A Review of Working Fluids and Flow State Effects on Thermal Performance of Micro-Channel Oscillating Heat Pipe for Aerospace Heat Dissipation

Abstract

A MCOHP (micro-channel oscillating heat pipe) can provide lightweight and efficient temperature control capabilities for aerospace spacecraft with a high power and small size. The research about the heat flow effects on the thermal performance of MCOHPs is both necessary and essential for aerospace heat dissipation. In this paper, the heat flow effects on the thermal performance of MCOHPs are summarized and studied. The flow thermal performance enhancement changes of MCOHPs are given, which are caused by the heat flow work fluids of nano-fluids, gases, single liquids, mixed liquids, surfactants, and self-humidifying fluids. The use of graphene nano-fluids as the heat flow work medium can reduce the thermal resistance by 83.6%, which can enhance the maximum thermal conductivity by 105%. The influences of gravity and flow characteristics are also discussed. The heat flow pattern changes with the work stage, which affects the flow mode and the heat and mass transfer efficiency of OHP. The effective thermal conductivity varies from 4.8 kW/(m·K) to 70 kW/(m·K) when different gases are selected as the working fluid in OHP. The study of heat flow effects on the thermal performance of MCOHPs is conducive to exploring in-depth aerospace applications.

1. Introduction

The working environment of aerospace equipment is special, including cold and heat exchange [1,2], heat transfer [3,4], energy mode conversion [5,6], and energy management [7,8]. The heat dissipation that is caused by heat conduction [9,10], heat convection [11,12], and radiation [13,14] during the operation of aerospace equipment should not be underestimated. Oscillating heat pipe (OHP) is the preferred heat dissipation technology for aerospace equipment, battery, and electronic equipment, due to its superior heat transfer performance, simple structure, and miniaturization [15,16]. The changes of the array layout, heating mode, working fluid, pipe wall material, and working fluid flow in the pipe are the primary factors that affect the OHP heat transfer performance [17]. The OHP has been studied using a variety of working fluids and nano-fluids have emerged as a research fad [18]. OHPs with a 3D structure have been mentioned in recent years [19], which can provide excellent performance in some space heat transfer scenarios [20]. The OHP turns [21], and the heating settings and pulsing heating techniques also have an impact on the thermal performance [22,23].

Based on the outcomes of the visualization, Senjaya et al. [24] developed a model for the creation and expansion of tube diameter bubbles. The liquid slug velocity was less than 0.2 m/s, which promotes bubbles formation. Ando et al. [25] tested the OHP with a check valve, which is located near the condensation section. The valve intermittently supplied liquid plugs to the evaporation section to start the pipe successfully. Wang et al. [26] investigated the intricate hydrodynamic processes and presented a new, closed OHP with periodic expansion and contraction condensation. The thermal efficiency was raised by 45% and the contraction condensation increased the oscillation frequency of the vapor plug/liquid slug. Shi et al. [27] performed an experiment investigation on a closed-loop OHP with a 0.5% mass-concentrated micro-encapsulated phase change material suspension. The vertical installation performed better for heat transfer under the effect of gravity. Lim et al. [28] adopted a randomly arranged pipeline layout OHP and studied the internal working fluid oscillation. The liquid slugs oscillated with large amplitude in every channel and the thermal performance was 32% higher than MCOHPs with a uniform channel layout. Liu et al. [29] made 15 turns of the anti-gravity OHP and carried out the research with the heat recovery rate of 1.66 times than that of pure copper tubes, which proved that gravity had a positive effect on the internal hydrodynamics.

The heat transmission capabilities of liquid metal and water as the working fluids of OHPs were compared by Hao et al. [30]. The heat transfer performance of the OHP is increased by 13% when the liquid metal is the working fluid instead of pure water. Schwarz et al. [31] proposed two design methods of floral OHP and star-shaped OHP and conducted experiments. The floral OHP increased the latent heat transfer, while the star-shaped OHP improved the convective heat transfer. The floral OHP design reduced the thermal resistance by 7% in the horizontal position and 12% in the vertical position. Kwon et al. [32] reported the influence of double-diameter tubes on single-turn OHPs. The circular flow was promoted by double-diameter tubes, which reduced the thermal resistance of the OHP by 45%. Liu et al. [33] designed and manufactured a new type of flat-plate OHP (FPOHP) with double serpentine channels. FPOHPs can successfully start at all tilt angles from 0° to 90°. The thermal conductivity of FPOHPs is 5.8 times than that of ordinary OHPs. Arai et al. [34] designed three kinds of polycarbonate OHPs of actual flow channel structures with an additive manufacture method and conducted research. By comparison of the effective thermal conductivity of different flow channel sizes at the same filling ratio, the effective thermal conductivity of a 0.8 mm square flow channel was about seven times that of a 2 mm square flow channel.

The working fluids and flow state have effects on the thermal performance of the micro-channel oscillating heat pipe, which are confirmed by all the above studies. In this paper, the effects of working fluids and the flow state on OHPs factors on the heat transfer performance are studied. The heat dissipation mechanism and technical characteristics of OHP are introduced. The effects of various filling working fluids including metal nano-fluids, filling working fluids, non-metallic nano-fluids, mixed nano-fluids, gas, organic solvents, mixed liquids, surfactants, and self-rewetting fluids (SRWFs) on heat dissipation are listed in detail. The influence of gravity on the flow, the flow pattern characteristics, and the two-phase oscillating flow are compared.

2. Heat Dissipation Mechanism and Characteristics of the MCOHP

2.1. Heat Dissipation Mechanism of the OHP

The working fluid is filled into a vacuum pipe with a certain proportion, which can be blended into various shapes and divided into an evaporation section, adiabatic section, and a condensation section [35]. The working fluid absorbs heat in the evaporation section and releases heat in the condensation section, which can transfer the heat generated at the evaporation section to the condensation section to achieve heat dissipation [36]. The mass forms randomly distributed air and liquid plugs under the effect of temperature difference between the cold and hot ends and the surface tension [37,38]. Due to the pressure difference between the evaporation section and the condensation section, the working fluid is driven to flow to the condensation section. The working fluid flows back to the evaporation section under the gravity action after the heat release [39]. The working process of the OHP is given in Figure 1.

There are many complex physical phenomenon and heat transfer processes in the internal operation of the OHP, including heat convection, latent heat transfer, pressure difference, temperature difference, inertial force, friction, gravity, and other factors [40], which are complex coupling processes [41,42]. When the OHP is placed vertically and the heater is located at the bottom, the thermal resistance representing the heat transfer performance of the OHP can be obtained from Equation (1) [43], after reaching a pseudo-steady state under each heating power condition.

$$R_{th}=\frac{{\overline{T}}_{\mathrm{evap}}-{\overline{T}}_{\mathrm{cond}}}{Q_{in}}=\frac{\frac{1}{t_a}\frac{1}{W}{{\displaystyle \int }}_0^{t_a}{{\displaystyle \int }}_0^W{T}_{Si}(x,L,t)dxdt-\frac{1}{t_a}\frac{1}{W}{{\displaystyle \int }}_0^{t_a}{{\displaystyle \int }}_0^W{T}_{Si}(x,0,t)dxdt}{Q_{in}}$$

(1)

where R_th is the thermal resistance, K/W. ${\overline{T}}_{\mathrm{evap}}$ and ${\overline{T}}_{\mathrm{cond}}$ are the average temperature of evaporation and condensation, respectively, K. Q_in is the input power, W. t_a is the time interval for time averaging, s. W and L are the width and length of the OHP, respectively, m. T_Si is the temperature of the silicon substrate, K. x is the horizontal coordinate, m. t is the time, s.

