIP et ESA MAP TOPDESS, dédiés au développement des systèmes des transferts thermiques diphasiques et s'intéressant à la gestion thermique systèmes dissipatifs de haute puissance par caloducs oscillants, appliqués pour les vols hors gravité. Ce travail concerne les études expérimentales menés sur diffèrents caloducs oscillant dit plats testés sous différentes conditions opératoires incluant l'orientation, le taux de remplissage, la température de refroidissement, la nature du fluide, le niveau de puissance appliqué et le niveau de gravité.Notons que de nouvelles expériences au sol sont programmées sur une nouvelle géométrie de prototype destiné à être testé à bord de l'ISS à l'horizon 2024-2026. De nouveaux essais lors de campagnes de vols paraboliques seront également nécessaires pour valider le nouveau dispositif dans des conditions de microgravité.

Table 2

 2

	.1. Comparison the two-phase cooling solutions.
	Technology	Maximum Heat Flux	Total Power
		[W.cm -2 ]	[kW]
	Two-Phase Microchannel [2.30]	250	-
	Two-Phase Spray / Jet [2.30]	1820	-
	Thermosiphon / Loop Thermosiphon [2.31]	100	10
	Heat Pipe [2.32]	100	0.2
	Loop Heat Pipe [2.33]	900	10
	Capillary Pumped Loop [2.34]	100	10
	Pulsating Heat Pipe [2.35]	30	5

Table 6 .

 6 1. Test operating conditions for ground tests.

	Orientation	Fluid	Filling ratio	Applied power	Condenser temperature
	Vertical	Water	50%	0 -200 W	20 o C; 40 o C
			59%
		Ethanol	50%	0 -200 W
		Methanol	50%	0 -200 W
			59%
		Water -20% ethanol	50%	0 -200 W
		Water -28% methanol	50%	0 -200 W
		Water -5% butanol	50%	0 -200 W
		Water -0.5% Tween 40 50%	0 -200 W
	Horizontal	Water	50%	0 -100 W	20 o C; 40 o C
			59%	0 -150 W
		Ethanol	50%	0 -200 W
		Methanol	50%	0 -200 W	20 o C
				0 -150 W	40 o C
			59%	0 -200 W	20 o C; 40 o C
		Water -20% ethanol	50%	0 -200 W
		Water -28% methanol	50%	0 -200 W
		Water -5% butanol	50%	0 -200 W
		Water -0.5% Tween 40 50%	0 -150 W	20 o C
				0 -200 W	40 o C
	6.1.1. Results for the FPPHP tested in vertical BHM
	6.1.1.1. Qualitative analysis of the FPPHP filled with pure fluids
	Here, in this first subsection, evolution of temperatures and pressure curves for the

FPPHP varying heat load level will be studied. All tests were carried out for the heat load range from 50 W up to 200 W (if there is no system failure during tests) with 50 W-steps.

Typical temperature and pressure responses as functions of time for the FPPHP filled with Chapter 6 -Ground Based and Parabolic Flight Tests Analyses 123 pure water, ethanol and methanol in saturation state are presented in Fig.

6

.1 -6.3, respectivelly.

Table 6 .

 6

		ρl	μl	σl	λl	∂Psat/∂T
		(kg.m -3 )	(kg.m -1 .s -1 )	(N.m -1 )	(W.m -1 .K -1 )	(Pa.K -1 )
	Water	988	5×10 -4	0.068	0.604	609
	Water -5% 1-butanol	978	5.1×10 -4	0.034	0.595	605
	Water -5% 2-butanol	979	5.5×10 -4	0.032	0.564	619-639
	Water -20% Ethanol	943	5.2×10 -4	0.035	0.478	680
	Water -28% Methanol	943	4.9×10 -4	0.044	0.515	-
	Water -0.5% Tween 40 ©	988	5.4×10 -4	0.035	0.604	-
	Results for FPPHP tested with water, water/ethanol and water/methanol mixtures as
	working fluids, are shown in

2. Main thermophysical properties of the tested fluids at T = 20 °C

[6.9 -6.11

].

