**5. Internal heat transfer of cooled turbine airofoils**

The need for the internal cooling of gas turbine blades is primarily defined by the magnitude of the incident heat load on the airofoils, which range from 0.5 to 5 MW/m<sup>2</sup> , and the requirements of the component durability for long operating hours against thermomechanical fatigue (TMF), low cyclic fatigue (LCF), creep, oxidation, and high cyclic fatigue (HCF). While the external airofoil profile defines the airofoil aerodynamic performance, the internal cooling geometry is defined by the amount of coolant required to maintain the airofoil at a certain material temperature and the temperature gradients across critical wall sections of the airofoil. **Figure 21** shows some typical examples of turbine vane and blade cooling designs. The internal heat transfer technologies used in these vanes and blades include impingement cooling, turbulators or ribs, pin or pedestal banks, dimples, shaped internal passages, and combinations of the above cooling features.

**Figure 21.** Typical blade cooling designs, (a) nozzle guide vane [22], (b) turbine vane [23], and (c) turbine blade [19].

#### **5.1. Convective cooling with jet impingement**

hot gas temperatures [4–6, 11, 16, 17, 20, 21]. Over the last decade, there has been a significant focus on airofoil, platform, and blade tip film cooling with more recent focus on advanced shapes of film cooling holes, such as three-dimensional shaped holes and trench holes. In a recent study [13], the multirow film cooling characteristic on a high lift vane and blade were demonstrated. **Figure 20** shows that the use of three-dimensional advanced fan shaped holes can provide high airofoil average film cooling effectiveness and the use of only one or two row of shaped holes located upstream of the suction side shoulder can provide high film cool-

**Figure 20.** Multi row film cooling characteristics on a gas turbine (a) vane and (b) blade [13].

ing effectiveness until the trailing edge.

126 Heat Exchangers– Design, Experiment and Simulation

**Figure 19.** Effect of film cooling effectiveness on hot gas temperatures.

Impingement cooling is widely used for the internal cooling of gas turbine components, particularly static airofoils (vanes), heatshields (casing segments), combustor liners, and fuel nozzles. The impinging jets are generally formed through cylindrical holes in a thin wall insert, which is positioned adjacent to the airofoil inner wall that is required to be cooled. They are normally directed as a single row of jets or as multiple rows of jets, and are generally injected normal to the target surface.

For the midchord areas of airofoils, impingement cooling is designed with multiple rows of jets and is directed on the pressure and suction sides of the airofoil. The efficiency of the impingement cooling is defined by several parameters such as the standoff distance of the impingement jet relative to the target surface, the axial and radial pitch of the neighbouring impingement hole, the arrangement pattern (e.g., in-line, staggered, or other combinations), and the amount of cross-flow from the upstream impingement jets. An overview of recent research into impingement cooling is given in Ref. [24] with impingement cooling applications detailed in Refs. [4–6, 27]. For relatively flat surfaces, **Figure 22** shows the impingement hat transfer for a flat surface with multiple impingement holes, which is based on the correlations in Ref. [25]. This figure highlights, that although jet impingement cooling is highly effective, the design of the impingement system requires careful consideration of several influencing parameters, such as the standoff distance from the target surface, the axial and lateral pitch of the impingement holes and the amount of crossflow from upstream jets.

**Figure 22.** Impingement heat transfer with multiple rows on flat target surfaces.

Airofoil curved leading edges are normally subjected to very high heat loads, and at these locations, internal impingement cooling in combination with turbulators and film cooling is quite common. **Figure 23** shows the dependency of the standoff distance and the leading edge curvature on the coolant Nusselt numbers at varying Reynolds numbers, based on the correlation in Ref. [26]. Highest stagnation heat transfer can be achieved if the jets are arranged very close to the target surface and the highest average heat transfer are achieved for airofoils with small internal leading edge diameters. At the internal leading edges, there are also several other additional factors that influence the airofoil heat transfer, such as showerhead film cooling, turbulators, surface roughness, and the amount of impingement crossflows [4–6].