It is assumed that the temperature distribution of the liquid film is linearly related to the thickness of the liquid film. The total heat transferred from the heating wall to the liquid film and the heat transferred from the liquid film to the cooling wall is calculated by Equation (2) [44].

$$\left\{\begin{array}{c}{Q}_{\mathrm{w},\mathrm{H}}=(T_{\mathrm{w}}-T_{sur,\mathrm{H}})\pi d{x}_{\mathrm{H}}{\lambda }_l/\delta \\ Q_{w,C}=(T_{sur,\mathrm{C}}-T_{\mathrm{w}})\pi d{x}_C{\lambda }_l/\delta \end{array}\right.$$

(2)

where Q_w,H and Q_w,C are the total heats transferred from the heating wall to the liquid film and from the liquid film to the cooling wall, W. T_w is the wall temperature, T_sur,H and T_sur,C are the temperature of liquid film during heating and that of cooling, respectively, K. λ_l is the thermal conductivity of the liquid, W/m·K. δ is the liquid film thickness, m.

In OHPs, the liquid phase is regarded as an incompressible flow and the gas phase as an ideal gas. The temperature and pressure in each vapor plug are uniform. The temperature changes for liquid plugs and vapor plugs can be calculated by Equation (3) [45].

$$\left\{\begin{array}{c}{c}_{p,l}{\rho }_l{A}_{cr}\frac{\partial T_l}{\partial t}=h_{w,l}\left(T_w-T_l\right)S+{\lambda }_l{A}_{cr}\frac{{\partial }^2{T}_l}{\partial l^2}\\ \frac{d{T}_v}{dt}=\left(Q_{v,sen}+Q_{v,lat}-P_v\frac{d{V}_v}{dt}\right)/c_{v,v}{m}_v\end{array}\right.$$

(3)

where c_p,l is the specific heat of the liquid, J/(kg·K). ρ_l is the density of the liquid, kg/m³. A_cr is the cross-sectional area of the flow path, m². T_w and T_l are the temperature of channel wall and liquid plug, respectively, K. h_w,l is the heat transfer coefficient between the channel wall and the liquid plug, W/(m²·K). S is the perimeter of the liquid plug, m. T_v is the temperature of the vapor plug, K. Q_v,sen and Q_v,lat are the total amount of sensible heat and latent heat, respectively, W. P_v is the pressure of vapor plug, Pa. V_v is the volume of the vapor plug, m³. c_v,v is the specific heat at constant volume, J/(kg·K). m_v is the mass of the vapor plug, kg.

2.2. Technical Characteristics of OHP

2.2.1. Excellent Heat Transfer Performance

Yu et al. [46] designed the 18-channel 3D-OHP with liquid metal as the working fluid, as listed in Figure 2a. The thermal resistance was as low as 0.0351 °C/W and the heat transfer performance was improved by 20.57% compared with the pure ammonia OHP. Thompson et al. [47] measured the thermal resistance data (0.08 °C/W) of a plate OHP with staggered micro-channels when the working fluid, heating mode, and heat pipe matched. Ji et al. [48] manufactured the high-temperature OHP for liquid metal. The minimum thermal resistance was 0.08 °C/W with a filling rate of 45% and a heating power of 3168 W. Ji et al. [49] tested the high-temperature vibration OHP of a liquid metal through different proportions of NaK. The minimum thermal resistance was 0.071 °C/W with 3528 W input power and a 90° inclination angle. Czajkowski et al. [50] studied the patterned OHP with a special rotation system. The thermal resistance was 0.05 °C/W when the rotation speed was higher than 200 rpm. Qu et al. [51] discovered the OHP of spherical Al₃O₃. The maximum thermal resistance of the OHP decreased by 0.14 °C/W with the filling rate of 70%. The thermal resistance of water was lower than that of ethanol [52]. Tokuda et al. [53] tested the double-loop closed OHP made of Incoloy 800 HT with sodium and the effective thermal conductivity was 2.6 × 10³ to 2.3 × 10⁴ W/(m·K). Zhao et al. [54] constructed a sizable OHP experiment to investigate the variables of thermal conductivity. The OHP had a significant thermal load with an effective thermal conductivity of 5676 W·m⁻¹·C⁻¹ when the filling rate was 40%. Lin et al. [55] took aluminum materials to make interconnected rectangular channels of the OHP for heat dissipation of high-power LED, as given in Figure 2b. The thermal resistance was 0.18 °C/W with the heating power of 110 W to achieve a good heat dissipation.

2.2.2. Simple Structure of the OHP with a Small Volume

The most prominent characteristics of OHPs are its miniaturization abilities and simple structure, as listed in Figure 3a. An OHP can oscillate by itself without the liquid suction core and the assistance of other equipment. Qian et al. [56] applied loop OHP to study the heat dissipation of grinding wheels. Monroe et al. [57] used a four-loop OHP to test the fluid stirring of magnets to collect energy with aluminum blocks to assist heating and cooling. Zhao et al. [58] carried out the experiment with three turns of the OHPs to achieve the thermal properties of the coupling phase change materials. Qu et al. [59] made three kinds of copper OHPs (a 2D-OHP, three-layer 3D-OHP, and four-layer 3D-OHP) and studied the coupled heat transfer of the phase change materials. Jin et al. [60] used high-temperature quartz glass to make transparent OHPs with high solar light transmittance, which can realize the experiment research of solar energy-absorbing nano-fluids. The highest thermal conductivity can be achieved when the OHP is filled with 3.0 wt.% nano-fluids. Alqahtani et al. [61] explored the influence of the bending degree of OHPs on heat transfer. There is no significant effect on the thermal performance when the bending angle increases. Iwata et al. [62] developed 10 laps of OHPs. A flexible and highly conductive tropical belt was formed, which can be used as a cooling device in a spacecraft. Wei et al. [63] clamped between commercial battery packs to simulate the thermal power generated by two adjacent battery modules, as shown in Figure 3b. The OHP was filled with the dual fluid mixed ethanol–water and the size was consistent with the length of the battery pack. The evaporation section was heated by the battery and the condensation section used fans to dissipate heat, which had a good battery cooling effect and provided a new idea for the battery cooling of electric vehicles.

3. Effect of Various Filling Working Fluids of OHPs on Heat Dissipation

The common working fluids used in OHPs are nano-fluids, gases, single liquids, mixed liquids, surfactants, and SWRF.

3.1. Metal Nano-Fluid

A metal nano-fluid refers to a new type of heat transfer medium with uniform, stable, and high thermal conductivity, which is prepared by dispersing metal nano-powders into the base liquid [64]. At an appropriate concentration, nano-fluids have better thermal conductivity [65,66] and higher heat transfer limits [67] than traditional working fluids. The common nano-particles are metals (Al, Ag, Cu, Fe, etc.) and metal oxides (Al₂O₃, Fe₃O₄, TiO₂, etc.).

Table 1. Thermal properties of metal nano-fluids.

Metal Nano-Category	Particle Size	Concentration	Liquid Filling Rate	Inclination Angle/°	Heating Power/W	Reduction of Thermal Resistance
Al2O3 [68]	−	0, 0.1, and 0.5 wt.%	50%	0, 90	10~80	15.8%
Al2O3 [69]	56 nm	0~1.2 wt.%	50%	90	20~140	25.7%
Ag [70]	50 nm	50, 200, and 600 ppm	−	−	314, 385, and 488	30%
Al2O3 [71]	10~30	0.5, 1, and 3 wt.%	20%, 40%, 60%, and 80%	10, 40, 70, and 90	20, 30, and 40	Improved thermal performance by 56.3%
ɤ-Fe2O3 [72]	20	2 vol.%	50%	90	0~160	12%
Fe2O3 [73]	20	2 vol.%	50%	0~90	10~90	24.1%
Fe3O4 [74]	5~20	90, 270, and 450 ppm	70%	90	20, 55, 90, 125, and 160	27.6%
Fe3O4 [75]	25	0.2 wt.%	50%	90	0~200	11%
NiFe2O4 [76]	25	1.5, 3 wt.%	−	90	200, 300, and 400	30.4%

demonstrates the metal nano-fluids effect on the heat transfer performance of OHPs. The heat transfer performance of the OHP can be significantly improved by the appropriate particle size [68,69], fluid concentration [70,71], filling rate [72], tilt angle [73], and heating power [74] of metal nano-particles. Furthermore, for metal nano-fluid OHPs, an applied magnetic field helps to reach the start-up faster at low heat input conditions [75,76].

