Microchannel Heat Sinks—A Comprehensive Review

Abstract

An efficient cooling system is necessary for the reliability and safety of modern microchips for a longer life. As microchips become smaller and more powerful, the heat flux generated by these chips per unit area also rises sharply. Traditional cooling techniques are inadequate to meet the recent cooling requirements of microchips. To meet the current cooling demand of microelectromechanical systems (MEMS) devices and microchips, microchannel heat sink (MCHS) technology is the latest invention, one that can dissipate a significant amount of heat because of its high surface area to volume ratio. This study provides a concise summary of the design, material selection, and performance parameters of the MCHSs that have been developed over the last few decades. The limitations and challenges associated with the different techniques employed by researchers over time to enhance the thermal efficiency of microchannel heat sinks are discussed. The effects on the thermal enhancement factor, Nusselt number, and pressure drop at different Reynold numbers in passive techniques (flow obstruction) i.e., ribs, grooves, dimples, and cavities change in the curvature of MCHSs, are discussed. This study also discusses the increase in heat transfer using nanofluids and how a change in coolant type also significantly affects the thermal performance of MCHSs by obstructing flow. This study provides trends and useful guidelines for researchers to design more effective MCHSs to keep up with the cooling demands of power electronics.

1. Introduction

Microchips have evolved at a rapid pace over the years. In 1970, the first microchips that were introduced to the market were quite big, about 100 square millimeters, and had about 2000 transistors [1]. Microchips continued to become smaller, while transistor counts increased, during the decade of the 1990s. A chip might consist of a few hundred thousand up to a million transistors. Microchips’ sizes also greatly decreased, with average sizes of around a few or tens of square millimeters. Further developments in transistor scaling and miniaturization were made in the 2000s. Transistor counts on microchips reached the millions to tens of millions. Microchips started crossing the 100 million transistors mark in the 2010s as a result of advanced manufacturing processor and transistor architectures, reaching billions of transistors on a single chip. Even smaller microchips now frequently have dimensions in the single-digit square millimeter range.

In 2018, the Intel Corporation introduced a 10 nm processor, which had over 100 million transistors but only 100 square millimeters of surface area [2]. The size miniaturization of microprocessors, along with the simultaneous increase in computing power, is a significant advancement in microprocessor technology [1]. However, this miniaturization and the rise in the computing power of microchips have resulted in a sharp increase in the heat flux generated by these microchips, as shown in Figure 1 [3].

Due to a lack of heat dissipation, this high heat flux develops hotspots in chips, which present a great threat to the safety of the microchip itself and reduce its life span. Thus, thermal management is a crucial aspect for designers to ensure the safe operation of these microchips for a longer duration [4]. Also, with the size of mechanical and electrical components changing from micrometers to nanometers, the efficient and sustainable cooling of these components is essential to ensure the seamless function of these microsystems. For the better and optimum functioning of microsystems, advanced cooling systems have been developed that are more effective than conventional cooling technologies. The rapid development in MEMS technology has inspired researchers to design novel cooling systems that dissipate more heat efficiently compared to conventional cooling systems [5]. As the miniaturization and power density of electronic devices have increased, cooling systems have evolved from fan cooling to more extensive heat transfer processors to withstand severe heat fluxes [6]. Several methods have been studied and developed, including microjet impingement, micro heat pipes, micro-electrohydrodynamics, and MCHSs. Among all these cooling methods, MCHSs are best for removing flux from microchips [7,8].

The arrangement or design of heat sinks within a thermal management system is referred to as heat sink topology in Figure 2. Here, the pie chart’s sections are divided into distinct categories. The term flow-based refers to heat sink topologies that optimize the passage of the cooling medium. The thermal performance-based category includes heat sink topologies created to maximize thermal performance and heat dissipation using materials with more conductive ability. Heat sink topologies that make use of special geometries to produce effective cooling are represented by the new geometry-based topology. The comparison-based section indicates a portion of the pie chart that compares or analyzes different heat sink topologies based on various factors, such as cost-effectiveness, thermal efficiency, or ease of implementation. The design-based section includes heat sink topologies that are determined by certain design factors, such as space limitations or system or device compatibility.

Figure 3 shows the flow type that researchers have used to study the thermohydraulic parameters of MCHSs.

Microchannel Heat Sink (MCHS)

The MCHS was introduced by D.B. Tuckerman and R.F.W. Pease at Stanford University in 1981 [11]. Since then, various academics have conducted extensive research on MCHSs [12]. The MCHS is a heat sink that has micro extrusions that allow the working fluid to flow through it for heat transfer in electronic devices [Figure 4]. Microchannel heat sinks can generate very high heat transfer rates, with an increase in heat transmission because of a small passage size and large area of surface to volume ratio [4].

To achieve the least pressure losses and thermal resistance, a few design variables must be taken into account while designing MCHSs. These include the material of the MCHS, the number of pin fins, shape, and fin alignment [5]. Material selection is primarily based on the thermal conductivities of materials [14]. An MCHS is made of good thermal conductors like copper and aluminum. Copper costs more in terms of money than aluminum, but it is roughly twice as conductive compared to aluminum and hence more efficient [15]. Aluminum can easily be formed by extraction and has less weight, and hence faces less stress on fragile parts. Because of these properties of aluminum, complex cross-sections can be made. Also, zinc is an acceptable material for heat sinks. The casting process when zinc is mixed with an alloy degrades its porosity, unlike copper and aluminum. But the thermal conductivity of zinc is lower than both copper and aluminum [15]. The inclusion of fins has a major effect on the effectiveness of MCHSs. Most heat exchangers require a large heat transfer per volume ratio for electronics cooling; as the number of fins increases, this ratio will correspondingly rise [16]. However, increasing the quantity of fins also increases the amount of pressure loss, which requires extra pumping power [17]. Different shapes of fins i.e., pin fins, wavy fins, straight fins, fluted fins, etc., are used in MCHSs [18]. Other non-standard fins are also present, and new geometries are also being designed. These non-standard fin shapes are unique or specialized arrangements made to meet certain thermal management requirements or to solve particular heat dissipation challenges. The most commonly used pin fins include square, hexagonal, cylindrical, and elliptical [18]. Straight fins with rectangular cross-sections are also very common.

A forced convection system is one that uses a pump to aid with fluid flow [19]. In that case, the cross-sectional area for flow, volumetric flow rate, and pressure drop are all system limitations. Otherwise, these are the design specifications. Resistance to the coolant movement is called pressure drop. It does not create much difference, but still, the selected MCHS shape changes the overall pressure losses of the system [20]. So, the MCHS selected should have lower pressure losses than the pump delivery pressure. The turbulence of flow is determined by the flow rate of the incoming liquid. When a fan is specified, its flow regime and fluid velocity are known [21]. Thus, increasing the velocity at the entrance raises the volumetric flow rate, which in turn promotes turbulence. If the flow is laminar with the given geometry, the thermal resistance will be enhanced. Specifically, for an entrance velocity of 2.5 m/s and a channel length of 50 mm, the thermal resistance decreases from 0.035 K/W to 0.022 K/W, demonstrating the significant impact of flow velocity on thermal performance. The cross-sectional area for flow can be designed based on the rate of flow requirements. If the cross-sectional measurement of the channel becomes smaller, more input is required to pump fluid at the required rate [20].

2. Methods Used to Enhance the Thermal Efficiency of MCHSs

As the heat flux removal requirements for MCHSs rise, several strategies are developed to improve the MCHS’s capability of dissipating more heat effectively. These methods can be divided into two categories: active methods and passive methods [22]. Figure 5 illustrates the details of these strategies.

2.1. Active Techniques

Active methods increase the thermal efficiency of smooth rectangular MCHSs by utilizing outside energy sources [8]. Go JS et al. [23] investigated the effects of vibration induced by flow on thermal performance in 2003 in a micro fin array. They reported that a larger heat exchange rate was achieved through raising the micro fin’s vibration displacement because of vibration-induced chaotic convection. They compared the thermal performance based on the thermal resistance encountered in plain wall and micro fin array heat sinks. The evaluated heat transfer rate was found to be an increase of 5.5 and 11.5% at air velocities of 4.4 and 5.5 m/s, respectively. Hessami et al. [24] investigated the impact of the flow pulsation technique on smooth MCHS performance. Their research revealed that raising the frequency and decreasing the intensity of flow pulsation led to improved heat removal.

Active techniques are those techniques which use an external power source or energy for thermal performance enhancement in microchannel heat sinks. The major reason to use both or either active techniques or passive techniques is to disrupt the flow and secondary flow generation. Due to their compact size, active techniques have been less often in microchannels comparing to passive techniques. Different approaches have been used by researchers, like electrostatic forces, flow pulsation, and vibration, etc. Krishnaveni et al. [25] proposed a rectangular MCHS by inducing chaotic mixing in the flow through the application of an electrostatic field. Morini et al. [26] numerically studied the electro-osmotic flow (EOF) in MCHSs with rectangular and trapezoidal cross-sections, using the Poisson–Boltzmann and Navier–Stokes equations. The results demonstrated the Reynolds numbers are in general low for EOF and for this reason: in the analysis of the heat transfer, the contribution related to axial conduction along the silicon wafer can be important because it changes the axial distribution of the heat flux. Han et al. [27] have evaluated the thermal performance of pure electro-osmotic flow (EOF), pure pressure-driven flow (PDF), and combination flow (CF) using numerical simulations. They observed a notable disruption in fluid motion with the aid of EOF, particularly in smaller hydraulic diameter MCHSs.

Narrein et al. [28] numerically investigated a helical-shaped microchannel heat sink with working fluid (Al₂O₃/water). They compared the performance for steady and pulsatile flow, and it was observed that pulsatile flow has a better heat transfer with a marginal reduced pressure drop. Nandi and Chattopadhyay [29] implemented pulsatile flow in a microchannel in a wavy shape. They concluded that flow pulsation was able to enhance heat transfer with a reduced pressure drop even at a low range of Re.

2.2. Passive Techniques

Passive techniques rely on natural mechanisms like convection, conduction, and radiation to remove heat, without the need for mechanical devices like fans or pumps. These techniques facilitate heat transfer using the built-in properties of the heat sink and its immediate surroundings. In MCHSs, passive cooling techniques are used to dissipate heat in various applications, including the electronics, automotive, and aerospace sectors. Through a decrease in or elimination of the dependence on active cooling components, they provide the benefits of silent operation, less power consumption, and increased system reliability. Some of the passive techniques for heat transfer in MCHSs are discussed below.

