# Flow Boiling Heat Transfer Performance and Boiling Phenomena on Various Straight Fin Configurations

^{*}

## Abstract

**:**

^{2}∙K, outperforming the trapezoidal fin in terms of heat transfer capability. As the hydraulic diameter reduces, the thermal boundary layer becomes denser, providing a more distributed saturated region. This leads to the increase in the heat transfer coefficient up to 22.5% and 17.1% for rectangular and trapezoidal fins, respectively. Additionally, the efficiency analysis shows that, albeit increasing the mass flux and reducing the gap increase the average cooling performance, but the pressure drop jumps up to 48%, reducing the efficiency of the heat removal system.

## 1. Introduction

^{2}, exceeding the single-phase capability [5]. This growth in cooling demand has been led by the miniaturisation and increase in computer power, resulting in the rapid growth of heat flux released from computer chip [6]. However, although the current cooling capacity of flow boiling is within the remit of industrial requirements, future heat generation in the computer system will exceed the capability of current two-phase cooling technology. Hence, to fulfil that demand, boiling heat transfer enhancement remains the core interest of the research in cooling technology.

_{h}of an Ω-shaped microchannel on the heat transfer performance. The study tested three microchannel variations with the D

_{h}of 590 μm, 781 μm, and 858 μm. The study showed that the microchannel with the D

_{h}of 781 μm provided a large heat transfer coefficient and a moderate pressure drop.

## 2. Materials and Methods

#### 2.1. Experimental Facility

_{e}) × 4 mm (w

_{e}) × 48 mm (h

_{e}) and channel dimensions of 60 mm (w

_{c}) × 18 mm (h

_{c}). To mimic the heat flux from the industrial application, total power of 2100 W was installed under the heating block. The supplied heat was controlled by a voltage regulator and monitored by a watt meter. Six k-type thermocouples were attached close to the surface of the heating block to estimate surface temperature. Meanwhile, two other k-type thermocouples used to measure fluid temperature were attached to the top of the channel. Furthermore, two pressure transducers were placed close to the inlet and outlet of the channel to measure the pressure drop caused by the system. Additionally, a transparent window was installed to observe and visualise the bubble dynamics during the experiment.

#### 2.2. Test Samples

#### 2.3. Measurement and Data Acquisition

_{s}), the measured data from six thermocouples were incorporated. By employing the steady-state conduction equation, heat flux and calculated thermal resistance of the modified surface, the uniform surface temperature was estimated as follows:

_{b}) is estimated by employing Equation (2) to account for the combined effects of uncertain parameters, specifically ${q}^{\u2033}$ (3.3%), ${T}_{s}$ (1.5%), and ${T}_{sat}$ (1.8%). Hence, the average uncertainty of h

_{b}is estimated at around 5.4%.

## 3. Results and Discussion

#### 3.1. Boiling Curve and Cooling Performance

^{2}∙K. Upon comparing the average heat transfer of different mass flux variations, the HTC of the R1 test case was 31.19% higher than T1.5. Both rectangular and trapezoidal fins demonstrated the enhancement of the heat transfer coefficient as the fin gap reduced. This was the result of the higher total area (see Table 1), whereby the 1 mm gap test cases had 1.1% and 1.2% wider areas for rectangular and trapezoidal fins, respectively. This small area increase, however, accounted for a significant enhancement in the heat transfer coefficient by 22.5% for rectangular and 17.1% for trapezoidal fins.

^{2}∙s to 19.7 kg/m

^{2}∙s. This pattern was anticipated since flow boiling relies on fluid flow to enhance the cooling performance. As the fluid velocity increases, the bubble lifts off easier from the surface, avoiding the concentration of thermal resistance due to bubble crowding on the boiling surface [29].

_{h}reduced, the thermal boundary layer within the vertical surface became closer. This resulted in a more saturated region across the flow cross-section. This led to more active nucleation sites and bubble generation across the boiling surface, which produced better boiling heat transfer performance. This mechanism is illustrated in Figure 6. This result supports the thermal boundary layer breakage concept proposed by Prajapati et al. [20]. The authors noted and depicted that certain fin designs result in an incomplete development of the thermal boundary layer, hindering the smooth flow of bubbles. Even though the low void fraction prevents the visual observation of this phenomenon, a similar mechanism can result in a uniformly distributed saturated region. The uneven distribution of the boundary layer causes a variation in heat transfer performance, as the Onset of the Nucleation Boiling (ONB) typically only occurs in the saturated region. However, this mechanism only occurs on the low void fraction and sub-cooled boiling regime, whereby the ONB is not yet observed in most locations.