Karthikeyan et al. [77] conducted an experimental study on a COHP with colloidal nano-fluids of Cu (average nano-particle size is 100 nm) and Ag (average nano-particle size is 60 nm). Compared with deionized water, the Ag nano-fluid increased the OHP heat transfer limit by 33.3% and the evaporation wall temperature was lower. The shape of the nano-particles also had an impact on the thermal resistance. Kim et al. [78] found that the OHP thermal resistance of Al₂O₃/acetone nano-fluids with spherical, brick, and cylindrical nano-particles decreased 33%, 29%, and 16%, respectively. The nano-particles effect on the thermal performance of the OHP was revealed by Jafarmadar et al. [79] with Al₂O₃, CuO, and Ag. The flow, heat transfer, and entropy generation of the OHP in the case of pure water were checked. The entropy produced by Ag was the highest. The volume concentration of nano-particles was 0.5~1%, which can minimize the generation of entropy and proper thermal operation. Goshayeshi et al. [80] studied the influence of nano-fluids on the flow and thermal properties of OHPs with Fe₂O₃/kerosene, as Figure 4 displayed. The five-flow modes were obtained of the evaporation section when the filling rate of the Fe₂O₃ nano-fluids was 50% (average nano-particle size of 20 nm with the concentration of 5 vol.%). With the increase in heat (10~80 W), the bubble flow, slug flow, foam flow, annular steak flow, and annular flow will gradually appear in the evaporation section. When the liquid plug speed was ≤ 0.15 m/s, bubbles with a diameter equal to the inner diameter of the pipe were generated. Subsequently, [81] compared Fe₃O₄/water and the effect of -Fe₂O₃/kerosene nano-fluids on the heat transfer performance. Fe₃O₄/water and -Fe₂O₃/kerosene nano-fluids reduced the thermal resistance by 30.8% and 16.7%. Gandomkar et al. [82] studied the glass and copper OHP of ferromagnetic fluid under different magnetic fields through visual experiments. The place with the magnetic field had a smaller thermal resistance and the best thermal performance of the copper OHP. The performance without a magnetic field was the best for the glass OHP. Monroe et al. [83] examined the performance of solenoid-assisted OHPs for CoFe₂O₄ nano-fluids. A ring magnet was used to magnetize and the CoFe₂O₄ nano-fluids improved the heat transfer of heat pipes by 58%.

3.2. Non-Metallic Nano-Fluid

Non-metallic nano-fluids are SiC, CNT, graphene, CaCO_3, and other compounds.

Table 2. Thermal properties of non-metallic nano-fluids.

Non-Metallic Nano-Fluids	Concentration	Liquid Filling Rate/%	Inclination Angle/°	Input Power/W	Reduction of Thermal Resistance
MWCNTs [84]	0.05, 0.1, 0.2, and 0.3 wt.%	60	90	50 ~400	About 36.2%
Graphene [85]	0.025, 0.05, 0.075, and 0.1 wt.%	−	0~90	20~120	48.4%
Graphene oxide (GO) [86]	0.01 and 0.03 vol.%	100	90	50 ~400	Maximum heat transfer enhancement 25%
C60 [87]	0.1, 0.2, and 0.3 wt.%	50	50	10~60	36%
Hydroxylation MWNTs [88]	0.1~1 wt.%	50	90	–	34%
Graphene Nano-sheets [89]	1.2, 2, 5.7, 9.1, 13.8, and 16.7 vol.%	45, 55, 62, 70, and 90	90	10 ~100	83.6%
Graphene oxide [90]	0.25, 0.5, 1, and 1.5 g/L	50	90	10~70	42%
Oligographene (FLG) [91]	0.1, 0.3, 0.5, 0.75, and 1 mg/mL	55	90	20~60	25.16%
Carbon nano-tubes (CNTs) [92]	0.05, 0.1, 0.2, 0.3, and 0.5 wt.%	35	90	8~56	About 66.6%
Graphene oxide [93]	0.02~0.1 wt.%	20, 50, and 80	90	10~30	54.34%
SiO2 [94]	0.5, 1, 1.5, and 2 wt.%	50	90	10~50	40.1%

highlights the influence of non-metallic nano-fluids on the heat transfer performance of OHPs. In the OHP, heat transfer occurs due to repeated pressure fluctuations, with higher heat transfer occurring with more repetitions of pressure fluctuations. To provide a higher frequency of pressure and an average pressure inside the OHP, Tanshen et al. [84] used an aqueous solution of 0.2 wt.% of multi-walled carbon nano-tubes (MWCNTs) to experimentally investigate the thermal resistance and pressure fluctuations inside the OHP. Sadeghinezhad et al. [85] found through experimental studies that the deposition of graphene formed a coating on the surface of the sintered core in the evaporator section. This coating increased the surface wettability and thus improved the thermal performance of the heat pipe. Kim and Bang [86] discovered that the capillary limit of heat pipes containing graphene oxide/water nano-fluids was higher than that of the aqueous heat pipes. This is because the nano-particle coating changes the effective capillary radius and the bending moon surface, leading to an increase in the maximum fluid flow rate through the core structure. On the other hand, Wu et al. [87] showed in their study that the variation of thermal load has a greater effect on the thermal performance of the OHP than the variation of concentration. Beyond this, the addition of nano-particles to the working fluid can significantly enhance the heat transfer characteristics of the OHP and further improve the heat dissipation capacity of the OHP [88]. Zhou et al. [89] indicated that the addition of a graphene nano-sheet nano-fluid to distilled water can alleviate drying and improve the heat transfer performance of OHPs. Nazari et al. [90] reported that the addition of graphene oxide flakes improved the thermal conductivity and viscosity of the base fluid. Furthermore, the high concentration of nano-fluids reduces the thermal properties of OHPs compared to pure water, which is attributed to the increase in the dynamic viscosity of the nano-fluid. To prepare graphene nano-fluids with excellent stability, Xu et al. [91] as well as Zhou et al. [92] chose to use ethanol–water mixtures as the base fluid. In addition to this, the addition of appropriate graphene oxide nano-particles improved the OHPs initiation performance [93]. Zhang et al. [94] revealed that the addition of nano-particles promoted the phase transition of the work fluid in the OHP on the one hand, while increasing the transient velocity and driving force of the work fluid on the other hand. These are conducive to the reflux of condensate, and they can effectively avoid the dry-out phenomenon.

Sadeghinezhad et al. [95] studied the thermal properties of copper sintered heat pipes with graphene nano-fluids at different dip angles of 0°~70°and liquid filling rates of 30~60%. The maximum thermal conductivity of the graphene nano-fluids heat pipe (5 vol.%) was increased by 105% and the thermal resistance was reduced by 26.4%. Khajehpour et al. [96] discovered the performance of the L-shaped OHP with SiO₂ nano-fluids with different nano-particle sizes (11~14 nm and 60~70 nm). The experiment thermal resistance increased with the nano-particle size. For SiO₂ nano-fluids at 11~14 nm (0.5 wt.%), the maximum reduction of thermal resistance was about 24% at the vertical position under a heat load of 10 W and a liquid filling rate of 100%. Li et al. [97] studied the thermal performance of OHPs with aqueous ethylene glycol-based graphene nano-fluids. The minimum thermal resistance of 0.36 K/W was achieved at a thermal load of 85 W and a liquid filling rate of 35% for 2 g/L graphene nano-fluids. Choi [98] tested the thermal performance of thermosyphon heat pipes with cellulose nano-fluids, which increased the boiling heat transfer coefficient by about 71.74%.