2.2.1. Flow Disruption Technique

One strategy for improving heat transmission in MCHSs is fluid flow obstruction. This technique involves inserting small obstacles into the flow channel to impede the flow and create turbulence. Reducing the thickness of the thermal boundary layer and raising the convective heat transfer rate lead to improved heat dissipation. MCHSs can use various flow obstacles, comprising micro fins, micro pin fins, micro dimples, micro cavities, micro ridges, etc. These microstructures are made to create swirls and vortices, which encourage heat transmission. The effectiveness of the flow obstruction method depends upon several factors, including flow rate, the thermal characteristics of the fluid, and the obstruction size, shape, and distribution.

2.2.2. Ribs

One common method of flow interruption on MCHS walls is the insertion of ribs. On the channel side walls, the small projections known as ribs can be positioned in several configurations, including parallel or staggered patterns. The flow turbulences created by these ribs reduce the heat of the thermal boundary layer and raise the heat transfer coefficient. Julian Wang et al. [30] investigated the influence of microscale ribs and grooves on the heat flux rate. They observed that, in comparison to a smooth microchannel, the Nusselt number is increased by 1.1–1.55 times when ribs and grooves are added to the microchannel. The friction factor for a rib-grooved microchannel with compared rib heights of 0.6, 0.73, and 0.85 is more than 0.38 times, 2.16 times, and 4.09 times higher than the smooth microchannel, respectively, according to their investigation into the impact of altering the relative rib height on the friction factor. Xia et al. [31] studied the effect of arc-shaped grooves and ribs on the thermohydraulic properties of a microchannel. Statistical methods were utilized in conjunction with the two target functions and the design elements of relative groove altitude, relative rib height, and relative rib width to reduce thermal resistance and pumping power at a fixed volume flow rate. It was found that relative rib height showed the highest performance, with overall thermal resistance decreasing linearly as rib height increased. The impacts of various cylindrical ribs and cavities were studied on the thermal and flow characteristics of MCHSs by Faraz Ahmad et al. [32]. For rib spacing (Sr) = 0.4 mm, the side wall ribs had the maximum thermal enhancement factor of 0.98, while the base wall ribs had the lowest thermal enhancement factor of 0.90. The all wall rib arrangement achieved a Nusselt number of 12.58, which is 71% higher than the Nusselt number of 7.35 for the smooth MCHS.

Ahmad, F et al. [12] studied the influence of various arrangements of ribs and cones on the walls of MCHSs and found that conning at 45 degrees decreases the pressure drop by 85%, while reducing the maximum Nusselt number by 25%. Furthermore, it was discovered that hexagonal ribs were the most effective of the various layouts considered. Shahzad Ali et al. [33] analyzed the outcome of adding different arrangements of hydrofoil ribs on the thermo-fluid characteristics of an MCHS. They found the best performance was obtained at rib spacing (Sr) = 0.4 mm, and the staggered arrangement performed better than the aligned ribs configuration. A maximum performance evaluation factor of 1.06 was found in the case of the aligned configuration at Re 1000. Aatif Ali Khan et al. [34] numerically analyzed MCHSs with six configurations of ribs. It was concluded that circular ribs have the lowest and rectangular ribs have the greatest pressure drop, while triangular–circular ribs have the highest thermal performance factor. They noted that triangular ribs have the least thermal resistance across a range of Reynolds numbers and rectangular ribs have the highest thermal resistance. Yao Hsien et al. [35] investigated the influence of ribs and grooves on heat transfer and friction in a rectangular microchannel. They found that the convective heat transfer coefficient was enhanced 1.40 times, and the friction factor ratio increased by up to 1.30 times. They concluded that the discrete angled arrangement of grooved ribs exhibited the best thermal performance among the configurations studied. Lau et al. [36] studied the effect of staggered ribs on turbulent heat transfer in a microchannel. It was concluded that discrete ribs with rib angles-of-attack of α = 60 degrees and 90 degrees create a very high heat transfer from ribbed walls, while discrete ribs with angles-of-attack of α = 45 degrees possess the greatest thermal efficiency. The thermal performance of cooling channels in turbine airfoils is improved by replacing angled complete ribs with these discrete ribs. Shizhong Zhang et al. [37] studied the effect of trefoil ribs walls in MCHSs, and side wall trefoil ribs provided the greatest performance factor, measuring 1.60. The impact of bidirectional ribs on the heat transfer performance of microchannel heat sinks was studied by Guilian Wang et al. [38]. It was concluded that adding bidirectional ribs increases the Nusselt number by 1.4–2 and 1.2–1.42 times over vertical and spanwise ribs at the expense of an increased pressure drop. Sadiq Ali et al. [39] analyzed the thermohydraulic characteristics of MCHSs with trefoil-shaped ribs. They studied the various trefoil rib shapes and concluded that side wall trefoil ribs (MC-SWTR) had the highest thermal enhancement factor of 1.6, while the all wall trefoil ribs (MC-AWTR) had the lowest thermal enhancement factor of 0.87 at Re 1000.

M. M. U. Rehman et al [36] analyzed the influence of unique side wall ribs on the heat transfer and flow characteristics of MCHSs. Their research showed that hydrofoil ribs had the highest heat transfer improvement of 26% from various geometrical rib profiles, and as a result, the pressure drop of the hydrofoil ribs was also enhanced by 66% as compared to the smooth MCHSs. The effect of sinusoidal cavities and rectangular ribs on the heat transfer augmentation of a microchannel heat sink was investigated by Ihsan Ali Ghani. Et al [40]. It was concluded that the maximum overall performance enhancement factor of 1.85 with Re = 800 was obtained with a cavity amplitude of 0.15, rib width of 0.3, and rib length of 0.5. Lei Chai et al. [41] investigated the influence of rectangular ribs on the heat transmission properties of an interrupted microchannel and determined the best rectangular rib parameters. It was discovered that the new interrupted microchannel with a length of 0.5 mm had a maximum enhancement factor of 1.35, while MCHSs with a length of 0.1 mm had the lowest enhancement factor of 1.12. Y.F. Li et al. [42] investigated the effects of triangular and rectangular cavities and ribs on laminar flow parameters in microchannel heat sinks. They found that a microchannel with triangular cavities and rectangular ribs with a rib width of 0.3 and cavity width of 2.24 gained a thermal enhancement factor of 1.619 at Re = 500. Y.L. Zhai et al. [21] evaluated the performance of three types of microchannels and noted that the microchannel with circular ribs had the maximum thermal enhancement factor of 1.5 at Re = 600, while the microchannel with rectangular ribs had the lowest thermal enhancement factor of 1. Ayodeji S. Binuyo [43] investigated the effect of an interrupted microchannel on an Al₂O₃/water-based nanofluid. The study concluded that increasing the pulse amplitude increased the Nusselt number, and increasing the pulsating frequency improved the heat transfer enhancement compared to decreasing the pulse amplitude.

Abdelkader Korichi et al. [44] investigated the impact of heated obstructions placed on the top and bottom walls of a microchannel. Their analysis revealed that the presence of these obstacles enhanced heat transfer by up to 123.1% when increasing the Reynolds number from 50 to 500, and by 48.5% when increasing it from 500 to 1000, as indicated by the overall Nusselt number. Aparesh Datta et al. [45] studied the influence of triangular cavities and ribs on microchannel heat transfer performance. It was reported that back-to-back cavities with widths 0.75 times and lengths 0.0454 times the basic microchannel provided the best thermal performance. They also observed that a copper MCHS has the highest thermal performance of 2.31 and a silicon MCHS has a thermal performance 1.96. Faraz Ahmad et al. [46] examined a microchannel heat sink’s conjugate heat transfer capabilities with side wall ribs. It was concluded that elliptical ribs had the maximum Nusselt number because of a greater streamlining effect and that more than 96% of the losses were due to heat transfer in the microchannel heat sink.

Q. Zhu et al. [47] analyzed the effect of rectangular grooves and rectangular ribs on the thermohydraulic characteristics of MCHSs. Different configurations of grooves and ribs were investigated, and it was found that a grooved channel with rectangular ribs gave the optimum efficiency with a rib width of 0.25. In order to maximize the Nusselt number while maintaining the lowest possible friction factor, several rib designs were used. The Nusselt number and pressure drop were found to be significantly impacted by the spacing between the ribs, the angle at which the ribs were installed, and the geometry of the ribs (e.g., circular, cone shaped, micro-scale, trefoil ribs, and hexagonal ribs) [12,30,32,34,36]. The Nusselt number increases when the rib spacing decreases, but the pressure drop also rises as a consequence. Placing ribs at angles minimizes fluid flow resistance, which is required to overcome the problem of pressure drop.

The following tables serve to provide a concise overview of the nature of the research work and the key parameters related to the MCHS.

Table 1. Additional data for ribs.