_{h}. This means that even though the R1.5 test case had more contact area between the hot surface and working fluid, single-phase convection dominated the heat transfer process because a less saturated region was observed in the channel. Hence, the average heat transfer coefficient was lower. The summary of the average heat transfer coefficient in different D

_{h}values is presented in Table 3. The results clearly show a strong correlation between the average heat transfer coefficient and D

_{h}. As the D

_{h}increases, the heat transfer coefficient declines. In the present analysis, the determination of hydraulic diameter was carried out by employing a comparison between the cross-sectional area, taken perpendicularly to the direction of flow (A), and the wetted perimeter (P), as expressed by Equation (6). Herein, A and P are precisely defined at the uppermost section of the fin, where the fluid comes into contact.

#### 3.2. Boiling Phenomenon and Bubble Dynamics

^{2}∙s and 19.7 kg/m

^{2}∙s. A bubble detaches after sliding on the surface in the direction of the fluid flow. This bubble sliding is the distinction between the pool and flow boiling, whereby the bubble is unable to accumulate on a surface due to the fluid flow.

^{2}∙s and G = 19.7 kg/m

^{2}∙s, respectively) and was not linearly correlated with the fluid velocity. The results of the test case indicate that as the mass flux increases by half, the time taken for the bubble departure decreases by approximately 25%. This suggests that the effect of mass flux on bubble dynamics is non-linear and can have a significant impact on the overall heat transfer performance. As the fluid flows, the evaporative heat flux from the high mass flux is enhanced, allowing the average boiling heat transfer coefficient to rise. It should be noted that in this particular study, the observed vapour quality and Reynolds number were relatively small, ranging from 30 to 180. As a result, the impact of bulk fluid movement can be considered negligible.

#### 3.3. Enhancement Efficiency Analysis

^{2}∙s, respectively. The enhancement ratio of >1 means that the test case provided better cooling system efficiency than the reference test case, and vice versa. The enhancement ratios of all test cases are presented in Figure 9.

^{2}∙s. It is noteworthy that, although the heat transfer coefficient was at its lowest, the pressure drop in this test case was also the lowest. Compared to the rectangular fin with a gap of 1 mm, for instance, the heat transfer coefficient was 24% lower, whilst the pressure drop was 72% lower, resulting in the higher ‘efficiency’ of the enhancement. Meanwhile, the increase in the mass flux significantly raised the pressure drop up to 48%. On the other hand, the HTC enhancement of increasing the mass flux was restricted to below 30% (see Figure 5). Hence, the overall system efficiency reduced remarkably as the mass flux increased. This is a vital consideration in the design, whereby higher performance and high energy-saving efficiency should be considered together.

## 4. Conclusions

- The inclining trend of heat transfer coefficient with the rise in heat flux was indicated by the sharper boiling curves, whereby the lower surface temperature could be achieved at the same supplied heat flux. This was the result of the more bubbles being generated, increasing the contribution of the evaporative and quenching heat flux.
- The cooling performance improved as the fin gap reduced up to 22.5% and 17.1% for rectangular and trapezoidal fins, respectively. The hydraulic diameter had an important impact on the heat transfer coefficient, whereby the lower diameter contributed to the more distributed saturated region. As the hydraulic diameter decreased, there would be more covered area to fulfil the requirement of the ONB, resulting in better cooling performance by the bubble nucleation process.
- In general, the rectangular fin had a higher heat transfer coefficient than the trapezoidal fin. The effect of the total extended area was undermined by the effect of the hydraulic diameter. In order, the best test cases in terms of cooling performance were R1, T1, R1.5 and T1.5, with the highest observed at 5066.84 W/m
^{2}∙K. - Bubble sliding was observed during the boiling process. As the heat transfer coefficient increased, the bubbles were observed to exist in larger numbers but with a shorter period of detachment. It was found that there was no significant effect of mass flux on the bubble sliding distance.