3.3. Mixed Nano-Fluid

Both metallic and non-metallic nano-fluids are prepared by suspending a single nano-particle in a base solution to obtain a stable suspension. Mixed nano-fluids are made up of two or more different nano-particles [99]. Zufar et al. [100] studied the thermal performance of Al₂O₃-CuO/water mixed nano-fluids (0.1 wt.%) and SiO₂-CuO/water mixed nano-fluids (0.1 wt.%) under different heat inputs (10–100 W) with the liquid filling rates of 50–60%. A minimum thermal resistance of 0.27 °C/W can be obtained with SiO₂-CuO mixed nano-fluid. The thermal resistance of Al₂O₃-CuO and SiO₂-CuO mixed nano-fluids were reduced by 57% and 34%, respectively [101]. Moghadasi et al. [102] conducted a 3D numerical study on the laminar flow and heat transfer of Al₂O₃-CuO/water mixed nano-fluids in a U-shaped bend in porous media. The temperature and velocity contour of different volume fractions with water at the base fluid and mixed nano-fluids are given in Figure 5a,b, where φ is the volume fraction and r_p is the porosity. Nano-fluids are applied in the presence of porous media as fluids accumulate near the walls and enhance the heat transfer. When the volume fraction changed from 1% to 5%, the velocity distribution improved, and the temperature gradient increased. Xu et al. [103] studied the performance of thermosyphon OHPs mixed with Al₂O₃-TiO₂/water (0.2 vol.%). The conditions under different filling rates (30~70%) and coolant flow rates (0.4~0.56 L/min) were compared with 25% (Al₂O₃) and 75% (TiO₂) mixed nano-fluids, which achieved a thermal resistance reduction of 26.8% and increased the thermal efficiency by 10.6%. Mukherjee et al. [104] configured the SiO₂-ZnO/water mixed nano-fluids of different mass fractions (0.025~0.10 wt.%) and Reynolds numbers (7743~23,228). The thermal conductivity at 60 °C can be increased by up to 30% with a mass fraction of 0.10 wt.%. Veeramachaneni et al. [105] fabricated a rectangular flat loop OHP for electron cooling applications with Cu-graphene/water mixed nano-fluids (0.1~0.2 vol.%). For a mixed nano-fluid with a volume concentration of 0.02%, the capillary limit increased by 36.97% and the wall temperature of the evaporation section decreased by 9.8%. The mixed nano-fluid with a copper/graphene ratio of 30:70 can obtain a minimum thermal resistance of 0.1 K/W.

3.4. Gas Working Fluid

The gas can be applied as the working fluid of OHPs if the temperature of the working environment is too low. The researchers explored various gases of neon, argon, nitrogen [106,107], helium [108], and hydrogen [109,110]. Liang et al. [111,112] experimentally tested OHPs with neon, and the maximum effective thermal conductivity was 22.18 kW/(m·K) at an optimal filling rate of 24.5%. Barba et al. [113] found that the expansion and contraction of gases play an important role in the circulation of working fluids. The circulation was hindered when the filling rate was too low and the OHP could not be started normally. Sun et al. [114] simulated OHPs with hydrogen as the working fluid. The influence of hydrogen on the latent heat transfer was 45~51%, which is proportional to the volume fraction of gas. Li et al. [115] researched nitrogen as the working fluid and the thermal conductivity of the bottom heated was about 16 k W/(m·K), which was about 32 times that of pure copper. Xu et al. [116] conducted experimental studies on low-temperature OHPs filled with helium. The effective thermal conductivity was 4.8~13 kW/(m·K) with an inclination angle of 30° and a liquid filling rate of 70.8%. Fonseca et al. [117] took the experiments between 77 K and 80 K with nitrogen. The temperature difference between the OHP sections was small and the maximum thermal conductivity was 70 kW/(m·K), with a liquid filling rate of 20% [118]. The thermal performance of more than 2000 helium working fluid OHPs under the filling rate of 20~90% was tested. The liquid filling rate was 69.5% and the maximum effective thermal conductivity was 50 kW/(m·K). The non-condensable gas of the OHP reduced the evaporation amount to slow down the circulation of the working fluid, which can weaken the oscillation and reduce the heat transfer performance [119,120]. The OHP with a heat flux constant and wall temperature constant is displayed in Figure 6a. The effect of non-condensable gases on OHPs is demonstrated in Figure 6b. (Q is the heat flux, T is the temperature, and ω is the mass concentration). The higher the temperature of the evaporation section, the less influence of the non-condensable gas. Chen et al. [121] conducted a series of experiments to study the thermal performance of ethane OHPs (EOHPs) in the medium and low temperature regions (−90~0 °C). The liquid filling rate of the best performance of EOHPs was not affected by the operating temperature and heat input, which was always maintained at about 30%. The lowest corresponding thermal resistance was 0.02 °C/W at the inclination angle of 30° and the temperature of −80 °C. At a high heat input of 30~50 W, the latent heat of vaporization was the main characteristic that determined the thermal performance of the EOHP.

3.5. Organic Solvent

Table 3. Heat transfer performance with different organic solvent liquids as working fluids.

Working Fluids	Filling Rate/%	Input Power/W	Lowest Thermal Resistance/°C·W−1	Lowest Thermal Resistance Obtained
Ethanol [122]	0, 25, 37.5, 50, 62.5, 75, and 100	–	0.95	50%
Methanol [123]	20~95	5~100	0.2	95% and 100 W
Ethanol [124]	50	15~50	0.6244	50 W
Acetone [125]	50 ± 5% and 70 ± 5%	60~300	0.092	70% and 260 W
Lonic liquids [126]	65 ± 5	50~250	0.15	44.4% and 250 W
Acetone [127]	50	10~200	0.14	200 W
LiCl solution [128]	45, 55, 62, 70, 80, and 90	10~100	0.9	62% and 10 wt.%
Acetone [129]	0~100	10~120	0.39	60% and 100 W
R1233zd(E) [130]	40~70	0~200	0.1184	50% and 70 W

summarizes some trends in the heat transfer performance of OHPs with different organic solvent liquids as the working fluids. From the information in Table 3, the working fluids, filling rate, input power, lowest thermal resistance, and lowest thermal resistance obtained from the nine papers are compared. At a low heat input, the heat transfer depends strongly on whether the oscillations are triggered or whether the oscillatory flow is triggered fast. In contrast, the effect of viscosity on the heat input and the effect of the latent heat of vaporization increases at a high heat input. It can be concluded from Table 3 that the choice of the working medium and filling ratio should be determined according to the actual situation. In terms of taking full advantage of low thermal resistance, 50% is a good filling ratio for ethanol, which should have a low heat input. In the case of a large filling ratio and large heat input, methanol is a better choice. Acetone has a good fill ratio of 50–70% with high heat input. The best filling ratio for lonic liquids is about 44% with high heat input. The LiCl solution performs well at a 62% filling rate. R1233zd (E) with a filling rate of about 50% at moderate heat input is also an option.