S. No	Authors	Geometry	Nature of Work	Substrate Material	Coolant Type	Key Findings
1	D. B. Tuckerman et al. [11]	Rectangular	Experimental	Silicon	Water	Maximum thermal resistance of 0.1 °C/W for 1 cm2 area Thermal resistance depends on water flow rate
2	Guilian Wang et al. [30]	Rectangular with microscale ribs and grooves	Experimental and Numerical	Silicon	Water	Nusselt number is 1.55 times and friction factor is 4.09 greater than smooth channel
3	Y. L. Zhai et al. [21]	Rectangular with rectangular, trapezoidal, and circular ribs	Numerical		Water	Maximum Nu value of 15 at Re 600 Thermal enhancement factor of 1.6 Field synergy number of 14 Entropy augmentation number of 0.94
4	G. D. Xia et al. [31]	Rectangular with arc-shaped ribs and grooves	Numerical	Silicon	Water	Relative rib height shows best performance in terms of thermal resistance
5	Faraz Ahmad et al. [32]	Rectangular with cylindrical ribs and cavities	Numerical	Silicon	Water	Maximum Nu value of 20 at Re 1000 Maximum thermal enhancement factor of 1.02
6	Shahzad Ali et al. [33]	Rectangular with hydrofoil ribs	Numerical	Copper	Water	Nu value of 15 at Re 1000 Thermal enhancement factor value of 1.07
7	Aatif Ali Khan et al. [34]	Rectangular heat sink with different rib configurations	Numerical	Silicon	Water	Rectangular ribs show greatest pressure drop of 100 kPa Nu value of 8.5 at Re 500 Thermal enhancement factor of 0.91
7	Yao Hsien et al. [35]	Rectangular with ribs and grooves	Experimental	Copper	Water	Heat transfer coefficient increased by 40% (Nu value of 300) Thermal enhancement factor of 2.8 for discrete ribs and grooves
9	Lau et al. [36]	Square with staggered ribs	Experimental	Copper	Water	Discrete ribs show superior performance than ribbed walls
10	Shizhong Zhang et al. [37]	Rectangular with trefoil-shaped ribs	Numerical	Copper	Water	Maximum Nu of 30 at Re 1000 Thermal enhancement factor of 1.5 Maximum entropy augmentation number of 0.72
11	Guilian Wang et al. [38]	Rectangular with bidirectional ribs	Experimental and Numerical	Silicon	Water	Thermal enhancement factor greater than 1 Nu 1.42 times higher with bidirectional ribs (Nu value of 13)
12	Sadiq Ali et al. [39]	Rectangular with trefoil-shaped ribs	Numerical	Copper	Water	Maximum pressure drop of 60 kPa Maximum Nu value of 29 Thermal enhancement factor of 1.60 at Re 1000
13	M. M. U. Rehman et al. [48]	Rectangular with side wall ribs	Numerical	Copper	Water	Maximum thermal enhancement factor of 1.10 Maximum average heat transfer coefficient value of 19 at Re 1000 Highest pressure drop of 19 kPa at Re 1000
14	Ihsan Ali Ghani et al. [40]	Rectangular with rectangular ribs and sinusoidal cavities	Numerical	Silicon	Water	Performance factor value of 1.83 at Re 800 Average Nu value of 24 at Re 1000
15	Lei Chai et al. [41]	Interrupted with rectangular ribs	Numerical	Silicon	Water	Thermal enhancement factor of 1.3 Interrupted channel without ribs shows high heat transfer
16	Y.F. Li et al. [42]	Rectangular with rectangular ribs and triangular cavities	Numerical	Silicon	Water	Thermal enhancement factor of 1.6 at Re 500 Entropy augmentation number of 0.95 Average Nu value of 17
17	Ayodeji S. Binuyo [43]	Interrupted	Numerical	Silicon	Water	Highest Nu value of 14 Average friction factor value of 0.08
18	Abdelkader Korichi et al. [44]	Rectangular with heated obstacles	Numerical			Highest Nu value of 23 Increase in friction factor by factor of 19
19	Aparesh Datta et al. [45]	Microchannel with triangular cavities and ribs	Numerical	Silicon	Water	Highest Nu value of 18 Highest friction factor value of 0.80 Performance factor value of 1.45
20	Faraz Ahmad et al. [46]	Rectangular with side wall ribs	Numerical	Copper	Water	Base and side wall ribs show superior performance than all wall
21	Q. Zhu et al. [47]	Rectangular with rectangular grooves and different-shaped ribs	Numerical	Silicon	Water	Highest pressure drop value of 119 kPa Average Nu value of 22 Maximum thermal efficiency of 145%

describes the additional data for ribs as a flow obstruction technique. Explanation of computational domains is shown in Figure 6, Figure 7, Figure 8, Figure 9, Figure 10, Figure 11, Figure 12, Figure 13, Figure 14, Figure 15, Figure 16, Figure 17 and Figure 18.

2.2.3. Grooves

Another technique for improving the thermal performance of MCHSs is the introduction of grooves in MCHSs. Grooves can effectively increase the thermal characteristics of MCHSs. Grooves help to raise the surface area of the MCHS, which allows for a greater amount of heat dissipation. The increased surface area allows for more contact with the cooling medium, and fluid flow also increases due to the grooves in the MCHS, which can help to remove heat more efficiently. The grooves can create turbulence in the fluid flow, which can promote better mixing and heat transfer. Overall, the inclusion of grooves in MCHSs can lead to improved thermal performance, making them more effective at dissipating heat and cooling electronic devices. Hamdi E. Ahmed et al. [49] investigated the impact of design factors on grooved microchannel heat sinks (MCHSs). Their analysis showed a 2.35% increase in the friction factor, from 0.183 to 0.187, and a 51.59% improvement in the Nusselt number, from 22.4 to 33.9. These changes led to the most effective thermal design for MCHSs. Pankaj Kumar [50] studied heat transfer in trapezoidal microchannels with Reynolds numbers from 96 to 720. The study found a 12% increase in heat transfer efficiency for trapezoidal channels compared to rectangular ones, and a 28% enhancement with grooves in the trapezoidal channels. The simulations were conducted with a constant heat flux of 100 W/cm². The geometry with semicircular grooves had an extra pressure drop from frictional loss and 16% higher heat transfer than rectangular grooves. Guodong Xia et al. [51] numerically investigated MCHSs with triangular reentrant cavities to find the impact of geometric factors on heat transfer. Four design variables were taken for the optimum thermal design. For Re = 406.94, and Re = 611.25, 814.09, and 1015.96, the optimal parameters are discussed. The authors of [52] studied a grooved channel with curved vanes using holographic interferometry. The heat transfer was raised by factor of 1.5 to 3.5 around Re = 450, and the pressure drop increased by 3 to 5 times as compared to smooth MCHS. A range of groove structures, such as curved vanes [52], semicircular grooves, rectangular and trapezoidal grooves [45], and triangular reentrant cavities [51] are covered in this section. Semicircular grooves improve heat transmission at the expense of a higher pressure drop. Furthermore, variations in the Reynolds number also influence the performance of the grooves [51].

Table 2. Additional data for grooves.

S. No	Author	Nature of Work	Geometry	Coolant Type	Substrate Material	Key Findings
1	Pankaj Kumar [50]	Numerical	trapezoidal MCHS with groove structure	Water	Silicon	Performance factor of 0.62 Friction factor of 0.026 Maximum pressure drop of 140 kPa
2	Hamdi E. Ahmed et al. [49]	Numerical	grooved MCHS	Water	Aluminum	Maximum performance evaluation factor of 2 Average Nusselt number ratio of 1.9
3	Guodong Xia et al. [51]	Numerical	MCHS with triangular reentrant cavities	Water	Silicon	Highest thermal enhancement factor of 1.6 Friction factor value of 73
4	Cila Herman et al. [52]	Experimental	grooved MCHS with curved vanes	Air	Copper	Pressure drop increase by 3–5 times more than smooth channel Heat transfer increase by 1.5–3.5 times more than smooth channel

gives detailed information about the research work carried out on grooves. Explanation of computational domains is shown in Figure 19 and Figure 20.

2.2.4. Cavities and Dimples

Utilizing cavities and dimples on the surface of MCHSs is another method for improving heat transfer. Vortices and swirls in the flow are produced by cavities and dimples, which cause mixing and improve heat dissipation. Minghai Xu et al. [53] investigated the thermal performance of microchannel heat sinks with aspect ratios of 1:1 to 4:1 and dimple configurations (0.2 mm radius, 0.1 mm depth) at a Reynolds number of 500. The results show a significant increase in the average Nusselt number from 8.21 to 9.44 (15%) with dimples, and a 2% reduction in pressure drop, demonstrating enhanced heat transfer efficiency and reduced fluid resistance. Yu Chen et al. [54] conducted numerical investigations over the dimpled surface with turbulent channel flow. The Reynolds number ranged from 4000 to 6000. However, the Prandtl number remained unchanged at 0.7. The results showed an asymmetric dimple with a 15% stream-wise skewness and a depth ratio of ℎ/D = 15%. When compared to a symmetrical dimple with the same depth ratio, heat transmission increased considerably, while the pressure loss remained unchanged.

A microchannel with turbulent flow over a dimpled surface was experimentally studied by Moon et al. [55]. The thermochromic liquid crystal (TLC) technique was used to evaluate the local heat transfer coefficient. The channel heights of 0.37, 0.74, 1.11, and 1.49 were examined in a range of Reynolds numbers from 12,000 to 60,000. The Nusselt number of dimpled walls is 2.1 times that of a conventional channel, yet the friction factor is only 1.6 to 2.0 times that of a conventional channel in a thermally developed area. S. Gururatana [56] investigated a microchannel heat sink (MCHS) with dimples and found that the dimples have a significant impact when the Reynolds number exceeds 125. Specifically, when the Reynolds number increases from 100 to 300, the pressure drop rises from 0.1 to 0.4, while the heat transfer increases from 0.2 to 0.26. Heat transfer is enhanced by the turbulence generated by dimples and cavities. In this section, different studies are summarized, which include the effects of dimples at different Reynolds numbers, dimples with varying aspect ratios [53], turbulent flow channels [54], thermochromic liquid crystal (TLC) technology to analyze dimples and cavities [55], and varying channel heights. Heat transmission rises in response to an increase in Reynolds number [56].

Table 3. Additional data for cavities and dimples.

S. No	Author	Nature of Work	Geometry	Coolant Type	Substrate Material	Key Findings
1	Mohib-ur-Rehman et al. [57]	Numerical	MCHS with hemispherical shape protrusions/dimples	Water	Copper	Maximum pressure drop value of 50 kPa Friction factor value of 3.5 Entropy augmentation number of 0.85
2	Minghai Xu et al. [53]	Numerical	MCHS with dimples	Water	Copper	Highest Nusselt number value of 20 Performance improved by increasing number of dimples
3	Yu Chen et al. [54]	Numerical	MCHS with turbulent flow over dimpled surface	_	_	Highest performance ratio of 2.8
4	Moon et al. [55]	Experimental	Rectangular MCHS with concavities	_	_	Normalized Nu value of 2.1 Thermal enhancement factor of 1.75
5	Suabsakul Gururatana [56]	Numerical	Rectangular MCHS with dimpled surfaces	Air	_	Maximum pressure drop of 8 Pa at Re 350 LOCal Nu value of 17 Performance factor of 1.016

gives additional information for cavities and dimples as a flow disruption technique. Explanation of computational domains is shown in Figure 21, Figure 22, Figure 23, Figure 24 and Figure 25.

