**2. Micro-GT design specifications**

In a nutshell, the microturbines currently on the market have the design features described in the following references [1, 2]. They work according to a Brayton open cycle, which very often takes advantage of exhaust gases heat recuperation for air compressor discharge. The structure of these single-shaft microturbines includes an annular, or silo combustor, single-stage radial flow compressor as well as expander, and an optional recuperator. The electric power ranges between 30 and 330 kW. They work with low pressure ratios (typically in the range 3–5), high nominal rotation speeds (ranging between 43,000 and 116,000 rpm), and relatively low efficiency (17–20% for simple cycle machines and around 30% for recuperated machines). The design life is usually between 60,000 and 80,000 hours with overhaul [3]. Turbine entry temperatures range between 700 and 1000°C, while exhaust temperatures are in the range of 260–310°C (for recuperated machines).

Accordingly, **Table 1** reports the main data of a micro-GT unit designed in a previous work [4], which can be assumed as the typical machine and reference for subsequent calculations.


**Table 1.** Main data for the design of the reference micro-GT unit.

### **2.1. Bearing external loads**

generation as well as cogeneration, and they are involved in the air compression as well as conditioning market. Recently, manufacturers are addressing their efforts to new market areas, for example, in powering hybrid electric vehicles as well as autonomous robots in the

In the following, such microturbines are intended as autonomous power generators. Despite their name, excluding portable devices and MEMS, typical shaft diameters of commercial

In the design of a microturbine, the bearing choice is not a trivial issue due to the high rotation speed and working temperature. In these operating conditions, the "classical" engine design criteria for choosing the most adequate bearing type (rolling or sliding bearings), which are based on dimensions and/or nominal power rate, cannot be adopted. According to such traditional design criteria, the choice for machines of small and large dimensions is unavoidable, i.e., rolling and sliding bearings are employed, respectively. For intermediate power machines, on the contrary, specific choices have to be carried out: on the one hand, rolling bearings have smaller overall dimensions and purchasing costs, and on the other hand, slid-

Therefore, for microturbines, a comparative analysis of the available support systems, particularly focused on studying the influence of operating conditions (speed, temperature, and loads), is required. Accordingly, the present chapter provides a detailed comparison of bearing technologies by means of literature review and analysis of published data. Cutting-edge (e.g., magnetic, air, and ceramic bearings) and well-established solutions (e.g., steel rolling element and oil-lubricated slide bearings) are both considered in order to provide a large perspective. On the basis of such a comparative study, an innovative design of a support system for micro-GTs capable of overcoming the limits of modern units is proposed. It employs different types of bearings and requires a proper design of their arrangement and coupling in order to take the maximum advantage of the peculiarity of each bearing. After a brief description of the existing bearing arrangements, the present chapter deals with the conceptual design of the

In synthesis, the present chapter defines the bearing requirements on the basis of design specifications of a typical micro-GT unit. Afterward, it studies the suitability of different bearing types for the operating condition requirements by means of a literature review. Finally, the innovative support system is proposed and compared with the current technical solutions.

In a nutshell, the microturbines currently on the market have the design features described in the following references [1, 2]. They work according to a Brayton open cycle, which very often takes advantage of exhaust gases heat recuperation for air compressor discharge. The structure of these single-shaft microturbines includes an annular, or silo combustor, single-stage radial flow compressor as well as expander, and an optional recuperator. The electric power

microturbines are roughly 10 mm, and their electric power is in the order of 100 kW.

case of small-size machines.

2 Bearing Technology

ing bearings are more reliable.

new support system.

**2. Micro-GT design specifications**

**Table 2** gathers the bearing loads computed for the reference unit. The total radial load is due to the shaft weight, and it is directed downward. The positive direction of the axial forces listed in **Table 2** is from compressor to turbine, as also shown by the direction of *z*-axis in **Figure 1**.

The total thrust load is computed as the algebraic sum of compressor and turbine axial forces. Both of the forces are the resultant of the axial thrusts exerted on the blades and on the backside of the impellers by the working fluid.