## Author Contributions

## Funding

## Data Availability Statement

## Conflicts of Interest

## Nomenclature

a | Sample length (mm) |

A_{base} | Area of fin base (mm^{2}) |

A_{ext} | Area of fin extended surface (mm^{2}) |

A_{tot} | Total area of fin surface (mm^{2}) |

b | Sample width (mm) |

CHF | Critical heat flux |

D_{h} | Hydraulic diameter (mm) |

G | Mass flux (kg/m^{2}·s) |

h | Fin height (mm) |

h_{b} | Boiling heat transfer coefficient (W/m^{2}·K) |

P | Wetted perimeter (m) |

q″ | Heat flux (W/m^{2}) |

R_{E} | Enhancement ratio (dimensionless) |

s | Fin width (mm) |

T_{s} | Surface temperature (K) |

T_{sat} | Saturation temperature (K) |

Δp | Pressure drop (Pa) |

ΔT_{e} | T_{s} − T_{sat}, Excess temperature (K) |

σ | Uncertainty |

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**Figure 2.**(

**a**) Evaporator chamber of the experimental facility and (

**b**) schematic diagram and sensors placement.

**Figure 3.**(

**a**) Illustration and (

**b**) example of actual image of straight fin samples employed in this study.

**Figure 7.**Bubble (indicated by red dotted circles) dynamics of the (

**a**) R1, (

**b**) R1.5, (

**c**) T1 and (

**d**) T1.5 test cases at power input 170 W and G = 6.6 kg/m

^{2}∙s.

Test Case | Profile | s (mm) | h (mm) | A_{ext} | A_{base} | A_{tot} | A_{tot}/A_{base} |
---|---|---|---|---|---|---|---|

R1 | Rectangular | 1 | 0.95 | 3602.4 | 4779 | 8381.4 | 1.75 |

R1.5 | Rectangular | 1.5 | 0.925 | 3507.6 | 4779 | 8286.6 | 1.73 |

T1 | Trapezoidal | 1 | 0.95 | 2777.0 | 4779 | 7556.0 | 1.58 |

T1.5 | Trapezoidal | 1.5 | 0.925 | 2685.45 | 4779 | 7464.45 | 1.56 |

Properties | Values |
---|---|

Boiling point (K) | 334.15 |

Specific heat (J/kg∙K) | 1170 |

Latent heat vaporisation (kJ/kg) | 112 |

Thermal conductivity (W/m∙K) | 0.068 |

Liquid density (kg/m^{3}) | 1418.64 |

Vapour density (kg/m^{3}) | 0.98 |

Kinematic viscosity (m^{2}/s) | 3.008 × 10^{−7} |

Test Case | $\mathbf{Average}\mathbf{Heat}\mathbf{Transfer}\mathbf{Coefficient}$(h_{b}, W/m^{2}·K)
| Hydraulic Diameter (D_{h}, mm) |
---|---|---|

R1 | 4505.64 | 1.31 |

R1.5 | 3635.95 | 1.66 |

T1 | 4070.53 | 1.60 |

T1.5 | 3433.79 | 1.90 |

**Table 4.**Bubble (indicated by red dotted circles) lift-off mechanism in the R1 test case and input power of 170 W.

Time (s) | G = 13.1 kg/m^{2}∙s | G = 19.7 kg/m^{2}∙s |
---|---|---|

0.4 | ||

0.8 | ||

1.2 | ||

1.6 |

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## Share and Cite

**MDPI and ACS Style**

Pranoto, I.; Rahman, M.A.; Wicaksana, C.D.; Wibisono, A.E.; Fauzun; Widyatama, A.
Flow Boiling Heat Transfer Performance and Boiling Phenomena on Various Straight Fin Configurations. *Fluids* **2023**, *8*, 102.
https://doi.org/10.3390/fluids8030102

**AMA Style**

Pranoto I, Rahman MA, Wicaksana CD, Wibisono AE, Fauzun, Widyatama A.
Flow Boiling Heat Transfer Performance and Boiling Phenomena on Various Straight Fin Configurations. *Fluids*. 2023; 8(3):102.
https://doi.org/10.3390/fluids8030102

**Chicago/Turabian Style**

Pranoto, Indro, Muhammad Aulia Rahman, Cahya Dhika Wicaksana, Alan Eksi Wibisono, Fauzun, and Arif Widyatama.
2023. "Flow Boiling Heat Transfer Performance and Boiling Phenomena on Various Straight Fin Configurations" *Fluids* 8, no. 3: 102.
https://doi.org/10.3390/fluids8030102