2.2.5. Micro Fins

Another method of disrupting the flow in microchannel heat sinks is micro fins. Small protrusions called pin fins are positioned in the flow route. The pin fin’s turbulence in the flow enhances heat transmission by expanding the surface area available for heat transfer and promoting mixing. Dogan et al. [16] experimentally studied the heat transfer using a horizontal microchannel with longitudinal fins with varying heights and spacing, for an extensive range of altered Rayleigh numbers. Their findings indicate that maximal heat transmission occurred at the spacing of S = 8–9 mm, and that the ideal spacing relies on the Rayleigh number. Nawaz Khan et al. [17] numerically analyzed the different configurations of fins with varying heights mounted at the channel’s base wall. The study found that when the Reynolds number was between 150 and 350, the maximum Nusselt number was obtained by using full-length fins upstream, and when the Reynolds number was higher, the maximum Nusselt number was obtained by using full-length fins at the microchannel’s inlet and exit. The lowest pressure drop was achieved using a full-length fin at the center of the microchannel. An offset strip fin microchannel was studied numerically for convective heat transfer by Minghai Xu et al. [53]. The geometric characteristics of the strip fins, such as fin thickness (W_w), channel width (W_c), fin depth (H_c), fin interval (L1), and fin length (Ls), were used to study the efficiency of an offset strip fin MCHS. Two dimensionless numbers were used to find the effect of the fins on pressure drop, which were K = L₁/L_s and M = L/(L_s + L₁). The results concluded that K = 1 was optimal for the better heat transfer of an offset strip fin MCHS. Ping Li et al. [58] numerically studied and optimized a dimpled and pin-finned water-cooled microchannel. The Reynold’s number ranged from 50 to 300, and the findings demonstrated that for all Reynolds values, the Nu increased with an increasing pin diameter and decreasing separation. The thermal performance was a maximum of 10.35 at Re = 200. Wazir et al. [59] studied the effect of full-length round pin fins on the thermal and flow characteristics of MCHSs. They found that MCHSs with an MC (mixed configuration) achieved a highest thermal enhancement factor of 1.4 and a Nusselt number increase 2.3 times that of smooth MCHSs. Micro fins with different shapes, heights, and spacing [17] have been analyzed. It has been found that having full-length fins in the middle of the channel lowers the pressure drop, and a full-length pin fin at the start and at the end improves the heat transfer. Additionally, the Rayleigh number determines the appropriate fin spacing in the microchannel heat sink [16].

Details are given in

Table 4. Additional data for micro fins.

S. No	Author	Nature of Work	Geometry	Coolant Type	Substrate Material	Key Findings
1	Nawaz Khan et al. [17]	Numerical	Rectangular MCHS with pin fins configuration of varying height	Deionized ultra-filtered water	Copper	Highest pressure drop of 7000 Pa Nu value of 13 Thermal enhancement factor of 1.4
2	Dogan et al. [16]	Experimental	Rectangular MCHS with longitudinal fins	Air	Silicon	Maximum convection heat transfer coefficient value of 27 Highest Nu value of 250
3	F. Hong et al. [60]	Numerical	Rectangular MCHS with offset strip fin	Water	Silicon	Pumping power value of 0.21 W Maximum presssure drop value of 120 kPa
	Ping Li et al. [58]	Numerical	Rectangular MCHS with dimple and pin fin	Water	-	Performance factor of 1.9 Highest Nu ratio value of 2.3

on the application of micro fins in microchannel heat sinks.

2.2.6. Channel Curvatures

Changing the channel curvatures is another method for enhancing the MCHS’s thermal performance. The curvature of the channels affects the thermal performance of microchannel heat sinks. Centrifugal forces are experienced by the fluid flowing through a curved channel, which may enhance the fluid velocity along the channel’s outside walls. This higher velocity encourages a more efficient heat transfer between the fluid and the channel walls, enhancing the heat sink’s total heat transfer capabilities. The channel walls’ surface area may also be increased by the curvature, increasing the area where heat can be transferred. Chu et al. [61] examined four sets of triangular microchannels experimentally and noted how heat transmission and pressure drop were affected. Reynold’s number increases cause the friction factor to drop nonlinearly. The Nusselt number increases as the Reynold number increases in the low Reynold number zone, which is less than 100. Following Re = 30, the Nusselt number increases slowly. The correlation between the numerical and experimental results is within 15%. Deng et al. [62] found that in a reentrant microchannel heat sink, the Nusselt number increased by up to 39% from 10 to 13.9, and thermal resistance decreased by up to 22% from 0.21 C/W to 0.16 C/W, within a Reynolds number range of 150 to 1100. The pressure drop and heat resistance produced by the reentrant microchannels were similar to or slightly larger than those of their regular rectangular counterparts. Wu H. et al. [63] experimentally investigated thirteen different trapezoidal silicon microchannels. The Nusselt number and friction constant rose as the surface roughness and hydrophilic characteristic increased. The Nusselt number increases more slowly when the Reynolds number exceeds 100.

A wavy microchannel heat sink within the range of Reynolds numbers from 100 to 1000 was investigated using the finite volume method (FVM) by Deng, B et al. [64]. The amplitudes of the wavy microchannel ranged from 125 to 500 μm. The results showed that wavy microchannels outperform straight microchannels with the same cross-section in terms of heat transmission. The improvement in heat transmission outweighs the pressure drop penalty resulting from the wavy microchannels. The amplitude of the wavy microchannel grows as heat transport occurs. Except for microchannels with a wavy amplitude of 0.25, all the wavy microchannels had a better performance than traditional rectangular microchannels. The impact of geometrical factors on heat transport in microchannels for Reynolds numbers 100–1000 was investigated numerically by Gunnasegaran P. et al. [65]. A step MCHS was the best channel for the highest hydraulic performance, while a zigzag MCHS was the best channel for the highest thermal performance when compared to conventional straight MCHSs. Converging–diverging microchannels were examined by H. Ghaedamini et al. [66], who found that converging–diverging microchannels improved the performance factor by 0.8, 1.0, and 1.2 at Reynolds numbers of 200, 400, and 600, respectively, compared to conventional microchannels. X. F. Peng et al. [67] experimentally studied rectangular microchannels with hydraulic diameters of 0.133–0.367 mm and a height to width ratio H/W = 0.333–1. Laminar heat transfer took place at Reynolds numbers 200–700, whereas fully turbulent convective heat transfer formed at Reynolds numbers 400–1500. It was observed that geometric parameters greatly influenced the heat transfer and fluid flow. Laminar convective heat transmission occurred when the aspect ratio was close to 0.75, but turbulent heat transfer functioned best between 0.5–0.75. A tree-shaped microchannel net was numerically investigated by Xiang-Qi Wang et al. [68]. The pressure drop and uneven temperature difference in a typical microchannel led to the conduct of this investigation. According to the findings, these channel networks outperform traditional parallel channel nets in a number of areas, such as having a more consistent and lower temperature distribution and greater stability when a channel segment becomes accidentally blocked.

A trapezoidal microchannel of different aspect ratios for fully developed, single-phase laminar flow without slip circumstances was quantitatively studied by John P. Mchale et al. [69]. A numerical setup was used for a silicon MCHS and an experimental setup was used for a polycarbonate–Al MCHS; 45 degrees and 54.7 degrees were selected as the side wall angles. The research presents the local and average Nusselt numbers as a function of dimensionless length and aspect ratio. It was investigated how the Prandtl number affects the thermal admission condition. The fully formed friction factors were calculated and linked as a function of the channel aspect ratio. The fluid moving over a curved channel faces centrifugal forces, which might increase the fluid velocity near the outside walls of the channel. The total transfer of heat capabilities of the heat sinks are improved by this increased velocity, leading to improved heat transfer between the fluid and the channel walls. The surface area for heat transfer is increased by surface roughness [63] and wavy microchannel heat sinks [64], which results in greater heat transmission. Explanation of computational domains is shown in Figure 26, Figure 27, Figure 28, Figure 29 and Figure 30.

Table 5. Additional data for channel curvature.

S. No	Author	Nature of Work	Geometry	Coolant Type	Substrate Material	Key Findings
1	Chu et al. [61]	Experimental	Triangular MCHS	Water	Silicon	Friction factor value of 21 Highest Nu value of 0.7
2	Daxiang Deng et al. [62]	Numerical and Experimental	MCHS with semi-closed omega-shaped configuration.	Deionized water	Oxygen-free pure copper	Pressure drop of 8 kPa Maximum thermal resistance value of 0.41 °C/W
3	Ping cheng et al. [63]	Experimental	Trapezoidal MCHS	Water	Silicon	Nu value of 4 at Re 700 Apparent friction factor value of 32
4	Deng, B et al. [64]	Numerical	Rectangular MCHS with various wavy amplitudes	Water	Aluminum	Maximum friction factor value of 0.55 Pressure drop value of 6.2 kPa
5	Gunnasegran et al. [65]	Numerical	Zigzag, curvy, and step microchannel heat sinks	Water	Aluminum	Heat transfer coefficient value of 9.64 Maximum pressure drop of 2800 Pa Poiseuille number value of 25
6	H. Ghaedamini et al. [66]	Numerical	Converging–diverging MCHS	Water	Silicon	Highest Nu value of 30 Performance factor of 1.18
7	X. F. Peng et al. [67]	Experimental	Square- shaped MCHS	Water	Silicon	Highest Nu value of 4.61 Heat transfer coefficient value of 30,500 W/m2
8	Xiang-Qi Wang et al. [68]	Numerical	Rectangular (tree-shaped) MCHS	Water	Silicon	Pressure drop value of 320 Pa Studied temperature distribution
9	John P. Mchale et al. [69]	Numerical	Trapezoidal and square-shaped MCHS	Water, ethylene glycol, air, and mercury	Silicon and polycarbonate aluminum	LOCal Nu value of 70 Average Nu value of 90

shows more information on the channel curvature described above.