Takawale et al. [131] studied the performance of FPOHPs and capillary tube OHPs (CTOHPs) under different heat inputs (20 W~180 W) and liquid filling rates (40%, 60%, and 80%). After ethanol was filled into the OHP, the thermal resistance of the FPOHP and CTOHP decreased by 83% and 35%, respectively. Bastakoti et al. [132] tested the heat transfer performance of OHPs with methanol, ethanol, cetyltrimethylammonium chloride (CTAC), and deionized water as the working fluids. The heat pipe charged into the CTAC had the lowest thermal resistance of 0.30 K/W. The thermal resistance of the OHP with deionized water, methanol, ethanol, and acetone as the working fluid tended to increase after the heating power reached 65 W identically [133]. Bae et al. [134] established a numerical model of the OHP and simulated the change in the liquid film thickness. The numerical model was based on a 1D piston flow hypothesis. Figure 7a is a schematic diagram of two vertical heat pipes. Figure 7b is the plug flow on the z-axis. Figure 7c is a piston flow. Figure 7d is the liquid plug. Figure 7e is the plug (bubble). Figure 7f is the liquid film. Figure 7g is the pipe wall. The simulation results had an error of less than 20% compared with the experimental data, which proved that the oscillation prediction of fluids needs to consider membrane dynamics.

Sun et al. [135] studied the hydro-thermodynamic behavior of the ethanol-based bubble distribution, bubble motion, and temperature of the working fluid. The proportion of small-sized bubbles increased with the improvement in the liquid filling rate and heating power. The proportion of medium-sized and large-sized bubbles decreased when the oscillation frequency and amplitude of bubbles increased. The high boiling point fluid working fluid heat pipes was studied by Mahapatra et al. [136] using the Buckingham’s pi theorem to perform a dimensionless analysis of the heat transfer performance. High boiling point working fluids alleviated locally high heat flux densities. Xue et al. [137] conducted a novel full-visualization experiment on ammonia water OHPs with a high-speed camera. As the heating power increased, the flow pattern changed from a stopper flow to a ring flow and the proportion of evaporation heat increased from 7.7% to 32.4%. Liu et al. [138] analyzed the starting performance of the OHP based on system identification theory. The working fluids with small dynamic viscosity, small specific heat, and large saturation pressure gradients favored the start-up of OHPs. Hao et al. [139] studied the effect of polytetrafluoroethylene with ionized water, ethanol, and acetone as the working fluids on the heat transfer performance. When acetone was used as the working fluid, the liquid plug oscillation amplitude and speed were the highest and the thermal resistance was 30~63% lower than that with water as the working fluid.

3.6. Mix Liquids

Different pure working fluids have their own advantages under working conditions. Non-azeotropic mixtures have the characteristics of phase change and temperature fluctuation, which can make the heat source and working fluid well-matched [140]. The mixture plays the superior characteristics of each different components, which cause the OHP to achieve better start-up and heat transfer performances [141]. Zhu et al. [142] concluded that the thermal resistance of the OHP was filled with a ketone-pure–water mixture. When the filling rate was high, the thermal resistance of the OHP filled with pure water and acetone was 45.8% and 38.7% lower than that of the ketone-pure–water mixture. The mixture had better resistance to dryness at a low liquid filling rate [143]. Shi et al. [144] studied the OHP with ethanol–water, ethanol–methanol, and ethanol–acetone as working fluids with different mixing ratios. When the filling rate was increased to 62%, the heat transfer performance of OHPs with pure working fluids was better than that with mixed working fluids. When the filling rate reached 70%, the thermal resistance of the different working fluids tended to be approximated with the increase in the heating power. When the filling rate was low, the methanol working fluid can inhibit the drying up of the OHP [145]. Xu et al. [146] tested the effect of HFE-7100 and the lowest thermal resistance was 0.1634 °C/W with the mixing ratio of the working fluid was 1:2. Chang et al. [147] obtained the internal pressure of methanol-deionized water OHPs with different mass ratios. When the mass ratio of binary working fluid methanol and deionized water was 1:5, the starting performance of the OHP was the best with the temperature of 80 °C, the thermal resistance of 0.114 °C/W, and the heat flux density of 1.47 W/cm².

Markal and Varol [148] studied the effects of the volume mixing ratio, inclination angle, and fill ratio on the OHP thermal performance of ethanol (E)–pentane (P) mixtures under different heat inputs. The ethanol–pentane mixture exhibited lower thermal resistance and had the best thermal performance with the filling rate FR of 30%, the dip angle IA of 90°, and the mixing ratio E:P of 1:3. Under the same filling rate and inclination conditions, the ternary mixture of deionized water (W), methanol (M), and pentane (P) had a better thermal performance when the mixing ratio was 1:2:3, as shown in Figure 8. The best thermal performance occurred when the filling rate was 50% [149]. Compared with the two binary mixtures of ethanol–pentane and methanol–pentane, the overall performance of the ternary mixture was low. Markal and Varol [150] also compared the effects of pentane–methanol, methanol–hexane, and water–methanol–pentane mixtures on OHP thermal properties. The immiscible pentane–methanol (P:M = 1:1) mixture had better thermal properties than the mixture of hexane–pentane.

3.7. Surfactants and Self-Rewetting Fluids

When the working fluid flows in the MCOHP, the flow of the working fluid will be affected by resistance due to the presence of surface tension. The surface tension of the working fluid can be reduced if surfactants are added. The capillary resistance can be reduced and the heat transfer performance of the OHP can be improved [151,152]. Hao et al. [153] conducted a series of experiments to study the effects of super-hydrophilic and hydrophilic surfaces on the segment plug motion of OHPs. The influence of surface wetting characteristics on the gas–liquid interface at the end of the plug is shown in Figure 9. The length of the film in super-hydrophilic OHPs is significantly increased. Compared with the copper OHP, the thermal resistance of the super-hydrophilic and hydrophilic OHP were reduced by about 5~15% and 15~25%, respectively. Xing et al. [154] obtained OHPs with a cetyltrimethylammonium bromide (CTAB) solution as the working fluid, which can reduce the surface tension of the solution and the contact angle. The thermal resistance of 0.25 wt.% of the CTAB solution is reduced by 48.5% with the filling rate of 50%. The addition of surfactants can increase the critical heat flux density of the heat pipe by 1.26 times with the enhancement of pressure fluctuations [155]. Compared with deionized hydraulic working fluid heat pipes, the thermal resistance of the cetyltrimethylammonium chloride working fluid was reduced by 4.78% [156], which prevents drying up. Bao et al. [157] experimentally proved that the thermal resistance of OHPs that had surfactants as the working fluid was reduced by a maximum of 27.8%.

Abe et al. [158] proposed the concept of SRWFs by the physical properties of dilute aqueous solutions with high carbon alcohols. A SRWF enhances the heat transfer performance and heat transfer limit of OHPs with an incensement of the surface tension and a reduction of the contact angle [159]. Hu et al. [160] used a heptanol–aqueous solution to study the enhancement effect of SRWFs. The characteristics of SRWFs caused the working fluid to be spontaneously wetted in the overheated part of the tube. Wu et al. [161] applied butanol at a concentration of 6% as a working fluid for performance testing. The critical heat load was 650 W, and the total thermal resistance was 0.25 °C/W with a reduction of 60%. The SRWF nano-fluids exhibited an excellent heat transfer performance over the entire heat load range, with a maximum enhancement rate of approximately 15% [162]. The influence of a SRWF nano-fluid base prepared by mixing graphene oxide dispersion with an n-butanol–aqueous solution on OHPs had been studied [163]. The following percentages, 0.07 wt.% and 0.7 wt.% were the optimal concentrations of graphene oxide and n-butanol. The heat transfer performances were increased by 16% and 12% compared with the SRWF and nano-fluid. Savino et al. [164] performed a microgravity heat pipe experiment. They researched working fluids including aqueous alcohol solutions, multi-component brine and nanoparticle suspensions. It was shown that self-rewetting brines and self-rewetting nanofluid brines have good thermal properties. Wang et al. [165] conducted numerical studies on CLOHP with different wettability (contact angles of 5°, 33°, 147°, and 175°). Figure 10 shows the volume fraction of the liquid and vapor distribution at a heat load 20 W. Compared with superhydrophobic surfaces, CLOHPs on super-hydrophilic surfaces had a 10.8% reduction in thermal resistance at an input heat load of 20 W.