  Test parameters are presented in table 6.3. Note that only tests cases that could be exploited in this part are presented in this table. Finally, FPPHP filled with water, water with 16% of ethanol and with 28% of methanol (in weight) solutions have been tested under Lunar, Martian and micro-gravity conditions (all parameters are presented in table 6.4) during ESA 74th PFC.

	6.2.1. Experimental procedure Molybdenum FPPHP (N = 7) covered with sapphire transparent plate and filled with ethanol has been tested during ESA 69 th PFC. Several devices filled with water, water/5% 1-butanol mixture and water with presence of non-condensable gas (NCG) have been tested № Parabola PHP type Working fluid Heat Load Parabolic flight campaign 69 th 16-20 Mo/Sa Ethanol FR = 40 % 100 W 26-30 150 W Parabolic flight campaign 71 st (day 1) 6-10 Co/Sa Water + NCG FR = 50% 50 W 10-15 Co/Sa Water + NCG FR = 50% 100 W 16-20 Co/Sa Water + NCG FR = 50% 150 W Parabolic flight campaign 71 st (day 3) 6-10 Co/Co Co/Sa Water -5% 1-butanol FR = 50% Water FR = 50% 50 W 10-15 Co/Co Co/Sa Water -5% 1-butanol FR = 50% Water FR = 50% 100 W Co/Sa Water FR = 50% during ESA 71 st PFC. Table 6.3. Experimental matrix for microgravity tests (69th and 71st PFCs) 16-20 Co/Co Water -5% 1-butanol FR = 50% 150 W

th , 71 st and 74 th ESA Parabolic Flight Campaigns (PFC). Present section is focused on temperature and pressure signals analyses obtained during parabolic flight tests in the context of various gravity levels, heat load and working fluid thermodynamic properties influence on the FPPHP operation.

Table 6 .

 6 4. Experimental matrix for 74th PFC tests (copper-copper FPPHP)

	№ Parabola Working fluid	Heat Load Gravity level
	Day1
	0-5	Water FR=50%	50 W	Martian (0.38g)
	6-10			Lunar (0.18g)
	10-15			μ-gravity (0.01g)
	16-20		150W	μ-gravity (0.01g)
	21-25			Lunar (0.18g)
	26-30			Martian (0.38g)
	Day 2
	0-5	Water -16% Ethanol FR = 50%	50 W	μ-gravity (0.01g)
	26-30		150 W	μ-gravity (0.01g)
	Day 3
	6-10	Water -28% Methanol FR = 50%	50 W	μ-gravity (0.01g)
	21-25		150 W	μ-gravity (0.01g)

Pressure [hPa] / Heat load[W]

Pressure [kPa] / Heat load[W]

Acknowledgements

So, a review on the experimental and numerical studies, intended to shed light on the different parameters influencing the pulsating heat pipe operation, will be presented in the following section.

PART II FLAT PLATE PULSATING HEAT PIPE EXPERIMENTAL STUDIES

The uncertainty analysis also have been provided. The figures, used in following chapter are described during analysis of the first experimental results. Repeatability analysis provided, which also prove the system reliability.

Chapter 7

Definition of Flow Transition in Microgravity via High-Speed

Visualization

The visual analysis of the two-phase flow inside the PHP could give some answers on the questions related to the flow behavior, the influence of sensible heat transfer on the overall performance of PHP, which is still contradictory [7.1; 7.2; 7.3; 7.4; 7.5; 7.6] Par conséquent, les principaux objectifs de cette thèse sont de caractériser les performances de transfert de chaleur de caloducs oscillants plats fonctionnant sous gravité normale et réduite, de comprendre le comportement des écoulements diphasiques à l'intérieur de ces systèmes et d'améliorer les performances de transfert de chaleur en utilisant des fluides caloporteurs ayant différentes propriétés.