2.2.7. Thermal Enhancement Techniques and Analysis for Smooth Microchannel Heat Sinks

In this section, the other work related to primary flow channels without using any obstructions will be discussed. Weilin Qu et al. [20] have examined the heat transfer and properties related to pressure drop of a single-phase microchannel heat sink. They noticed the laminar to turbulent flow transition within Reynold numbers ranging from 139 to 1672. The other parameter which has also been investigated was the Nusselt number. At the bottom wall of the channel, the Nusselt number was lower as compared to the channel side wall. A theoretical and experimental investigation of gas flow in the microchannel was performed by John C. Harley et al. [70], who concluded that the Knudsen number was less than 0.38 and the data were within 8% of the theoretical predictions of the friction constant. Through a triangular microchannel of silicon, the characteristics of heat transfer of water as a coolant were evaluated by I. Tiselj et al. [71].They concluded that the water temperature, as well as the temperature of the heated wall, do not change linearly along the channel. Todd M. Harms et al. [72] studied the single-phase forced convection of two configurations in a deep rectangular microchannel. They found that their microchannel system established for developing laminar flow outperforms the comparable single-channel system designed for turbulent flow, and lowering the channel width and enhancing the channel depth provides a better flow and good performance in terms of heat transfer. K.C. Toh. et al. [73] investigated fluid flow and heat transfer characteristics in a microchannel heat sink. They observed that the cold flow friction factor corresponds to the fully developed friction factor with parameters of the microchannel heat sink; by decreasing the Reynolds numbers, there is a reduction in the friction factor with heating. Gabriel Gamrat et al. [74] investigated the effects of entrance and conduction on liquid flow in the laminar region of a microchannel heat sink. They found that the reduction in channel spacing from 1 mm to 0.1 mm had no effect on heat transfer properties, and the effects of conduction were stronger at a low Reynolds number range.

The enhancement in heat transfer using the redevelopment of a thermal boundary layer was studied by J.L. Xu et al. [75]. They found that the increase in Nusselt numbers and a shorter effective flow length at the microchannel heat sink thermal entrance zone improved the heat transfer and caused significantly reduced pressure drops. The effect of geometrical parameters on the single-phase forced convective heat transfer in microchannel heat sinks was investigated by X. F. Peng et al. [76] They concluded that the properties of liquid convection are different from those of channels of a similar size, changing from a laminar flow regime at Re 300 to a completely turbulent flow regime at about Re 1000. Amy Rachel Betz et al. [77] found that segmented flow increases the Nusselt number by over 100%, reaching values of 10.1–22.9, and enhances heat transfer by 140%, compared to single-phase flow conditions. Azad Qazi Zade et al. [78] explored the effects of a growing gaseous slip flow in microchannels. Their findings indicated that the Nusselt number experienced changes of 15% in the fully developed region and 20% in the entry region.

Baoqing Deng et al. [64] studied an enhanced porous medium model for a microchannel heat sink and extended the porous medium to the substrate. They also derived an analytical solution for the hydrodynamically and thermal fully developed flow.. The effect of two-phase flow on the thermal performance of a heat sink was studied by Tom Saenen et al. [79], and they verified the numerical code using the method of manufactured solutions and revealed that the numerical order in space and time is consistent with the expected values from the theory, second-order, and first-order, respectively. Yang et al. [80] studied a single-phase hybrid microchannel heat sink with secondary oblique channels. They found that the pressure drop reduced by 11% and thermal resistance reduced by 24% compared to smooth MCHS. They concluded that their suggested heat sink exhibited a maximum chip temperature of 53 °C with a pressure drop of 3.77 kPa. Zhi-Qiang Yu et al. [81] conducted a comprehensive review on microchannel heat sinks for electronic cooling applications. They summarized the heat transfer applications, coolants, materials, and performance of different channels with single-phase and phase-change flow in MCHSs. Min Yang et al. [82] studied the Pareto frontier for the performance evaluation of MCHS configurations. They used pumping power and thermal resistance as performance evaluation criteria. They concluded that hybrid MCHSs exhibit lower pumping power and superior thermal performance.

Table 6. Additional data for optimization.

S. No	Authors	Geometry	Nature of Work	Substrate Material	Coolant Type	Key Findings
1	Wang X.Q. et al. [68]	Pin fins	Numerical	Copper	Air	Pressure drop value of 330 Pa
2	Mchale et al. [69]	Pin fins	Numerical	Aluminum	air	Friction factor value of 24 Nu value of 70
3	Harley, J.C. et al. [70]	Plate fin heat sink	Numerical	Aluminum alloy	air	Khudsen number less than 0.38
4	Ping et al. [58]	MCHS with dimple and pin fin	Numerical		air	Performance factor of 1.9 Highest Nu ratio of 2.4 Maximum friction factor ratio of 1.9
5	Tiseli I. et al. [71]	Plate fin heat sink	Analytical	Copper	Ambient air	Nu value of 5.1 Average temperature of 330 K
6	Harms et al. [72]	MCHS variable pin fin configuration	Numerical	Copper	water	Nu value of 110 Pressure drop value of 60 kPa
7	Toh, K.C. et al. [73]	Rectangular microchannel	Numerical	Silicon	Water	Thermal resistance value of 0.30 cm2.K/W Friction factor value of 150
8	Gamrat, G. et al. [74]	Rectangular microchannel	Numerical	Silicon	Water	Nu value of 30 Poiseuille number value of 40
9	Xu, J.L. et al. [75]	Rectangular MCHS	Numerical	Copper, aluminum, silicon	Water	Nu value of 16 Friction factor value of 0.02
10	Peng, X. et al. [76]	Rectangular MCHS	Numerical	Silicon	water	Nu value of 10 Heat transfer coefficient value of 7000 W/m2.K
11	Betz, A.R. et al. [77]	Rectangular MCHS	Numerical	Silicon	Water	Nu value of 14 Pressure drop of 38 kPa
12	Toh, K.C. et al. [73]	Rectangular MCHS	Numerical	Silicon	Air	Friction factor value of 70 Thermal resistance of 0.28 cm2/K/W
13	Zade, A.Q. et al. [78]	Rectangular MCHSs	Numerical	Copper	Water–air	Nu value of 33 Friction factor value of 70

shows the work of different researchers on the optimization of microchannel heat sinks, and

Table 7. Additional data for channels without obstructions.

S. No	Authors	Geometry	Nature of Work	Substrate Material	Coolant Type	Results Obtained
1	Yang, D et al. [18]	Rectangular Microchannel	Experimental and Numerical	Silicon	Water	Average Nu value of 30 Average temperature of 120 °C
2	Deng, B. et al. [64]	Rectangular Microchannel	Experimental	Silicon	Water	Thermal resistance value of 0.19 K/W
3	Saenen. T. et al. [79]	Triangular Microchannel	Experimental and Numerical	Silicon	Water	Average temperature of 128 °C
4	Sharma, C.S. et al. [83]	Rectangular Microchannel	Experimental	Silicon	Water	Pressure drop of 1600 Pa Fluid outlet temperature of 65 °C Thermal resistance of 0.285 °C cm2/W
5	Chen, C.W. et al. [84]	Rectangular Microchannel	Numerical and Experimental	Bronze Block	Water	Thermal resistance of 0.4 °C/W/cm2 Large flow power can develop low thermal resistance MCHSs
6	Japar, W.M.A.A. et al. [85]	Rectangular Microchannel	Experimental	Silicon	Water	Nu value of 10 Pressure drop value of 70 kPa Nu ratio of 2.3
7	Chai, L. et al. [86]	Rectangular Microchannel	Experimental	Silicon	Water	Pressure drop of 60 kPa Friction factor value of 60 Nu value of 15
8	Memon, S.A. et al. [87]	Rectangular Microchannel	Experimental	Aluminum	Water	Pressure drop of 2.8 Pa Minimum average temperature of 311 K
9	Yang, M. [88]	Rectangular Microchannel	Numerical	-	-	Pumping power of 2.1 W Thermal resistance of 1.3 K/W
10	Gao, W. et al. [89]	Rectangular Microchannel	Numerical	-	-	Pressure drop of 32,000 Pa Pumping power of 0.01 W Nu value of 22
11	Kuppusammy, N. et al. [90]	Array Mini Channel	Experimental and Numerical	Copper	Water	Nu value of 12.5 with pressure drop of 20 kPa
12	Bahiraei et al. [91]	Microchannel Heat Sink	Numerical	Silicon	Water	17% improvement in convective heat transfer coefficient
13	Min Yang et al. [82]	Microchannel Heat Sink with Manifold and Oblique Channels	Experimental	Copper	Water	Pressure drop of 4.2 kPa Friction factor value of 17

discusses research performed on microchannel heat sinks without any flow obstruction. Explanation of computational domains is shown in Figure 31, Figure 32, Figure 33, Figure 34, Figure 35 and Figure 36.

Saenen, T. et al. [79] examined an MCHS using a cooling system with a two-phase flow. Simple algorithms were used for the results prediction. The equations were solved using finite volume discretization, and the numerical codes were validated using experimental results. The correlation between the findings was good and accurately predicted the model of the system. Sharma et al. [83] analyzed the fluid flow heat transfer properties in a microchannel heat sink using hot water as a coolant for an even hotter electronic chip. There was laminar flow inside of the channel and turbulent flow at the inlet and outlet manifolds of the heat sink. For the analysis of local entropy generation, a Reynolds number of 2400 and a high Reynolds number of 11,200 were used. The results concluded that due to viscous dissipation, there is an increase in the Reynolds number and in the entropy generation. Hot water as a coolant reduces entropy generation. The effect of substrate thickness on heat transfer in MCHSs was numerically examined by Ali Kosar [14]. Different materials, including Polyimide, Silica Glass, Quartz, Steel, silicon, and copper of different substrate thicknesses, t = 100 µm–1000 µm, were analyzed. The Nusselt number was related as a function of the Biot number Bi and relative conductivity k. The Nusselt number for Silica Glass is 50% higher than for a Polyimide MCHS. Zhigang et al. [92] used a Reynolds number ranging from 101–1775 to analyze the effect of thermal property variations on heat transfer. Several methodologies, as well as experimental results, correlations, and data from the simplified theoretical solution, were used to compare the obtained local and average flow and heat transfer properties. The forced convection in microchannels was analyzed by Chien-Hsin Chen [19]. Numerical simulations were made on a fluid-saturated porous medium modeled microchannel. The results concluded that with an increasing aspect ratio and porosity, the overall Nusselt number increased. Chih-Wei Chen et al. [84] used the simulated annealing approach. Three design variables, including channel height, fin width, and channel width, were analyzed to find the optimal performance of the microchannel heat sink. Explanation of computational domains is shown in Figure 37, Figure 38, Figure 39, Figure 40 and Figure 41.

Table 8 shows other techniques used for microchannel heat sink modeling and analysis.

Table 8. Additional data for other techniques.