More recently, there have been several studies on the use of narrow channel impingement passages and inclined impingement jets in variable shaped passages. Such design configurations can provide higher internal heat transfers and have been mainly driven by the introduction of near wall cooling features in gas turbine blades. Such configurations can be manufactured with 3D printing technologies such as selective laser melting (SLM) and direct laser melting (DLM), which allows greater manufacturing flexibility with geometrically complex cooling passages.

**Figure 23.** Impingement heat transfer on curved leading edge surface.

overview of recent research into impingement cooling is given in Ref. [24] with impingement cooling applications detailed in Refs. [4–6, 27]. For relatively flat surfaces, **Figure 22** shows the impingement hat transfer for a flat surface with multiple impingement holes, which is based on the correlations in Ref. [25]. This figure highlights, that although jet impingement cooling is highly effective, the design of the impingement system requires careful consideration of several influencing parameters, such as the standoff distance from the target surface, the axial and lateral pitch of the impingement holes and the amount of crossflow from upstream jets.

128 Heat Exchangers– Design, Experiment and Simulation

Airofoil curved leading edges are normally subjected to very high heat loads, and at these locations, internal impingement cooling in combination with turbulators and film cooling is quite common. **Figure 23** shows the dependency of the standoff distance and the leading edge curvature on the coolant Nusselt numbers at varying Reynolds numbers, based on the correlation in Ref. [26]. Highest stagnation heat transfer can be achieved if the jets are arranged very close to the target surface and the highest average heat transfer are achieved for airofoils with small internal leading edge diameters. At the internal leading edges, there are also several other additional factors that influence the airofoil heat transfer, such as showerhead film cooling, turbulators, surface roughness, and the amount of impingement crossflows [4–6].

**Figure 22.** Impingement heat transfer with multiple rows on flat target surfaces.

More recently, there have been several studies on the use of narrow channel impingement passages and inclined impingement jets in variable shaped passages. Such design configurations can provide higher internal heat transfers and have been mainly driven by the introduction of near wall cooling features in gas turbine blades. Such configurations can be manufactured with 3D printing technologies such as selective laser melting (SLM) and direct laser melting (DLM), which allows greater manufacturing flexibility with geometrically complex cooling passages.

For narrow channel impingement, it was recently highlighted in Ref. [28] that in addition to the high heat transfer from the target surface, the heat transfer from the impingement cavity side walls can also be significant. **Figure 24** shows that for a narrow channel with in-line impingement holes, the heat transfer from the side walls can be up to 50% of that from the target plate.

**Figure 24.** Impingement heat transfer in narrow channel passages [28].

The use of inclined impingement jets on shaped turbulators in irregular shaped passages can also result in very high heat transfer. In a recent study [29, 30], and as shown in **Figure 25**, it was highlighted that directed inclined impingement can result in relatively high heat transfers from the target walls and additionally produces intense convective fluid mixing within the passage. In a turbine blade, such a combination can result in greater total heat removal by the coolant from the hot airofoil walls. The use of directed impinging cooling jets in the leading edge passages was demonstrated in Ref. [31]. **Figure 26** shows that directing a double impingement jets on a curve leading edge with showerhead cooling results in high heat transfer coefficients at the pressure and suction surfaces, and additionally generates significant turbulent mixing within the leading edge passage.

Based on the many different design variations of impingement cooling, it is expected that impingement cooling systems will continue to play a significant role in gas turbine heat transfer technology.