4. Effect of In-Tube Flow State on Heat Dissipation Properties

The heating method, effect of gravity, characteristics of the flow pattern, and the oscillatory characteristics have influences on the heat dissipation properties.

4.1. Different Heating Method of Evaporation Section

An OHP has a very flexible use, which is reflected in its ability to heat in different positions; for example, its heating methods can be pulse heating, alternating heating, and continuous heating, and it can have a number of evaporation and condensation sections. Lin et al. [166] used vertical bottom heating to study the heat transfer performance of OHPs. The pulse heating method had a lower temperature difference than the continuous heating method [167]. Zhao et al. [168] found that the advantages of thermal resistance were pulse heating, alternating heating, and continuous heating. Taft et al. [169] compared the heat transfer performance of OHPs under DC and pulse modulation input modes. Chu et al. [170] used asymmetric heating to study a 3D helix OHP. The non-uniform heating method of multiple heat sources has also been studied in series with a two-channel flat OHP [171], as given in Figure 11a.

Mangini et al. [172] studied the heat transfer performance of hybrid OHPs in the super/microgravity environment and the non-uniform heating configuration promoted the net circulation of the fluid in the preferential direction, which improved the thermal performance relative to uniform heating. Peng et al. [173] took the bottom heating method to perform a numerical simulation study on the completely non-linear thermo-mechanical finite element model OHP. Qu et al. [174] used vertical and horizontal heating methods to study 3D-OHP. Yasuda et al. [175] studied flat-plate OHPs made of aluminum alloy by bottom heating and top heating. Lim et al. [176] used local heating to study the flat MOHP. As listed in Figure 11b, the heating method will directly affect the internal flow.

4.2. Flow State

E et al. [177] established a CLOHP model using the VOF (volume of fluid) method as the solution scheme to numerically simulate the liquid vapor in the two-phase conversion process. The distribution and fluctuation relationship of the pressure and vapor flow mode during the start-up under different vacuum degrees was determined. The interphase mass transfer due to evaporation and condensation in the VOF method can be applied to energy jump conditions, the Tanasawa model, and the Lee model [178]; the relevant equations are given in

Table 4. Numerical model of interphase mass transfer.

Content	Remarks
Energy jump conditions [179] $$\dot{m}=-\frac{\left(L_v+(c_{p,l}-c_{p,v})(T_{sat}-T_{\mathrm{int}})\right)}{{[-k_{lv}\nabla T\cdot \overrightarrow{N}]}_{\Gamma }}$$ (4)	$\dot{m}$ is the phase change local mass flow rate, kg/(m3·s). Lv is latent heat. cp,v is constant pressure specific heat of vapor, J/(kg·K). Tsat is the saturation temperature associated with the considered pressure, K. Tint is the local interface temperature, K. klv is thermal conductivity, W/(m·K). $\overrightarrow{N}$ is the normal vector pointing in the direction of the gas phase at the Γ of the interface.
Tanasawa model [180] $$\dot{m}=\frac{2\gamma }{2-\gamma }{\left(\frac{M}{2\pi R_g}\right)}^{1/2}\frac{{\rho }_v{L}_v(T_{if}-T_v)}{T_v^{3/2}}$$ (5)	γ is the adjustment factor. M is the molecular weight. Rg is a general gas constant. is 8.314 J/(mol K). ρv is the density of vapor, kg/m3. Tif is the interface temperature, K.
Lee model [181] $$\left\{\begin{array}{c}{\dot{m}}_{lv}={\alpha }_l{\rho }_l\frac{c_{p,l}{T}_{sat}}{L_v}\cdot \frac{T_l-T_{sat}}{T_{sat}},T_l>T_{sat}\\ {\dot{m}}_{vl}={\alpha }_v{\rho }_v\frac{c_{p,v}{T}_{sat}}{L_v}\cdot \frac{T_{sat}-T_v}{T_{sat}},T_{sat}>T_v\end{array}\right.$$ (6)	${\dot{m}}_{lv}$ is the mass transfer of each time step in the evaporation process, kg/(m3·s). ${\dot{m}}_{vl}$ is the mass transfer of each time step in the condensation process, kg/(m3·s). αl and αv are the volume fraction of liquids and vapors.

.

The heat exchange units established by Nuntaphan et al. [182] can be used to evaluate the efficiency of the heat exchanger and the heat transfer coefficient of the air side. Qian et al. [183] proposed a novel heat transfer prediction model based on an extreme gradient boost algorithm and studied the design and cooling method of OHP prototypes during optimized processing. Sun et al. [184] established a model to study the oscillatory motion characteristics of liquid plug and vapor plug/bubble in OHPs. Nemati et al. [185] used numerical models to study the heat transfer mechanism of OHPs and predict heat transfer capacity, which simulated the oscillation behavior of the liquid plug, considering the thickness of the liquid film in the evaporation and the decrease in the liquid film thickness caused by evaporation. Daimaru et al. [186] proposed a numerical simulation method for OHPs with a check valve and new modeling features, including the pipe wall energy equation. The check valve model included pressure loss, the detailed boiling algorithm, and the pressure loss of a bending surface transformation. The temperature error of the heating section was less than 1.7 °C. Adachi et al. [187] developed transient models of fluid conditions to reproduce transmission lines. Odagiri et al. [188] combined thermo-fluid behavior in channels with thermal diffusion in OHP casing segments, which was in good agreement with the multi-branch OHP experiment results with a channel diameter of 1 mm and several turns of 42 turns.

4.3. Gravity Effect

Gravity has an important effect on the flow and circulation of the working fluid of OHPs. The gravity prevents the working fluid from flowing to the evaporation section in the top heating mode. The gravity promotes the flow of the working fluid to the evaporation section in the bottom heating mode. For micro-gravity research, the European Space Agency used parabolic flights of aircrafts to create different gravity environments. The variation of the tilt angle changed the flow pattern inside the OHP, resulting in different performance levels [189]. Mameli et al. [190] studied the OHP under different gravities. The change in gravity had a greater effect on the heat transfer performance of the OHP in the case of vertical heating compared with horizontal heating. The FC-72 working fluid was investigating the effect of gravity on the heat transfer performance at different heating powers [191]. The thermal resistance of vertical heating was lower than that of horizontal heating under the influence of gravity. The vertical heating was not as stable as the horizontal heating during operation. Ayel et al. [192] reported that closed-loop FPOHPs can respond to gravity changes more quickly and reach a steady state. Mangini et al. [193] tested the sudden loss of buoyancy and activated the oscillating segment plug/stuff flow state in micro-gravity, which had the lowest starting power. Cecere et al. [194] discovered that FPOHPs with SRWF (butanol–water) were easier to keep working under micro-gravity and a low heating power. Xing et al. [195] examined the effect of gravity on OHPs with a surfactant solution as the working fluid. The influence of gravity on the OHPs of the CTAB solution was relatively small. When the heating power was higher, the heat transfer performance of the CTAB solution OHP was stronger and the thermal resistance was reduced by 51%. Pagliarini et al. [196] trained the OHP used for the International Space Station in microgravity, where the two states of the working condition included intermittent flow (episodic fluid motion occurring in some channels) and full activation (steady fluid movement throughout the adiabatic section). The fully activated state is given in Figure 12. No significant variation between channel behaviors is observed with stable oscillations, high heat flux amplitude, and oscillation frequency. The heat flux amplitude increased almost linearly, 1500 W/m² at 202 W, with the power input from 1100 W to 202 W.