S. No	Author	Geometry	Nature of Work	Substrate Material	Coolant Type	Key Findings
1	Saenen, T. et al. [79]	Rectangular MCHS	Numerical	Silicon	Air	Maximum temperature of 126 °C
2	Wang S.l. et al. [93]	Rectangular MCHS with turbulent flow	Numerical	Copper	Hot Water	Studied Dean vortex Secondary branches weaken Dean vortex
3	Ali Kosar [14]	Rectangular MCHS	Numerical	Polyimide, Silica Glass, Quartz, Steel, Silicon, Copper	Water	Nu value of 4 Friction factor of 0.7
4	Tsung-Hsun Tsai et al. [94]	Rectangular MCHS with porous medium	Numerical	Silicon	Cu–Water CNT–Water	Pressure drop of 350 kPa Thermal resistance of 0.073 °C/W
5	Zhigang et al. [92]	Rectangular microchannels	Numerical	Copper, Silicon, Stainless Steel	Water	Local friction coefficient of 1 Heat transfer coefficient of 200 kW/m2.K Nu vlaue of 25
6	Chen, C.-H [19]	Rectangular MCHS	Numerical			Nu value of 280 Dimensionless velocity of 0.6
7	Chih-Wei Chen et al. [84]	Rectangular MCHS		Silicon	Water	Thermal resistance of 0.4 °C/W/cm2

Table 8. Additional data for other techniques.

S. No	Author	Geometry	Nature of Work	Substrate Material	Coolant Type	Key Findings
1	Saenen, T. et al. [79]	Rectangular MCHS	Numerical	Silicon	Air	Maximum temperature of 126 °C
2	Wang S.l. et al. [93]	Rectangular MCHS with turbulent flow	Numerical	Copper	Hot Water	Studied Dean vortex Secondary branches weaken Dean vortex
3	Ali Kosar [14]	Rectangular MCHS	Numerical	Polyimide, Silica Glass, Quartz, Steel, Silicon, Copper	Water	Nu value of 4 Friction factor of 0.7
4	Tsung-Hsun Tsai et al. [94]	Rectangular MCHS with porous medium	Numerical	Silicon	Cu–Water CNT–Water	Pressure drop of 350 kPa Thermal resistance of 0.073 °C/W
5	Zhigang et al. [92]	Rectangular microchannels	Numerical	Copper, Silicon, Stainless Steel	Water	Local friction coefficient of 1 Heat transfer coefficient of 200 kW/m2.K Nu vlaue of 25
6	Chen, C.-H [19]	Rectangular MCHS	Numerical			Nu value of 280 Dimensionless velocity of 0.6
7	Chih-Wei Chen et al. [84]	Rectangular MCHS		Silicon	Water	Thermal resistance of 0.4 °C/W/cm2

Figure 37. Rectangular MCHS with porous medium by Tsung-Hsun Tsai et al. [94]. Reproduced with permission from Tsung-Hsun Tsai et al. [94]; published by Elsevier, 2007.

Figure 37. Rectangular MCHS with porous medium by Tsung-Hsun Tsai et al. [94]. Reproduced with permission from Tsung-Hsun Tsai et al. [94]; published by Elsevier, 2007.

Figure 38. Rectangular microchannels of copper, silicon, and stainless steel by Zhigang et al. [92]. Reproduced with permission from Zhigang et al. [92]; published by Elsevier, 2007.

Figure 38. Rectangular microchannels of copper, silicon, and stainless steel by Zhigang et al. [92]. Reproduced with permission from Zhigang et al. [92]; published by Elsevier, 2007.

Figure 39. C rectangular MCHS using water–air coolant by Zade A.Q. et al. [78]. Reproduced with permission from Zade A.Q. et al. [78]; published by Elsevier, 2011.

Figure 39. C rectangular MCHS using water–air coolant by Zade A.Q. et al. [78]. Reproduced with permission from Zade A.Q. et al. [78]; published by Elsevier, 2011.

Figure 40. Rectangular microchannel by Chien-Hsin Chen et al. [19]. Reproduced with permission from Chien-Hsin Chen et al. [19]; published by Elsevier, 2007.

Figure 40. Rectangular microchannel by Chien-Hsin Chen et al. [19]. Reproduced with permission from Chien-Hsin Chen et al. [19]; published by Elsevier, 2007.

Figure 41. Rectangular MCHS by Guodong Xia et al. [95]. Reproduced with permission from Guodong Xia et al. [95]; published by Elsevier, 2011.

Figure 41. Rectangular MCHS by Guodong Xia et al. [95]. Reproduced with permission from Guodong Xia et al. [95]; published by Elsevier, 2011.

2.2.8. Induced Secondary Flow

This section covers secondary channels’ potential to improve the effectiveness of heat transfer in microchannel heat sinks. Secondary channels are smaller channels that are located inside the main microchannel. The secondary flow created by these channels increases the microchannel heat sink’s ability to transfer heat more effectively. Secondary channels decrease the thermal boundary layer’s thickness, improve the surface area available for heat transfer, and promote better flow mixing using secondary flow streams. For use in microchannel heat sinks, secondary channels can be arranged by a variety of methods. One common arrangement is straight parallel channels that are positioned transverse to the main microchannel. The secondary passages create vortices in the flow, which enhance mixing and heat transfer. The addition of curved secondary channels provides various arrangements, such as a serpentine or meandering design, to the main microchannel. These curved channels create secondary flows that increase heat transfer efficiency and decrease the thermal boundary layer thickness. For optimal heat transfer efficiency, the size, curvature, and spacing of the curved channels can be adjusted. Perforated plates are another secondary channel design for microchannel heat sinks. The heat sink is filled with thin plates with tiny holes, called perforated plates, which make up the primary microchannel. The perforated plates induce turbulence in the flow, improving heat transfer. For multilayer arrangements, MCHSs can also include secondary channels. Secondary channels are positioned in between layers of microchannels that are piled on top of one another. Secondary channels assist microchannel heat sinks to transfer more heat because they produce secondary flows that improve heat transfer. One of the challenges associated with the utilization of secondary channels is the fabrication of secondary channels with high precision and accuracy. To accomplish the intended heat transfer enhancement, the secondary channels must be positioned with a high degree of precision and accuracy. High precision and accuracy microchannel fabrication often costs a lot and calls for sophisticated fabrication methods like photolithography and MEMS technology.

Another challenge is pressure losses occurring due to the use of secondary channels. The performance of the flow rate and heat transfer may be impacted by the pressure drop in the channel, which may raise the overall pressure losses. To obtain the best overall performance, an optimized design with secondary channels should balance the pressure losses with the improvement of heat transfer. Also, secondary channels can lead to flow maldistribution and thermal nonuniformity in microchannel heat sinks. The secondary channels can induce complex flow patterns, resulting in maldistributed flow patterns, resulting in non-uniform heat transfer and temperature distribution. Using secondary channels in microchannel heat sinks presents several challenges in terms of fabrication, pressure loss, flow maldistribution, and thermal non-uniformity; therefore, the best design with secondary channels should consider these issues to achieve the best overall performance. Japar, W et al. [8] examined the effectiveness of the secondary channel numerically at a Reynolds number (Re) ranging from 100 to 450. A comparison analysis of related geometries was used to examine the efficiency of the suggested MCHS with rectangular ribs, triangular cavities, and microchannels with rectangular ribs and triangular cavities. The outcome shows that the suggested design, when compared to other designs, has an exceptional overall performance because of the collective effects of thermal boundary layer redevelopment and the mixing of flow in the main channel. Ghani et al. [96] conducted a numerical simulation at Reynolds numbers ranging from 100 to 800 to investigate the properties of heat transfer and fluid flow in MCHSs with cavities in sinusoidal-shaped and rectangular ribs. The optimized results showed that the suggested geometry, with cavities in sinusoidal-shaped and rectangular ribs with parameters of a relative cavity amplitude 0.15, relative rib width 0.3, and relative rib length 0.5, has the best overall performance of 1.85.

The heat transfer of microchannel heat sinks with periodic expansion–constriction cross-sections was investigated both experimentally and numerically by Lei chai et al. [86]. The results showed that in comparison with the straight rectangular MCHS, the pressure drop in the suggested new heat sink is lower when Re < 300, but increases rapidly and was higher at 300 < Re < 750; heat transfer is significant due to the greater mixing of the cooling fluid. Memon et al. [87] studied computationally modeled secondary flow channels in different orientations, rectangular and trapezoidal. It is evident from the results that the trapezoidal microchannel heat sink had a higher average Nusselt number and enhanced heat transfer coefficient of 64% as compared to the conventional parallel orientation without secondary flow. M. Yang et al. [88] optimized microchannel heat sinks with secondary oblique channels, achieving an 18.83% reduction in thermal resistance compared to traditional manifold MCHSs while maintaining similar pressure levels. W. Gao et al. [89] suggested a multijet impingement heat sink with trapezoidal fins and secondary channels via using numerical simulations. The results predicted that the suggested heat sink performs well, with good heat transfer enhancement. N.R. Kuppusamy et al. [90] employed secondary flow in MCHSs in an alternating orientation by introducing a slanted passage in the channel wall. Due to disruption in the hydrodynamic boundary layer and redevelopment at the leading edge of the following wall, the results showed that the overall performance of the suggested MCHS increased by 1.46 times and the thermal resistance also reduced. The thermohydraulic attributes of a nanofluid containing graphene/silver nanoparticles in a MCHS with ribs and secondary channels were evaluated by Bahirae et al. [91]. The use of auxiliary channels, ribs, and nanofluid greatly enhanced the performance. The average convective heat transfer coefficient increased by 17% at Re = 100 when the concentration rose from 0 to 0.1% with an increase in Reynolds number.