**Figure 25.** Impingement heat transfer in irregular passages [29, 30].

**Figure 26.** Impingement heat transfer in leading edge channels [31].

#### **5.2. Convective cooling with turbulators**

The use of inclined impingement jets on shaped turbulators in irregular shaped passages can also result in very high heat transfer. In a recent study [29, 30], and as shown in **Figure 25**, it was highlighted that directed inclined impingement can result in relatively high heat transfers from the target walls and additionally produces intense convective fluid mixing within the passage. In a turbine blade, such a combination can result in greater total heat removal by the coolant from the hot airofoil walls. The use of directed impinging cooling jets in the leading edge passages was demonstrated in Ref. [31]. **Figure 26** shows that directing a double impingement jets on a curve leading edge with showerhead cooling results in high heat transfer coefficients at the pressure and suction surfaces, and additionally generates significant

Based on the many different design variations of impingement cooling, it is expected that impingement cooling systems will continue to play a significant role in gas turbine heat trans-

turbulent mixing within the leading edge passage.

130 Heat Exchangers– Design, Experiment and Simulation

**Figure 25.** Impingement heat transfer in irregular passages [29, 30].

fer technology.

The use of ribs or turbulators for cooling gas turbine blades is a major heat transfer technology and has been employed largely in rotating blades with radial passages as shown in **Figure 27**. The passages are typically arranged as multiple radial passages and commonly referred to as serpentine passages or multipass systems, and the turbulators are generally designed on the pressure and suction surfaces of the passages. The key function of the turbulators is to create regions of flow separation downstream of the turbulators, which promotes intense regions of turbulence, secondary flows, and rapid mixing between the air warmed by the heated walls and the core coolant flow. The nature of the flow structure and the amount of heat transferred in the passages with turbulators is dependent significantly on the turbulator shape, configuration pitch, height, angle of orientation, flow Reynolds number, rotation, passage shape, and its aspect ratio. There has been a significant amount of research conducted on the application of turbulators in gas turbine blades, including the effects of rotation, shapes, sizes, orientation, entrance length effects, position of film cooling holes, presence of bends and other enhancement devices, and operating parameters [4–6, 19, 33, 52].

To assess the relative impact of different turbulators on the heat transfer and frictional characteristics, **Figure 27** shows the ribbed wall heat transfer and passage frictional enhancement of turbulators in various aspect ratio passages, based on the correlation in Ref. [32]. **Figure 27** highlights that the turbulator angle and the shape of the passage have a major effect on both the passage heat transfer and pressure loss. Although smaller aspect ratio ducts (W/H = 0.25) generally give higher heat transfer enhancement on the ribbed walls and lower pressure losses, the passage average heat transfer values can be lower due to the larger perimeter of the nonribbed walls, which have much lower heat transfer enhancement. Although the results shown in **Figure 27**

**Figure 27.** Passage heat transfer and frictional losses due to turbulator design and passage shape.

are for low Reynolds number, and with idealized geometry, care needs to be taken when implementing such results in real turbine blades with cast geometries, where the dimensions, shape, and position of the turbulators can be different from the predicted idealized geometries.

The impact of rotation on the heat transfer from gas turbine blades can be significant and is dependent on several additional parameters such as the rotational and buoyancy numbers [34, 35]. **Figure 28** shows a schematic overview of the impact of rotation on the flow field in a twopass rotating passage of a gas turbine blade. Under rotating conditions, Coriolis and buoyancy effects can significantly alter the temperature and velocity profiles within the passages.

In a study from Ref. [35], **Figure 28** shows that with the coolant flowing radially upwards, the heat transfer with increasing rotation numbers, increases on the pressure side and reduces on the suction surfaces. Similarly, when the coolant flows radially inwards, the heat transfer increases on the suction side and reduces on the pressure side, especially for smooth passages. However, for passages with turbulators, the turbulators tend to dampen the effect of rotation and the heat transfer enhancement on the pressure and suction sides. This overall trend of rotation with different type of passages and turbulators has been observed by several studies [33–35], and these effects play an important role in the design of gas turbine blades.