4.4. Characteristics of Flow Pattern in Tube

Awareness of flow boiling and two-phase instability are important parts of understanding the complex phenomena and developing OHP technology, to explore the physical mechanisms controlling the complex unsteady flow boiling heat transfer and two-phase phenomena [197], which provide insight into the heat and mass transfer relationships in OHPs. With the change in the OHP working stage, the flow pattern of the working fluid in OHPs also changes, which can directly affect the flow mode of the working fluid and the heat and mass transfer efficiency of OHPs. Yuan et al. [198] established a flow model of the liquid plug based on the Lagrange method. Liquid plug oscillation amplitude and angular frequency depend on the geometry of the OHP and the liquid filling rate. When the flow type is slug flow, the sensible heat transfer can account for more than 80% of the total heat transfer. Karthikeyan et al. [199] employed high-resolution infrared thermography to measure the flow characteristics inside OHPs. The flow is the working fluid without internal oscillation, intermittent oscillation, or continuous local oscillation. With the increase in heating power, the thermal resistance decreased from 1.90 K/W to 0.24 K/W. Spinato et al. [200] used time-strip image processing techniques to study the two-phase flow of the OHP. Low amplitude/high amplitude oscillations, cyclic oscillations, backflow, and steady cycles were observed. The nucleation and rapid growth of bubbles in the U-bend of the evaporation section lead to the transition of the working fluid from the circulating to the oscillating state. Xian et al. [201] visualized the flow behavior in an OHP duct with pulsed heating. The liquid film was thinner in pulsed heating than in continuous heating. Under the condition of a short period of pulse heating, the proportion of bubble flows increased. The flow pattern is the same as for continuous heating. Pouryoussefi et al. [202] worked with numerical methods to simulate the chaotic behavior of fluids, as listed in Figure 13. Volume fraction contours are provided for different time points for two different operating conditions (red is vapor and blue is liquid). Figure 13a is the volume fraction diagram of the OHP under a time series of 0.8 s, 2.5 s, 3.8 s, 5.5 s, 13 s, and 18 s, with the evaporation temperature T_h = 145 °C, the condensation temperature T_c = 35 °C, and the liquid filling rate of 30%. Figure 13 b presents the volume fraction plots for the time sequences of 2.2 s, 3.8 s, 5.8 s, 10 s, 16.5 s, and 18.6 s, with T_h = 150 °C, T_c = 35 °C, and a 60% filling rate. The relevant dimension increases by promoting the filling rate and evaporation temperature.

Feldmann et al. [203] simulated turbulence in a pipe at different Womersley numbers and found that disturbance energy is required to trigger the non-linear transition process in the subcritical state. Pouryoussefi et al. [204] modified the chaotic flow in a heat pipe based on the VOF method. There was an upper limit to the accuracy of the simulation as the fluid filling rate and heating power increased. The optimal filling rate and minimum thermal resistance were measured to be 60% and 1.62 °C/W. Mangini et al. [205] demonstrated the reliability of the infrared visualization of two-phase flows with a maximum error of ±1.5 °C in combination with a high-speed camera capable of detecting the wetting and drying of liquid films. Xia et al. [206] studied the properties of unsteady flows in parallel micro-channels. Continuous two-phase unsteady boiling often occurred when the flow rate and heat flow density were greater than 607.6 kg/m² and 30 W/cm². This phenomenon can be suppressed by increasing the forced convection heat transfer of 100% or increasing the flow boiling heat transfer of 50%. Yoon et al. [207] investigated the oscillation frequency of the liquid plug with a MOHP. The heat input can change the vapor temperature and affect the oscillation frequency. The oscillation frequency increased with the increase in the heat input. The liquid plug with a longer total length of the heat pipe had a lower oscillation frequency. Ling et al. [208] found that the heat pipe temperature fluctuated greatly during the stable circulation of the working fluid. Noh et al. [209] used a numerical model to simulate the heat transfer between the tube wall and the liquid slug/vapor plug and proposed guidelines for designing heat pipes. Figure 14a is the liquid slug and vapor plug distributions for a two-turn heat pipe with an input power of 50 W (t = 145 s) and a twenty-turn OHP with an input power of 48 W (t = 80 s). Figure 14b illustrates the heat flux distribution for the two heat pipes and Figure 14c presents the wall temperature distribution for the two OHPs. After the pseudo-steady state, the liquid slug in the two-loop heat pipe oscillates and shows a net circulating flow.

Ahmad et al. [210] investigated the heat transfer performance as a function of the flow pattern using OHPs with ethanol. The thermal resistance at a 50% filling rate was as low as 1.6 °C/W. Vo et al. [211] performed visualization experiments on the OHP and found that the cyclic motion dominated the motion of the working fluid. A 3D computational fluid dynamics model of the OHP was developed and the k-ε turbulence model was applicable to the heat pipe simulation. Schwarz et al. [212] discovered through visualization experiments that the thermal resistance of the fluid inside the heat pipe during start-up is constant at 0.43 °C/W when there is no flow (25 W to 75 W). The thermal resistance dropped to 0.34 °C/W (200 W)~0.36 °C/W (100 W) at an average velocity of 240 mm/s.

4.5. Oscillatory Characteristics

The pulsed heat pipe is a kind of non-equilibrium, passive two-phase heat transfer device with complicated internal transmission process. The heat transfer performance of OHPs depends on the oscillation degree and stability of the self-sustained oscillating two-phase flow, which leads to its unique heat transfer characteristics [213]. Yoon et al. [214] observed that the internal oscillation frequency of the OHP was between 40 Hz and 50 Hz. Spinato et al. [215] examined the flow behavior inside the OHP using time-slot image processing technology. The flow pattern changed from oscillating to circulating state while the main frequency changed from 1.2 Hz to 0.6 Hz. Dilawar et al. [216] researched the oscillating two-phase flow in a micro-channel based on a numerical model. The pressure loss of the oscillating two-phase flow at the bend of the tube reduced the oscillation amplitude and weakened the heat transfer performance. Kato et al. [217] investigated the OHP consisting of a single straight pipe and open pipe. The heat transfer performance was enhanced with an increased amplitude of the oscillatory flow. Liquid exchange due to the oscillatory motion significantly enhanced the heat transfer with an effective thermal conductivity of up to 40 kW/(m·K). Miura et al. [218] studied liquid column oscillations with a forced oscillator. The evaporation of liquid film produced by the oscillation of working fluid is the main process of latent heat transfer. Both evaporation and condensation occur on the liquid film on the wall of micro-channel. Daimaru et al. [219] processed oscillation data based on a fast Fourier transform and mutual analysis. The vapor plug received or applied energy according to the direction of propagation. Das et al. [220] suggested a theoretical model considering the two-phase oscillatory equilibrium, which was calculated in good agreement with the experiment results. Jung et al. [221] evaluated the effect of the oscillation amplitude of a miniature OHP on the heat transfer performance. When the input power was 16 W, the time-resolved distribution of the heat flux and the corresponding flow visualization after the miniature OHP reached the pseudo-steady state are given in Figure 15. The total heat transfer rates of the evaporation and condensation sections were 14.4 W and 14.2 W, while the average value of latent heat ratio increased from 54.8% to 81.9%. When the input power increased from 7 W to 16 W, the oscillation amplitude increased from 3.4 mm to 8.3 mm, with a 13.5% reduction in thermal resistance. The oscillating motion of the liquid slug and vapor plug in the OHP promoted the heat transfer between the evaporation section and the condensing section. The oscillation amplitude and frequency are important parameters for estimating the heat flow density in OHPs [222].