Wang et al. [93] showed that by incorporating transverse gaps into the ribs, wavy configurations with secondary branches were created for the investigated microchannel heat sinks, which had symmetric and parallel wavy microchannels at Reynolds numbers ranging from 50 to 700. The new hybrid design’s performance was evaluated against the original wavy configurations. The findings showed that the new hybrid MCHS performed better overall with a small increase in pressure loss. Ma D. et al. [95] developed wavy sinusoidal microchannels with secondary channels. The testing findings demonstrated that the proposed microchannels could provide superior heat elimination performance to the standard rectangular microchannels because of the existence of the secondary channel. A new microchannel heat sink (MCHS) design with cavities in a trapezoidal shape, secondary channels, and ribs was proposed by SA Razali et al. [97]. The heat transfer characteristics and fluid flow were numerically analyzed for Reynold numbers ranging from 100 to 500. Four different microchannel heat sink geometries were placed under study. They found that the secondary MCHS with ribs in a rectangular shape is the best proposed design compared to the other four geometries, with a maximum performance factor of 1.78. Farzaneh et al. [98] investigated the effect of branching to develop a system configuration that had minimum resistance to heat flow. According to the examined data, raising the number of branches lowers both the temperature and pressure drop. According to the experiments, when compared to a case without a branch, ones with a greater number of branches have greater heat transfer characteristics. Huang et al. [98] numerically studied three types of MCHS for heat transfer characteristics, which were the Parallel-Slot Rectangular Microchannel Heat Sink, Stagger-Slot Rectangular Microchannel Heat Sink, and Trapezoidal Stagger-Slot Microchannel Heat Sink. The findings demonstrated that the Trapezoidal Stagger-Slot MCHS was the best in terms of total heat transfer capabilities and thermal performance, with a non-changing pumping power since it has the lowest drop in pressure and the best heat transfer qualities. The oblique fin microchannel heat sink was experimentally investigated by Poh Seng Lee et al. [99]. Due to the oblique fins, there is also the introduction of a secondary path for the flow. As a result, the average Nusselt number increased from 8.6 to 15.8, with a reduction in the total thermal resistance.

Yan Fan et al. [100] numerically studied the heat transfer in laminar flow through novel cylindrical discrete oblique fin heat sinks. Their results revealed that the average Nusselt number increased by 73.5%, from 14.6 to 25.4, and the average convective thermal resistance decreased by 61%, from 0.075 °C/W to 0.029 °C/W, compared to smooth microchannel heat sinks (MCHSs). Secondary channels reduce the thickness of the thermal boundary layers, increasing the surface area available for heat transfer, which promotes better flow mixing via secondary flow streams. In these microchannels with secondary flow, the pressure drop is low at lower Reynolds numbers but increases with increasing Reynolds numbers [41]. Secondary flow decreases thermal resistance and increases channel transport efficiency [88]. The disruption in the boundary layer caused by secondary flow improves the fluid mixing, thus increasing the Nusselt number. Explanation of computational domains is shown in Figure 42, Figure 43, Figure 44, Figure 45, Figure 46, Figure 47, Figure 48, Figure 49, Figure 50 and Figure 51.

Table 9. Additional data for secondary flow.

S. No	Author	Geometry	Nature of Work	Substrate Material	Coolant Type	Key Findings
1	Japer. W et al. [8]	Rectangular MCHS	Numerical	Copper	Water	Review of literature
2	Lei Chai et al. [41]	Rectangular MCHS with triangular cavities	Experimental and numerical	Silicon	Water	Thermal enhancement factor of 1.6 Nu ratio of 1.8
3	Ghani et al. [96]	Rectangular MCHS with ribs and cavities	Numerical	Copper	Water	Friction factor ratio of 4 Nu ratio of 2.2 Performance factor of 1.7
4	Bahirae et al. [91]	MCHS with ribs and secondary channels	Numerical	Copper	Graphene/silver nanoparticles	Secondary channel increases area, thus reducing pressure drop
5	Shou Lin Wang et al. [93]	Secondary flow branched design	Numerical	Silicon	Water	Nu value of 18 Performance factor of 2.3 Thermal resistance value of 1.2
6	GD Xia et al. [95]	Sinusoidal secondary flow MCHS	Experimental	Silicon	Water	Heat transfer coefficient value of 45 Pressure drop value of 240 kPa
7	SA Razali et al. [101]	MCHS design with trapezoidal cavities, ribs, and secondary channels	Numerical	Copper	Water	Performance factor of 1.8 Maximum Nu ratio of 2.3 Highest friction factor ratio of 5.7
8	Farzaneh et al. [97]	Square-shaped MCHS with one and two initial loops.	Experimental	Silicon	Water	Pressure drop reduced by 25%
9	Huang et al. [98]	Rectangular Parallel-Slot MCHS	Numerical	Copper	Water	Nu value of 21 Pressure drop ratio of 2000 Performance factor value of 1.95
10	Poh Seng Lee et al. [99]	Secondary flow MCHS with oblique fins	Experimental	Copper	Water	Average Nu value of 16 Thermal resistance value of 0.097 °C/W Pressure drop value of 3000 Pa
11	Yan Fan et al. [100]	Novel cylindrical discrete oblique fin MCHS with secondary flow	Numerical	Copper	Water	Pressure drop of 90 Pa Heat transfer coefficient value of 3000 W/m2.K

summarizes the use of secondary flows in microchannel heat sink improvement.

2.2.9. Nanofluids

The poor thermal conductivity of common fluids like water, ethylene glycol, or oil results in a low heat transfer efficiency in MCHSs. Nanoscale particles suspended in a basic fluid, like water or oil, make up nanofluids. More heat dissipation from MCHSs may arise from these nanoparticles’ increased thermal conductivity of the base fluid, but whereas nanofluids can significantly increase heat transfer in microchannel heat sinks, their use is also limited by several limitations. Nanoparticle agglomeration and sedimentation will reduce the thermal conductivity of nanofluids and clog microchannels as they settle over time. This may affect the long-term stability of the nanofluid and reduce its heat transfer efficiency. CuO–water nanofluids were found to lose up to 30% of their thermal conductivity after 30 days as a result of agglomeration in a study by Lyu et al. [102]. The higher density and viscosity of nanofluids compared to their base fluids causes a greater pressure drop in microchannels, which may increase pumping power requirements and limit the maximum flow rate.

S. Abubakar [103] investigated the impact of temperature on a microchannel heat sink with water and Fe₃O₄-H₂O₄ as fluids. To replicate conjugate heat transfer in a three-dimensional rectangular microchannel heat sink, a continuous heat flow of 9,000,000 W/m² was applied to the top of the silicon heat sink. The presence of nanoparticles lowers the temperature of the surfaces as the particle volume % of Fe₃O₄-H₂O₄ grows due to its higher dynamic viscosity and lower heat capacity compared to pure water. When using Fe₃O₄-H₂O₄ as a working fluid with a volume fraction of 0.4 in comparison with pure water, it shows a 0.04% temperature reduction, and there is direct relation between nanoparticle concentration and temperature. Three distinct nanofluids—ethylene glycol-based CuO, water-based CuO, and Al₂O₃—were used in a study by A. Sivakumr et al. [104] to examine the thermal performance of a serpentine-shaped microchannel heat sink in the Reynolds number range of 100 to 1300. The results showed that, in comparison to their base fluids, the heat transfer coefficients of the nanofluids CuO/ethylene glycol and Al₂O₃/water are higher, and that the heat transfer coefficient of the nanofluid CuO/ethylene glycol is higher than the heat transfer coefficient of the nanofluids CuO/water and Al₂O₃/water. L. Snoussi et al. [105] numerically studied laminar nanofluid flow in 3D rectangular copper MCHSs and constant heat flux. It was discovered that the heat transfer coefficient increased by 14% at a 2% Al₂O₃/water nanofluid concentration. Additionally, 4% more heat is transferred when utilizing Cu/water than Al₂O₃/water, which may be a result of Cu/water’s better thermal conductivity. In the study by Sarafraz et al. [106], the heat transfer coefficient (HTC) values for CNT/water nanofluid were reported as follows: At a Reynolds number of 200, the HTC was 3870 W/m²·°C, and at 1100, it increased to 6100 W/m²·°C. For a mass concentration of 0.1% CNT, at a heat flux of 90 kW/m², the HTC was 3580 W/m²·°C, while for a 0.15% concentration, it was 4645 W/m²·°C. In an experimental study, MR Thansekhar et al. [107] examined the effects of nanofluids with volume concentrations of 0.1% and 0.25% in Al₂O₃/water and SiO₂/water, as well as deionized water. The greatest heat transfer improvement achieved by the nanofluid Al₂O₃/water at a 0.25% volume concentration was 36.63% when compared to water. Al₂O₃/water outperformed SiO₂/water nanofluid in heat transfer because of the higher thermal conductivity of Al₂O₃ nanoparticles.

In double-layered MCHSs, A.A.A. Arani et al. [108] examined the heat transfer properties of nanofluid water/single-wall carbon nanotubes at Reynolds numbers 500, 1000, and 2000. The effect of the proportion of nanoparticles in water-based nanofluid in a Newtonian suspension was explored for values of 0.04, 0.08, and 0. 0. Their findings revealed that the ratio of the maximal to the lowest temperature difference for the microchannel’s bottom wall, as well as the thermal resistance ratio, decreased with an increasing nanoparticle volume percentage and decreasing truncated length. Arabpour A. et al. [109] reported that an MCHS cooled with SiO₂ nanofluids had a thermal performance enhancement of 3580 to 4645 W/m² °C compared to water. The CCZ-HS showed a thermal performance improvement of 3580 to 4645 W/m² °C over the CZ-HS. R. Vinoth et al. [7] examined the impact of a channel cross-section on an oblique finned MCHS’s heat transfer efficiency experimentally. The nanofluid used was Al₂O₃/water with a volume fraction of 0.25%. The three channel cross-sections of the oblique finned microchannels were square, semicircular, and trapezoidal. When nanoparticles are introduced to the base fluid, the rate of heat transmission increases by 4.6% when compared to water. It has been discovered that trapezoidal cross-section profiles are more effective than square and semicircular cross-section profiles in heat transmission and flow characteristics.

Abdollahi et al. [110] studied the heat transfer properties of a nanofluid flow by conducting a three-dimensional numerical simulation of an interrupted MCHS. With an Al₂O₃/water coolant nanofluid, an interrupted MCHS with elliptical and diamond ribs was taken into consideration. It was shown that the Al₂O₃/water nanofluid may increase the Nusselt number by more than 30% at a volume fraction of 5% nanoparticles. Based on the comparison with an unbroken, pure water-cooled MCHS, the findings showed that an elliptical rib MCHS with nanofluid can enhance microchannel performance. Using Al₂O₃ nanofluid combined with various base fluids, Altayyeb Alfaryajat et al. [111] quantitatively examined the heat transfer properties in a 3D rhombus MCHS. The Al₂O₃ content in these fluids was 4%, and the nanoparticle diameter was 25 nm. Al₂O₃/water was shown to have the lowest temperature, the best heat transfer properties, and the least amount of thermal resistance when compared to the other base fluids. MM Sarafraz et al. [112] evaluated the thermal performance of MCHS using Ag/water as a nanofluid by experimental means. The silver/water nanofluid was shown to have a greater heat transfer coefficient than the base fluid. The results showed that the heat transfer coefficient rose by 47% at a silver (Ag) to water ratio of 0.1. A novel multi-nozzle trapezoidal MCHS (MNT-MCHS) using Al₂O₃ and TiO₂ as coolants was numerically investigated by Ngoctan Tran. et al. [113]. The optimized findings indicated that higher thermal conductivities and bigger volume fractions of nanoparticles in nanofluids will result in a better thermal performance than lower thermal conductivities and smaller volume fractions of nanoparticles in nanofluids. The thermal performance of single-walled carbon nanotubes (SWCNTs) and multi-walled carbon nanotubes (MWCNTs) distributed in the two base fluids of water and kerosene in a fractal microchannel was quantitatively examined by Zongjie Lyu et al. [102] at Reynolds numbers of 1500 to 3000. According to the study, at the same Re number and nanoparticle volume percent, the water-based nanofluid’s performance assessment criteria was four times higher than the kerosene-based nanofluid’s. As a result, it was advised to cool SWCNT nanoparticles in fractal silicon microchannels using water as the working fluid.