Consequently, together with the area of the impeller backside, the pressure on the clearances between casing cover and impeller back shroud (backside pressure) plays an important role in determining the nominal axial thrusts, which are plotted in **Figure 1** as a function of the backside pressure. By means of CFD simulations, a backside pressure equal to 0.25 MPa (Case A) has been calculated so that both the impeller thrusts are directed toward the external side of the unit. Nevertheless, for different impeller geometries, which may yield different pressure drops between compressor delivery and clearances, compressor, turbine, and total thrust directions can reverse, as shown by **Figure 1**, for backside pressures lower than 0.15 MPa, and as a consequence, such cases (both thrusts directed toward the inner side of the machine) must be taken into account. To this purpose, for the same reference value of the total axial load (500 N) Case B is identified in **Figure 1**. The axial loads in both Case A and B are summarized in **Table 2**. An axial force reversal can also occur during transient operation, e.g., the start/stop phase of the unit.


**Table 2.** Loads acting on the bearings of the reference micro-GT unit.

**Figure 1.** Compressor, turbine, and total thrust acting on the shaft in nominal conditions for the reference micro-GT as a function of the backside pressure.

#### **2.2. Bearing requirements**

quence, such cases (both thrusts directed toward the inner side of the machine) must be taken into account. To this purpose, for the same reference value of the total axial load (500 N) Case B is identified in **Figure 1**. The axial loads in both Case A and B are summarized in **Table 2**. An axial force reversal can also occur during transient operation, e.g., the start/stop phase of the unit.

(N) −600

(N) 1200

(N) −1700

(N) 1100

**Design variable (unit) Value** Radial load (rotor weight) W (N) 40 Total thrust load (absolute value) Tref (N) 500 Axial thrust on compressor blades (N) 2100 Axial thrust on turbine blades (N) −3010 Case A Pressure on the back side of the impellers (gauge) (Pa) 250,000

Case B Pressure on the back side of the impellers (gauge) (Pa) 80,000

Compressor thrust Tc

Compressor thrust Tc

**Table 2.** Loads acting on the bearings of the reference micro-GT unit.

Turbine thrust Tt

4 Bearing Technology

Turbine thrust Tt

**Figure 1.** Compressor, turbine, and total thrust acting on the shaft in nominal conditions for the reference micro-GT as

a function of the backside pressure.

Requirements are related to the operating conditions of the units (operative speeds and temperatures) and bearing performance (duration, load-carrying capacity, and power loss).

The high rotation speeds of micro-GT shafts yield different problems (e.g., ball spinning, centrifugal loading, and skidding), which have greater effect on larger bearings. Therefore, an important requirement is that bearings have a high peripheral speed limit.

In the following, the effect of peripheral velocity is quantified in terms of the DN speed factor (where D is the diameter of the shaft housing expressed in mm, and N is the rotation speed expressed in rpm).

Micro-GT units must usually operate in high-speed conditions, which for rotors are usually characterized by speed factors ranging between 1 × 10<sup>6</sup> and 2 × 10<sup>6</sup> mm rpm. In ultra-highspeed rotor applications DN overcomes 2 × 10<sup>6</sup> mm rpm, as in the case of mobile power systems [5], which are beyond the scope of the present paper.

Indeed, in the case of the reference microturbine, the operative speed factor is roughly 1 × 10<sup>6</sup> mm rpm (see **Table 1**) so that bearings with DN limit lower than such operative value must be discarded.

In addition, the bearings must operate throughout the domain of possible temperature conditions of the microturbine. Such temperatures range between 100 and 1000°C in nominal operation, while during start-up, they extend down to room temperature. Expected working temperature of the bearings depends on their location and relevant thermal management, e.g., type of cooling or vent.

Bearing duration should preferably overcome 70,000 hours (the average duration of the units) and should not be limited by the number of start/stop cycles so that maintenance and bearing replacements could be minimized.

Reasonably, predicted axial load is a magnitude order higher than radial load (**Table 1**). The former load is significant as far as metal fatigue and wear are concerned, while the latter is so light that it may cause stability problem at high speed in self-acting radial bearings.

Since the state-of-the-art efficiency of microturbines is quite low, bearings must not reduce further this value. To this goal, they should be designed for the minimum power loss (e.g., friction) and their possible power input should not be significant.

Finally, depending on the shaft layout, resonance eigenfrequencies usually occur at a speed lower than the operating one; therefore, the vibration amplitude should be limited by the bearing damping.