Pai et al. [223] introduced a non-linear thermal model based on a U-shaped three-plug OHP. When the ratio of liquid slug mass or vapor plug length to the tube cross-sectional area decreased, the initial air pressure, liquid filling rate, gravity, and the oscillation frequency increased. As the temperature difference or heat transfer coefficient between the evaporation and condensation sections increased, the liquid filling rate or the initial temperature decreased. Perna et al. [224] tested the main frequency of pressure signal oscillation in the evaporation and condensation sections under micro-gravity. The main frequency of the pressure signal was in the range of 0.6~0.9 Hz, which increased with the heating power. Simplifying the entire OHP into a single unit can facilitate the study of various parameters of oscillation behavior [225]. Rao et al. [226,227] investigated self-sustained thermally driven oscillating in micro-channels. The motion of the meniscus generated a liquid film on the tube wall. The thickness of the liquid film and length at a given time determined the overall dynamics of the meniscus. Fourgeaud et al. [228] examined the variation of the liquid film thickness on the tube wall based on a single branch OHP. The film thickness was larger compared with the wedge-shaped film in a capillary OHP.

5. Bibliometric Study and Analysis

The OHP was first proposed in the 1990s and it has aroused wide attention in the academic circle. Scholars around the world have published a lot of research in this field. In order to understand the data from the relevant literature, we analyzed the literature on the oscillating heat pipe from the Web of Science Core Collection (WOSCC) during the 20 years from 2003 to 2022, through the research method of bibliometrics and through the data visualization analysis. Specific data are given as follows.

5.1. Publication Year and Number of Publications

In the WOSCC statistics, a total of 680 oscillating heat pipe literatures were published from 2003 to 2022. The number of papers published each year reflect trends in a particular field of research. As shown in Figure 16, the number of published articles in this field peaked at 65 in 2017. In the past two years, the number of published documents has fluctuated, but the overall trend is a slow rise.

5.2. Keyword Distribution

A keywords statistics of the above 680 literatures were carried out. A total of 41 keywords appeared more than 15 times. The top three keywords with the highest frequency are oscillating heat pipe, thermal performance, and flow, and the times are 135, 106, and 89, respectively. Figure 17 shows the connections among various words. The lines represent the number of times that these two keywords appear at the same time in a piece of literature. The thicker the lines are, the higher the number of times they appear together. The number of co-occurrences between these words is very high. Among them, the co-occurrence frequency of thermal performance and oscillating heat pipe is the highest, while the co-occurrence frequency of oscillating heat pipe and flow rank second. It can be seen that the OHP is closely related in the study of thermodynamics and fluid mechanics.

6. Current Research Shortcomings and Prospects for Future Works

6.1. Shortcomings of the Current Studies

(1)

Effects of channel layout

Studies of OHPs have focused on the effect of the fluid volume on the heat transfer performance, which neglected the synergistic effects of the layout optimization and internal flow [229].

(2)

Limitation of materials

The material of the OHPs tube wall affects the heat transfer performance and the layout plasticity of the OHPs [230]. The overwhelming majority of the tube wall materials in the current study are conventional materials of copper, aluminum, and stainless steel [231,232]. The thermal conductivity of copper is relatively high, and some composite metal tube wall materials may have better performances in OHPs.

(3)

Insufficient understanding of working fluid properties

The performance of the working fluid is the core material for OHPs, which directly affects the heat transfer [233]. Most of the working fluids studies have focused on water, ethanol, etc. [234,235]. For mixed nano-fluids as well as non-azeotropic mixture liquids, the effect of the concentration and mixing ratio on the optimal thermal performance of OHPs should be explored in depth. In addition, there is a lack of studies on the time-dependent properties of nano-fluid OHPs.

(4)

Inadequate recognition of the heat transfer mechanism

The operation mechanism of OHPs is complex. It is not beneficial to optimize the performance of OHPs with a single thermodynamic theory to explain and derive the mechanism [236]. The interaction and influence between vapor and liquid phases have not been deeply studied, which provides a lack of accurate predictive numerical models [237].

(5)

Limitations of the actual application environment

The superior heat transfer performance of OHPs should be focused on the practical use aspects. The performance of OHPs [238,239] in different use environments (e.g., rotating, centrifugal, antigravity, and horizontal) can vary significantly from operating under laboratory conditions. The actual application scenario environment will limit the operation under different working conditions [240].

6.2. Prospects for Future Research

(1)

Further study of mixed nano-fluids

The nano-fluids have great potential for effective heat transfer as OHP working fluids. Through the collaborative combination of different types of nano-particles, the mixed nano-fluids can achieve higher advanced thermophysical properties and stability of single nano-fluids [241]. The influence of the mixed nano-fluid on the thermal performance of OHPs can be further studied by the optimization, analysis, and improvement of the thermal efficiency of the mixed nano-fluid [242,243].

(2)

Study of non-azeotropic mixtures

Theoretical analysis shows that the non-azeotropic immiscible binary mixture can expand the operating temperature range of OHPs [244]. The non-azeotropic immiscible binary mixture can solve the problems of starting temperatures and the drying under the high heat flux of OHPs [245,246], which can improve the heat transfer performance and heat transfer limit of OHPs [247,248]. Future studies can be performed on non-azeotropic mixtures OHPs with different configurations by the analysis of the correlation among profile mixing ratio, filling ratio, and power input.

(3)

The improvement in the numerical model

VOF is a widely used method in order to study the complex coupling of pressure and temperature of OHPs, which can capture the phase distribution and interface dynamics [249,250]. Lee’s model is used to explain the phase transition principle. Numerical model improvements can be made to better cover the microscopic and macroscopic levels [251]. Dreiling et al. [252] proposed a closed-interface tracking CFD-VOF method. The effect of factors of the film around the bubble and the curvature of the interface, turbulence, or mass transfer strength parameters on the OHP can be systematically investigated. More stable and effective turbulence models that are applicable for OHPs need to be developed in the prospective.

(4)

The combination of artificial intelligence technology

Artificial intelligence techniques, in terms of efficiency and intelligence, have also been applied in the field of OHPs [253]. Many cases of research with deep learning algorithms optimize the parameters of OHPs [254]. With the development of AI technology, the computational work of optimization algorithms has become more comfortable. Wen et al. [255] modeled acetone OHP thermal resistance with human neural networks, and the R-squared of the models proposed with MLP and GMDH were 0.989 and 0.965, which were able to predict and simulate acetone-filled OHP thermal resistance. Jokar et al. [256] proposed OHP simulation and optimization with the genetic algorithm approach and the obtained the optimum filling rate of 38.25%.

7. Conclusions

This paper describes the various filling working fluids and in-tube flow states on the heat dissipation properties of OHPs. The effect of metal nano-fluids, non-metallic nano-fluids, mixed nano-fluids, gas working fluids, organic solvents, mixed liquids, and SRWF fluids are given. The different heating methods of the evaporation section, flow state, gravity effect, flow pattern, and oscillatory characteristics of the tube are illustrated. The following are the main conclusions.

(1)

With the addition of non-metallic nano-fluids in OHPs, the thermal resistance decreases from 24% to 83.6% with the change in the type, size, and concentration of nano-particles. The maximum heat pipe thermal conductivity was enhanced by 105% with the graphene nano-fluids. OHPs with gas as the working fluid can be used in the field of low temperature cooling. The effective thermal conductivity varies from 4.8 kW/(m·K) to 70 kW/(m·K) when different gases are selected as the working fluid in OHPs.

(2)

Compared with the pure working fluid, the thermal resistance of OHPs can be reduced by 68.9% with the right mixture type, filling rate, and mixing ratio. The surfactant and SRWF can be added to reduce the surface tension of the working fluid and the thermal resistance of the OHP can be reduced by 4.78% to 60%.

(3)

The change in gravity has a significant effect on the heat transfer performance of OHPs with vertical heating. A sensible heat transfer can account for more than 80% of the total heat transfer when the internal flow type of OHPs is the slug flow. The heat transfer performance is enhanced with the increase in the oscillatory flow, amplitude and the effective thermal conductivity can reach 40 kW/(m·K). The input power is increased from 7 W to 16 W, the oscillation amplitude is increased from 3.4 mm to 8.3 mm, and the thermal resistance is reduced by 13.5%.