A.A Razali. et al. [114] examined the impact of heat transmission using both computational and experimental methods to Al₂O₃ nanofluids after their being used as a working fluid in MCHSs. When Al₂O₃ nanofluids were used instead of water, it had been found that the rate of heat transfer was greater. However, the Al₂O₃ particle structure tends to be more intact, and the crystallite size increased after increasing the temperature in the microchannel, according to the X-ray diffraction method (XRD). In the study by AR Chabi et al. [115], they used CuO/water nanofluids with concentrations of 0.1% and 0.2% by volume to examine forced convective heat transfer in a microchannel heat sink (MCHS). The experimental results revealed that at a Reynolds number of 1150, the average heat transfer coefficient for a 0.2 vol% CuO nanofluid was increased by approximately 7200 W/m² °C compared to deionized water. The top limits of the particle volume fraction for the TiO₂/water nanofluid heat transfer performance in MCHSs were experimentally determined by Manay et al. [95]. An analysis was performed on the impact of Reynolds numbers ranging from 100 to 750 and the particle volume percentage at a constant microchannel height of 200 mm on the parameters of heat transfer and pressure drop. Heat transfer was seen to be improved by the water/TiO₂ nanofluid up to 2.0 vol% but declined after that point. Furthermore, the thermal resistance was computed, and it was shown that adding nanoparticles to the base fluid with an average diameter of less than 25 nm might lower the thermal resistance.

A numerical analysis of the performance of a nanofluid-cooled microchannel cardiac sink was conducted by Tsung-Hsun Tsai et al. [94]. Through the use of a porous media model, the velocity and temperature distributions within the MCHS were determined. Heat transmission between the working fluid and the MCHS increased significantly when this fluid was used. By increasing the aspect ratio and porosity from optimum values, nanofluid did not decrease thermal resistance. Nanoparticles improve the thermal conductivity of conventional fluids like water or oils, in addition to their convective heat transfer coefficient, leading to increased heat dissipation from microchannels. However, these nanofluids have some disadvantages, including the fact that nanoparticles settle in the channel over time, creating an additional layer and increasing thermal resistance. The more dense nanofluids cause a higher pressure drop and require more pumping power [102]. However, nanoparticles significantly enhance fluid heat transfer abilities, which is the main goal of microchannel heat sinks. Explanation of computational domains is shown in Figure 52, Figure 53, Figure 54, Figure 55 and Figure 56.

Table 10. Additional data for nanofluids.

S. No	Author	Geometry	Nature of Work	Substrate Material	Coolant Type	Results Obtained
1	S. Abubakar et al. [103]	Straight rectangular MCHS	numerical	silicon	pure water, Fe3O4-H2O4	Maximum average temperature of 316 K
2	A. Sivakumar et al. [104]	Serpentine-shaped MCHS	Experimental and numerical	copper	ethylene glycol-based CuO, water-based CuO, and Al2O3	Heat transfer coefficient value of 42 kW/m2.K Pressure drop of 0.30 kPa
3	L.snoussi et al. [105]	Rectangular cross-section 3D MCHS	numerical	copper	Al2O3/water, Cu/water	Heat transfer coefficient value of 16 kW/m2.K Friction factor value of 0.07
4	Sarafraz et al. [106]	Rectangular MCHS	experimental	copper	multi-walled carbon nanotubes	Heat transfer coefficient value of 7000 W/m2.K Thermal resistance value of 0.11 W/°C
5	Thansekhar et al. [107]	Rectangular MCHS	experimental	copper	Al2O3/water, SiO2/water	Heat transfer coefficient value of 180 W/m2.K Thermal resistance value of 1.1 K/W Pressure drop value of 0.035 bar
6	A.A.A Arani et al. [108]	Truncated double-layer MCHS	numerical	silicon	water/single-wall carbon nanotubes (SWCNT)	Maximum pressure drop value of 500 kPa Highest friction factor value of 0.21 Average Nu value of 3
7	Arabpour, A. et al. [109]	Zigzag flow channels	experimental	copper	SiO2/DI water	Nu value of 14 Enhancement ratio of 1.2 Maximum pressure drop value of 110 kPa
8	R. Vinoth et al. [7]	Oblique finned MCHS with square, semicircular, and trapezoidal cross-sections	experimental	copper	Al2O3/water	Nu value of 19 Pressure drop of 20 kPa Friction factor of 3.8
9	Abollahi et al. [110]	Interrupted MCHS with ellipse and diamond ribs	numerical	copper	Al2O3/water	Highest Nu value of 13 Maximum friction factor value of 0.34 Highest performance factor of 1.3
10	Altayyeb Alfaryajat et al. [111]	3D rhombus MCHS	numerical	-	Al2O3 mixed with four different base fluids pure water engine oil glycerin ethylene glycol	Heat transfer coefficient value of 27 kW/m2.K Highest friction factor value of 0.040
11	MM Sarafraz et al. [112]	Straight rectangular MCHS	experimental	copper	biologically produced silver/water	Highest heat transfer coefficient value of 7000 W/m2.K Thermal resistance value of 0.11 W/°C Maximum pressure drop of 55 kPA
12	Nagoctan Tran et al. [113]	Multi-nozzle trapezoidal MCHS	numerical	copper	Al2O3 and TiO2	Thermal resistance value of 0.35 °C/W Highest pressure drop of 55 kPa
13	Zongjie Lyu et al. [104]	Single-layer fractal MCHS	numerical	silicon	water and kerosene	Highest Nu value of 45 Pumping power value of 0.15 W Maximum performance factor of 2.3
14	A.A Razali et al. [114]	Rectangular MCHS	experimental and numerical	copper	Al2O3	Nu value of 6
15	AR Chabi et al. [116]	Rectangular MCHS	experimental	copper	CuO/water	Heat transfer coefficient value of 20,000 W/m2. K Average Nu value of 13
16	Many et al. [115]	Rectangular MCHS	experimental	copper	TiO2/water	Nu value of 12 Thermal resistance value of 0.8 °C/W

summarizes the data for the use of nanofluids in microchannel heat sinks.

3. Conclusions

After conducting a thorough analysis of the available literature on MCHSs, we conclude that shape and internal geometry of the cooling system are essential in enhancing their thermal performance. Different aspects such as temperature, flow uniformity, and heat transfer rate are influenced by the system’s geometry, as factors such as the heat transfer coefficient, Nusselt number, and pressure drop determine the cooling system’s overall effectiveness.

• The outcomes of this review paper suggest that flow disruption techniques, optimization, secondary flows, and the use of nanofluids are all promising approaches for optimizing the thermal performance of MCHSs.
• Flow disruption techniques such as ribs, grooves, cavities, dimples, and fins play an important role in enhancing the thermal performance of MCHSs by disturbing the flow of the coolant and promoting a better heat transfer between the coolant and the heat sink surface.
• Optimization is a crucial step in adjusting various design parameters of the heat sink to obtain an optimal performance, while using a secondary flow promotes better mixing of the fluid and enhances convective heat transfer.
• The use of nanofluids combined with flow disruption methods is also promising for enhancing MCHSs’ thermal performance by boosting the convective heat transfer coefficient and promoting improved coolant mixing.

Overall, this critical review provides valuable insights into the impact of channel geometrical parameters on the efficiency of MCHSs and highlights potential areas for future research. By optimizing their design and exploring innovative techniques, it is possible to improve the thermal performance of MCHSs and unlock new possibilities for various applications.

4. Future Recommendations

Heat management has always been a major challenge because of how quickly electronic devices have evolved. Therefore, future studies should focus on the following areas:

• The optimization of MCHSs of different designs should be carried out to obtain their ideal geometries and enhance their heat transfer efficiency. Future research should look at the use of multi-objective optimization approaches to balance various design features for microchannel heat sinks. We can achieve optimal solutions that go beyond conventional single-objective optimization techniques by taking into account several performance indicators at once, including the pressure drop, heat transfer coefficient, and uniformity of temperature distributions.
• Future research should explore the concept or the introduction of secondary flows in various designs of microchannel heat sinks, such as asymmetric, tapered, or curved channels. Researchers can find the best designs that best use secondary flows and enhance the overall performance by carefully investigating the effects of various geometries on heat transfer characteristics.
• Future research should investigate the manufacturing of MCHSs with an unconventional fin geometry using advanced manufacturing techniques like 3D printing and selective laser sintering. Complex and complicated fin structures that were previously challenging or impossible to construct using conventional techniques can now be made because of this technology. Researchers can investigate unconventional fin shapes, such as elliptical, triangular, and conical structures or customized arrangements, by using advanced manufacturing, to improve heat transfer efficiency.
• Future studies should focus on developing nanoparticles with tailored properties specifically for microchannel heat sink applications. They should explore the use of more nanomaterials besides traditional metallic or oxides nanoparticles. Researchers should examine the possibility of adding carbon-based nanomaterials to nanofluids, such as carbon nanotubes or graphene, evaluating their effect on thermal stability and heat transfer performance and their compatibility with microchannel heat sink materials.
• Researchers should study the advantages of including more than one layer in the microchannel structure, and investigate patterns with various channel widths, heights, or shapes across multiple layers. This can improve heat transfer uniformity, flow mixing, and efficient fluid distribution. They should explore the effects of multiple-layer arrangements on the pressure drop and overall thermal performance.