**Figure 28.** Impact of blade rotation on passage heat transfer [6, 19, 35].

The trailing edge regions of rotating blades are in general the most difficult to thermally design, largely due to the thin airofoil geometry, complex internal flow geometry, and the coolant flow conditions. The trailing edge region generally consists of coolant passages with very high aspect ratios, typically between 4 and 7, and which have cooling features such as turbulator and pedestals. Previous studies in such triangular and wedge-shaped passages with various turbulator shapes have been investigated by Refs. [36, 37, 38, 39]. They reported significant variation in the heat transfer distribution in both stationary and rotating cases. In a recent study [40], it has been shown that the impact of high Reynolds number typically found in heavy duty gas turbines can have a significant effect on the overall thermal performances of angled, broken, and chevron turbulators in a very large aspect ratio passage. **Figure 29**

**Figure 29.** Impact of turbulator design in trailing edge passage [40].

are for low Reynolds number, and with idealized geometry, care needs to be taken when implementing such results in real turbine blades with cast geometries, where the dimensions, shape,

The impact of rotation on the heat transfer from gas turbine blades can be significant and is dependent on several additional parameters such as the rotational and buoyancy numbers [34, 35]. **Figure 28** shows a schematic overview of the impact of rotation on the flow field in a twopass rotating passage of a gas turbine blade. Under rotating conditions, Coriolis and buoyancy effects can significantly alter the temperature and velocity profiles within the passages.

In a study from Ref. [35], **Figure 28** shows that with the coolant flowing radially upwards, the heat transfer with increasing rotation numbers, increases on the pressure side and reduces on the suction surfaces. Similarly, when the coolant flows radially inwards, the heat transfer increases on the suction side and reduces on the pressure side, especially for smooth passages. However, for passages with turbulators, the turbulators tend to dampen the effect of rotation and the heat transfer enhancement on the pressure and suction sides. This overall trend of rotation with different type of passages and turbulators has been observed by several studies

[33–35], and these effects play an important role in the design of gas turbine blades.

and position of the turbulators can be different from the predicted idealized geometries.

**Figure 27.** Passage heat transfer and frictional losses due to turbulator design and passage shape.

132 Heat Exchangers– Design, Experiment and Simulation

shows the complex flow structures and the high three-dimensional heat transfer distribution which exist within the high aspect ratio triangular passages with different turbulator shapes. **Figure 30** shows the comparison of the average thermal performances and shows that at high Reynolds numbers, the differences between the various designs are very similar. When considering the investigated turbulator design for gas turbine cooling applications, all three configurations show comparable levels of heat transfer performances.

For leading edge passages of gas turbine blades, the application of turbulators is also widespread. However, due to the leading edge geometry, the heat transfer is significantly different to that in midchord or trailing edge passages. Several studies show the impact of the turbulator geometry on the overall heat transfer in gas turbine blade leading edges [33, 41–43]. In a recent study by Saxer-Felici et al. [44], several turbulator geometries were tested at engine representative Reynolds numbers. **Figure 31** shows that the flow structure is significantly modified due to the presence of the turbulators, and this dominates the strength and distribution of the local and average heat transfer coefficients.

The final selection and implementation of turbulator designs in a turbine blade are dependent on several additional complex requirements. These include blade metal temperature, metal temperature gradients, cooling flow pressure margins, and the amount of required cooling flow. A further key criterion is for the fulfilment of the blade mechanical integrity, which is determined by the blade low cycle fatigue and creep behaviour, both of which are driven by the local metal temperature gradients and the absolute metal temperatures. An optimal balance of these factors is therefore needed to select the best turbulator concept for a blade design system.

**Figure 30.** Average heat transfer and frictional loss in trailing edge passages [40].

**Figure 31.** Heat transfer in leading edge passages [44].

#### **5.3. Convective cooling with pins and pedestals**

shows the complex flow structures and the high three-dimensional heat transfer distribution which exist within the high aspect ratio triangular passages with different turbulator shapes. **Figure 30** shows the comparison of the average thermal performances and shows that at high Reynolds numbers, the differences between the various designs are very similar. When considering the investigated turbulator design for gas turbine cooling applications, all three con-

For leading edge passages of gas turbine blades, the application of turbulators is also widespread. However, due to the leading edge geometry, the heat transfer is significantly different to that in midchord or trailing edge passages. Several studies show the impact of the turbulator geometry on the overall heat transfer in gas turbine blade leading edges [33, 41–43]. In a recent study by Saxer-Felici et al. [44], several turbulator geometries were tested at engine representative Reynolds numbers. **Figure 31** shows that the flow structure is significantly modified due to the presence of the turbulators, and this dominates the strength and distribu-

The final selection and implementation of turbulator designs in a turbine blade are dependent on several additional complex requirements. These include blade metal temperature, metal temperature gradients, cooling flow pressure margins, and the amount of required cooling flow. A further key criterion is for the fulfilment of the blade mechanical integrity, which is determined by the blade low cycle fatigue and creep behaviour, both of which are driven by the local metal temperature gradients and the absolute metal temperatures. An optimal balance of these factors is therefore needed to select the best turbulator concept for a blade design system.

figurations show comparable levels of heat transfer performances.

tion of the local and average heat transfer coefficients.

134 Heat Exchangers– Design, Experiment and Simulation

**Figure 30.** Average heat transfer and frictional loss in trailing edge passages [40].

**Figure 31.** Heat transfer in leading edge passages [44].

The use of pins and pedestals for enhancing the internal heat transfer in gas turbine blades and vanes is quite common particularly at the airofoil trailing edges, which generally demands aerodynamically small wedge angles and thin trailing edge diameters. Pin banks and pedestals are sometimes the only method of cooling in the space constrained narrow, converging trailing edges. They are also preferred from a manufacturing point as they tend to offer structural stability for casting. An additional advantage is that the pin banks also provide good mechanical integrity of the blades due to the robust structural support between the pressure and suction surfaces of the airofoil. In general, the pins are cylindrical in shape, tend to be thick relative to their height, has fillets imposed at the interface with blade walls, and are typically arranged in a staggered pattern. Although they provide high heat transfer due to a combination of high heat transfer coefficients from the base wall and the pins, they also entail high pressure losses. The latter disadvantage is generally not a major issue for convectively cooled blades, where there is availability of higher coolant pressure ratios relative to the surrounding hot gas pressure at the blade trailing edge.

There have been a large number of heat transfer studies on pin banks which have addressed the influences of pin geometry, channel shape, arrangements, pin shapes and combination of pins with dimples and turbulators [45–49]. For straight passages with pin banks, the local heat transfer generally increases from the first row of pins until the second to third row and then starts to decrease. For converging ducts, **Figure 32** shows that with a converging duct, the heat transfer increases in the downstream section. Additionally, by using a thicker pin in the rear portion of the passage, further increase in heat transfer can be achieved, which is driven by the increased Reynolds number.

In a recent study [50], several pin fin configurations were investigated in a trailing edge converging channel which consisted of cylindrical pins, conical pins, and a hybrid cyl-

**Figure 32.** Heat transfer in pin banks and pedestals [45, 46].

inder pin/turbulator configuration. **Figure 33** highlights the investigated geometries. **Figure 34** shows from both predictions and measurements that the flow is highly turbulent downstream of the pins and that the complex heat transfer distribution exists on both the endwall and the pins. High levels of local heat transfer occur at the leading edge stagnation point of the pins and at the leading edge endwall. Lower heat transfer coefficients were predicted and measured in the wake region of the pin trailing edge. Laterally, averaged local distributions of the heat transfer enhancement are shown in **Figure 34** for the tested geometries, and the nonuniform nature of the heat transfer in the pin banks are further highlighted.

**Figure 33.** Trailing edge passages with different pin bank configurations [45, 46].

**Figure 34.** Heat transfer in trailing edge passages with different pin bank configurations [50].
